US2484554A - Centrifugal impeller - Google Patents

Description

Oct. 11, 1949.. c. CONCORDIA ET AL CENTRIFUGAL IMPELLER 2 Sheets-Sheet 1 Filed Dec. 20, 1945 Inventor-s: Gh ar-les Concordia,

Millard F. Dowell. 10 5 mm Their Attorngy.

Oct. 11, 1949.

Filed Dec. 20, 1 945 Axis of impeller -|6 Relative Ar'ea Blade loading-Ap c. CONCORDIA ET AL 2,484,554

CENTRIFUGAL IMPEL-LE R 2 Sheets-Sheet 2 Fig.9. I

forward Inventors:

Ch ar'les Concordia,

Millar-d' F Dowel Their Attorngs grammatically :sents a modification of the impeller in Patented Oct. 11, 1949 2,484,554 CENTRTFUGAL IMPELLER Charles Concordia, Schenectady, N. Y., and Millard F. Dowell,

New York Lynnfield, General Electric Company,

Mass, assignors to a corporation of Application December 20, 1945, Serial No. 636,133 6 Claims. (Cl. 230134) This invention relates to impellers for turbomachines, particularly to a centrifugal impeller for imparting a pressure rise to a compressible fluid.

The object of the invention is to provide a turbo-machine impeller which combines the characteristics of several known types of impellers in a manner to produce improved'efficiency and other performance characteristics.

Another object is to provide a turbo-machine impeller, having blades of a novel shape specifically designed to produce a preselected desired loading of the blades.

Other objects and advantages will appear from the following description and the appended claims, taken in connection with the accompanying drawings, in which Fig. 1 is a perspective view of an impeller having blades designed in accordance with our invention; Figs. 2, 3 and 4 represent top, front and side views of a single blade; Figs. 5, 6, 7 and 8 are diagrammatic illustrations of previously known centrifugal impellers; Fig. 9 is a graphic representation of certain characteristics of an impeller embodying the invention; and Figs. 10 and 11 represent possible modifications of the basic impeller form shown in Fig. 1.

In order to explain clearly the shape and characteristics of our new impeller, reference will be made to various types of centrifugal impellers known to the prior art. Fig. 5 represents diagrammatically the most usual form of centrifugal impeller, having fiat blades each composed of straight-line elements parallel to, and lying in a radial plane through, the axis of rotation. (For simplicity, the impellers represented diain Figs. 5, 6, '7 and 8 have been shown with only a few blades). Fig. 6 repre- Fig. 5, known as a backward sloped impeller. The blades of this type are characterized by the fact that while they are flat and extend in a substantially radial direction, the plane of the blade does not pass through the axis of rotation but is inclined backwardly with respect to the direction of rotation, which direction is indicated by the arrow l8. The type of blade inclination represented in Fig. 6 will be referred to herein as "backward slope. Fig. '7 represents another known type of centrifugal impeller similar to those of Figs. 5 and 6, except that the blades are forward sloped." Here each blade is flat, lies in a plane perpendicular to the plane of rotation (that is, parallel to the axis of rotation) but not passing through said axis, being inclined forwardly with respect to the direction of rotation indicated by the arrow l8. Fig. 8 represents an impeller which is well known as a "mixed fiow impeller. This type is characterized by the fact that the blades are substantially fiat but the plane of the blade is tipped" forwardly in the direction of rotation so that the plane of the blade forms an acute angle with the plane of rotation. The mixed flow impeller is further characterized by the fact that it has an appreciable axial depth so that the tipped blades act partly like an axial flow compressor and partly like a centrifugal compressor. It is because of this combination of axial and radial flow that this type has been called a mixed flow impeller. To recapitulate, the term slope" will be used herein to denominate the type of inclination shown in Figs. 6 and "I, while the term tip will be used to mean blade inclination as in the case of the mixed flow impeller of Fig. 8.

