US2866619A - Rotary bladed fluid flow machines - Google Patents

Description

Dec. 30, 1958 A. D. s. CARTER 2,866,619

ROTARY BLADED FLUID FLOW MACHINES Filed