We have discovered by analysis, backed by long experience with known forms of centrifugal impeller, that optimum efliciency and other performance characteristics can be obtained from a turbo-machine impeller if due consideration is given to the blade loading. Blade loading may be defined as the differential static pressure existing across an infinitesimal increment of area of the blade at a given location along the fluid flow path through the impeller. The definition and significance of this factor may be explained by reference to the drawings, in which Fig. 1 shows an impeller comprising a blade support body 8 having a number of equally spaced substantially radially arranged blades 3. The mean flow path through the fluid passages defined by the blades of the impeller shown in Fig. 1 may be represented by the dot-dash line I, which may be considered to be, the locus of the centers of gravity of cross sections of the flow passage, taken in a direction perpendicular to the mean fiow path at the given section. At the point I) on the mean flow path i in Fig. 1, the blade loading would be represented by the difference in the static pressure exerted by the flowing fluid on an infinitesimal increment of blade area represented, to an enlarged scale, by the rectangle 2.

By reference to well-known aerodynamic prinv ciples governing fluid flow in such an impeller, it can be shown that for a blade of a given preselected geometric shape, the blade loading can be determined from the diiferentialequation in which p is the density in mass units 1', distance of a given axis 01 rotation in particle of fluid from the w, average angular velocity of a particle of fluid I about the axis of rotation at radius r or at time t du/dt, time rate of change of angular velocity u, that is, angular acceleration of a fluid particle about the axis of rotation V7, radial component of the absolute velocity in space V of a fluid particle at radius r from the axis of rotation Ap, blade loading A0, angular-distance between correspondingmean flow paths of adjacent passages (equal to 21f radians divided by the number of blades).

Then, noting that density p increases as the static pressure of the fluid increases and that .angular velocity u of a fluid particle at a given point in the flow passage depends both on the rotational speed of the impeller and on the direction of the flow path at the given point, this expression can be solved by various known methods (as for instance by graphical integration) to obtain the blade loading Ap at the given point.

Then the increment of pressure rise produced by the impeller blades between two points on a given flow path, for instance the points a and b on path I in Fig. 1, can be represented by where in which a. is the value of a: at point a dr, differential of the radius r Ap, blade loading from Equation (1) dp, acute angle which a plane tangent to the blade at a given point makes with a plane defined by the rotation Vz+r, component of the absolute velocity in space of a fluid particle at a given point, lying in a plane through the axis of rotation and having a 2 component in an axial direction and an r-component in a. radial direction (measured parallel with and perpendicular to the axis of rotation respectively) dV=+r, diiferential of the velocity component The above expressions to suggest one possible will be recognized and method of analysis and readily understood by those skilled in fluid mechanics. Further details 1 of the method of calculating the design of the blade are not necessary to the disclosure of our invention.

(in accordance with an increase in distance from given point and the axis of are given here merely the third term being negligible. With jection views of one are twenty-two identicalcurved blades. spaced circumi'erentially' and having portion 4,

the axis of rotation) with a resulting increase in the centrifugal force exerted on the fluid. Likewise, the second term represents the increment in pressure (which increment may be positive or negative) due to the angle of slope" of the blades measured to a radial plane through the axis of revolution. This second term will be positive for a "forward sloped impeller as in Fig. 7, and negative for a "backward sloped impeller"'as .in Fig. 6. The third term represents the eiiect of the variation in flow path area, disregarding all tangential components of fluidmotion and considering only the difiusing eliect of the flow passage, that is, the conversion of kinetic energy third term also becomes important. It maybe noted in passing that for a pure axial flow compressor (not shown in the drawings) the above expression for the blade loading is substantially the second term plus the third term, with the first term. having only an incidental effect.

A consideration layer characteristics of centrifugal impellers led us to the conclusion that best efliciency could .be-

obtained if the blade loading, as defined above,

could be made substantially zero at the inlet tothe impeller and caused to increase smoothly but rapidly to a maximum value which is maintained over-a major portion of the flow path and then caused to drop smoothly but rapidly to substantially zero at the exit of the im peller flow passages. We have found that a composite impeller combining in a particular way the characteristics of the'mixed flow impeller of Fig. 8, the forward sloped impeller of Fig. 7, and the backward sloped impeller of Fig. 6, produces such a blade loading and gives superior performance. The manner in which these three known types of impeller blade shapes are combined will be readily apparent from the perspective view of a complete impeller shown in Fig. 1, taken in connection with the orthographic problade shown in Figs. 2, 3, and .4.-

' Fig. 1 illustrates a complete impeller, consisting diameter at one axial end and a minor diameter at the other axial end Ythe axial length of the body being approximatelyof the same magnitude as the major radius. This-body has a concave outer surface (appr ximately a hyperboloidal surface) on which the blades aresecured and is provided with an axial bore 9 for mounting the impeller on a shaft. Supported on the body 8 3, equally an inlet an intermediate portion 5, and a tip or discharge portion 6. The extent of each of these threev distinct blade portions is clearly indicated in Figs. 1-4. Inlet part 4 extends from the inlet blade impeller as in Fig. 5,

(but usually negative), while the of the fundamental. boundarya solid body of revolution 8 having a major.

edge I to a point approximately one-third the length of the mean flow path I from the inlet edge I. This corresponds to the region of rapid change of curvature indicated by the shade lines near the middle of the blade in Fig. 3. Likewise the intermediate portion merges into the tip portion 6 at the location indicated by the shade lines adjacent the :blade tip in Fig. 3.

From a consideration of Figs. 1-4, it will be readily apparent that the inlet portions I are in the form of mixed flow impeller blades, similar to those represented in Fig. 8; the intermediate blade portions 5 are in the form of forwardly sloped impeller blades, as shown in Fig. '7; while the tip portions 6 are in the shape of backwardly sloped impeller blades, as in Fig. 6. These blade portions must of course be modified slightly at the place where they join an adjacent portion so that the mean flow path through the fluid passages formed between adjacent blades will be a smooth curve.

The inlet portion 4 is designed in accordance with the well-known principles governing the design of the mixed flow impeller so that fluid approaching the inlet edge 'I of the blades, in a substantially axial direction, is smoothly accelerated by the blade in a tangential and axial direction, and perhaps also given some very slight acceleration in a radial direction. The intermediate blade portion 5 defines roughly 50 per cent of the length of the mean flow path and furnishes the major portion of the work energy imparted by the impeller to the fluid. The comparatively short tip portion 6 serves to gradually relieve the loading, to a value approaching zero at the exit edge It.

For convenience, our new blade shape may be referred to as an S-blade", because of its characteristic reverse curvature.

The upper portion of Fig. 9 represents graphically the blade loading of our S-blade impeller as a function of the radial distance from the axis of rotation. While Figs. 2, 3 and 4 are to substantially the same scale as Fig. 1, the abscissa of Fig. 9 is to approximately double the scale of Figs. 1-4. The heavy curve shown represents the loading along the mean flow path I. It will be understood by those skilled in the art that similar loading curves can be calculated for flow paths occupying other positions in the fluid passage. for instance one lying along the forward portion of the fluid passage (relative to the direction of rotation) which maybe represented by the line I I in Fig. 1. Likewise the blade loading can be calculated for a flow path lying along the r arward portion of the fluid passage, as for instance that represented by the line I! in Fig.

4 1. Similarly, load curves can be calcualted for flow paths lying adjacent the curved outer surface of the body 8, or adjacent the curved inner surface of the shroud which forms the fourth wall of the fluid flow passages.

No shroud has been illustrated for the impeller of Fig. l; but it will be readily understood by those skilled in the art that the impeller shown is what is known as an open impeller which, when assembled in its casing. is closely surrounded. by a stationary curved wall forming a close clearance with the free edges l3 of the blades. An open impeller assembled with its casing and other component parts is shown in U. S. Patent 2,377,740, filed March 31, 1944 in the name of J. S. Alford. It will be obvious that our invention may be incorporated also in a shrouded or "closed" type impeller, in which the fourth wall 'iiuid flow path, deflned by the inlet edges I of the blades, occurs at a radius roughly 25 per cent of the tip radius. The termination of the mixed flow or inlet portion 4 of the blade corresponds roughly to point It in Fig. 9, at a radius of about 40 per cent of the tip radius. The loading starts at zero, or a very low initial value, at inlet edge I; and by the time the fluid reaches point ll, the blade loading has increased to very nearly its maximum value, which is maintained at a high, substantially uniform, value throughout the intermediate. blade portion 5. The intermediate portion terminates approximately at point l5; and from that point to the blade tip II), the loading decreases smoothly but rapidly to substantially zero, under the influence of the backward sloped tip portion 6.

For purposes of comparison, curves are shown in Fig. 9 representing typical blade loadings for the known types of impeller represented in Figs. 5-8. The forward sloped impeller gives the highest blade loading; the backward sloped impeller gives much lower loadings; while the plain radial blade produces a loading curve lying between those for the forward sloped and backward sloped impellers respectively. The mixed flow impeller has a distinctive blade loading curve quite different from the other known types represented. .In producing our new blade shape we have utilized the characteristics of the forward sloped blade to keep the blade loading uniformly high over the intermediate portion of the blade, which does the major portion of the work. The characteristics of the mixed flow impeller are employed to increase the loading from a low initial value to a value in the neighborhood of that produced by the intermediate portion. By merging the forward sloped intermediate portion smoothly with the backward sloped tip portion, the loading curve is caused to drop smoothly to a very low value, or to zero.

We have discovered by analysis and experiment that optimum results are obtained when the mixed flow portion 4 of our S-blade defines from 20 to 40 per cent of the mean fluid flow path, while the backward sloped tip portion 6 defines from 10 to 30 per cent of the mean flow path. It has also been found that best results are obtainedwhen the length component of the mean flow path in a direction parallel to the axis of rotation is from 40 to 60 per cent of the outside diameter of the impeller, that is, when the axial depth of the impeller is roughly equal to the tip radius of the blades. For convenience, this axial length component of the fluid flow passages may be referred to as the Z-dimension. This is the same e which appears in Equation (2) above.

It is well known to those skilled in this art that energy losses due to turbulence are likely to occur when the shape of the walls defining the fluid flow passage is such as to cause a change in direction of the fluid. This tendency becomes more pronounced as the rate of curvature increases, by reason of separation of the boundary layer from the passage wall. It has been found that such turbulence losses can be reduced by causing the fluid flow passage to converge slighty concurrent with the change in direct on. In accordance with this principle, the flow passages defined by our S-shaped blades decrease in crosssectional area as represented by the curve shown in the lower part of Fig. 9. It will be seen that the cross-sectional area of the passage progressively decreases throughout the mixed flow portion 4, in which part of the flow path the maximum change in direction of the fluid flow occurs. The cross-sectional area of the intermediate and tip portions of the flow passage may remain substantially constant, as represented by the curve in Fig. 9, or may continue to decrease slightly, or may in some impellers increase again. We have found that best results are obtained if the passage area decreases to a value between 70 and90 per cent of the area at the inlet to the impeller by the time the heavily loaded intermediate portion 5 of the blade is reached. The relative area curve shown in Fig. 9 reaches a minimum of approximately 80 per cent of the inlet area at a location slightly beyond the point M on the blade loading curve in Fig. 9.

The relative cross-sectional area of the fluid flow passages through our impeller may be varied either l) by reducing the height of the passage (measured in a direction normal to the surface of i the blade support body 8), or (2) by thickening the blades so'as to reduce the width of the passage (measured from one blade to the next adjacent blade in a direction perpendicular to the mean flow path). Either or both of these methods, may be used in making this relative cross-sectional area of the flow path vary in the desired manner. Consideration of Fig. 1 will reveal that the blades are thickest in the region where the mixed flow blade portion 4 merges into the forward sloped portion 5. This thickening of the blades is an important factor intended to reduce the cross-sectional area in accordance with the relative area curve in Fig. 9.

Figs. 1-4 illustrate our basic blade shape with the blades arranged relative to each other and to the support body 8 so that the arbitrary "center-line" i1 shown in Fig. 2 passes through the axis it of the impeller. Figs. 10 and 11 show possible modifications, in which the center-line l'l is "sloped relative to the direction of rotation (indicated by' the arrow It). In Fig. 1 blade 3 is "backwardly' sloped by the amount of angle l9; that is, the center-line l'lforms the angle is witha' radial plane. In Fig. 11 the arbitrary center-line l l'is forwardly sloped, making the angle 20 with a radial plane.

Sloping the blade backwardly'as'T in Fig. 10 has the effect of decreasing the average blade loading; while sloping the blade'iorwardly as in Fig. 11 has the efiectof increasing the blade loading. If our S-shaped blade is sloped backwardly far enough, in the manner of Fig. 10 but to a greater degree, the intermediate portion of the blade may become substantially radial. Likewise, if the blade is sloped forwardly a suficient amount, in the manner of Fig. 11, the discharge portion 6 of the blade may closely approach the radial direction. Many such modifications have been studied by us, and all have their place in the design of various impellers of special characteristics.

While our improved turbo-machine blade shape has been specifically described as applied to a compressor impeller, it will be, apparent to those skilled in the art that the invention is also applicable to turbine rotors, in -which case nozzles would be arranged to direct-motive fluid radially, inward to the blade tip portions- 6. Thus the fluid would flow through the impeller in the reverse direction, as compared with the flow in a compressor impeller; and the pressure gradients, blade loadings, etc. would be similar qualitatively to those obtaining in a compressor impeller incorporating our invention. a

What we claim as new and desire to secure by Letters Patent of the United States, is:

1. In a compressor impeller arranged to receive fluid with a substantially axial velocity adjacent the axis of rotation and to discharge it at a ma terially greater radius from the axis of rotation, walls defining a passage having a mean flow path consisting of a first portion forming 20 to 40 per cent of the length of the path nearest the axis of rotation and tipped forwardly in the direction of rotation to form an acute angle with the plane of rotation, a second portion forming 10 to 30 per cent of the length of the flow path most remote from the axis of rotation and sloped backwardly relative to the axis, and a portion intermediate the first and second portions forming the remainder of the flow path and sloped forwardly relative to the axis, said passage having a cross-sectional area which progressively decreases from the inlet of the first portion to a value of from 70 to per cent of the inlet area at the juncture of the first portion with the inter mediate portion.

2. In a centrifugal impeller arranged to receive fluid with a substantially axial velocity adjacent the axis of rotation and to discharge it at a materially greater radius from the axis of rotation, walls defining a fluid flow passage, said walls including a plurality of curved blades circumferentially spaced from each other, each blade having an inlet portion in the form of a mixed flow impeller blade defining 20 to 40 per cent of the length of the fiow path from the inlet thereto, a

discharge portion in the form of a backward sloped impeller blade defining 10 to 30 percent of the length of the flow path adjacent the exit thereof, and an intermediate portion in the form of a forward sloped impeller blade defining the remainder of the flow path connecting the inlet portion and the discharge portion, the width and thickness of said blades being dimensioned so that the cross-sectional area of the fluid passagedefined by adjacent blades progressively decreases from the inlet of the mixed flow portion to a value of from 70 to 90 per cent of the inlet area at the juncture of the mixed flow and forward sloped portions.

3. A turbo-machine impeller comprising a blade support body having a substantially hyperboloidal outer surface and a plurality of circumferentially spaced radially extending blades secured to said surface, each adjacent pair of blades defining a fiow path having a first radially inner terminal portion constituting 20 to 40 per cent of the total length of the flow path and having a mean fiow path substantially straight with a small radial component and substantial tangential and axial components, a second radially outer terminal portion forming 10 to 30 per cent of the total length of the flow path and having a mean flow path substantially straight with a negligible axial component and a substantial radial component, and an intermediate portion connecting the first and second portions and having a mean flow path substantially straight with substantial axial and radial components, the intermediate portion being modified at its respective ends to cooperate with the first and second portions so as to form a smooth reversely curved flow path having a, blade loading which is. maintained at a substantially uniform high value throughout the intermediate portion of the flow path and falls off smoothly throughout the first and second terminal portions to minimum values approaching zero at the impeller inlet and exit respectively.

4. A turbo-machine impeller comprising a blade support body having a substantially hyperboloidal outer surface and a plurality of circumferentially spaced radially extending blades secured to said surface, each adjacent pair of blades defining a flow path having a first radially inner terminal portion constituting 20 to 40 per cent of the total length of the flow path and having a mean flow path substantially straight with a small radial component and substantial tangential and axial components, a second radially outer terminal portion forming to 30 per cent of the total length of the flow path and having a mean flow path substantially straight with a negligible axial component and a substantial radial component, and an intermediate portion connecting the first and second portions and having a mean flow path substantially straight with substantial axial and radial components, the blades being dimensioned so that the cross section area of the fiow path defined therebetween decreases progressively throughout the first portion to a value at the juncture with the intermediate portion of from 70 to 90 per cent of the area at the other end of the first portion, the intermediate portion being modified at its respective ends to cooperate with the first and second portions so as to form a smooth reversely curved flow path having a blade loading which is maintained at a substantially uniform high value throughout the intermediate portion of the flow path and falls off smoothly throughout the first and second terminal portions to minimum values approaching zero at the impeller inlet and exit respectively.

5. A turbo-machine impeller comprising a blade support body having a substantially hyperboloidal outer surface and a plurality of circumferentially spaced radially extending blades secured to said surface, each adjacent pair of blades defining a flow path having a first radially inner terminal portion constituting 20 to 40 per cent of the total length of the flow path and having a mean fiow path substantially straight with a small radial component and substantial tangential and axial 60 components, a second radially outer terminal portion forming 10 to 30 per cent of the total length I of the flow path and having a mean fiow path substantially straight with a negligible axial component and a substantial radial component, and an intermediate portion connecting the first and second portions and having a mean flow path substantially straight with substantial axial and radial components, the intermediate portion being modified at its respective ends to cooperate with the first and second portions so as to form a smooth reversely curved flow path having a blade loading which is maintained at a substantially uniform high value throughout the,intermediate portion of the flow path and falls off smoothly throughout the first and second terminal portions to minimum values approaching zero at the impeller inlet and exit respectively, the axial length of the complete flow path being from 40 to per cent of the outer tip diameter of the path.

6. In a centrifugal impeller adapted to receive fluid with a substantially axial velocity adjacent the axis of rotation and to discharge it at a materially greater radius from the axis of rotation, walls defining a fluid flow passage having an inlet portion the mean flow path of which has a substantial tangential component in the direction of rotation, said walls including a blade support body having a substantially hyperboloidal outer surface and a plurality of circumferentially spaced radially extending blades secured to said body, the width and thickness of said blades being so dimensioned that the cross-sectional area of the fiow passage decreases progressively throughout the inlet portion to a value at the end of said portion of the order of per cent of the inlet area, whereby turbulence and boundary layer separa tion are reduced.

CHARLES CONCORDIA. MILLARD F. DOWELL.

REFERENCES CITED The following references are of record in the file of this patent;

UNITED STATES PATENTS Number Name Date 673,244 Davidson Apr. 30, 1901 1,314,049 Criqui Aug. 26, 1919 1,959,703 Birmann May 22, 1984 2,399,852 Campbell et al. May 7, 1946

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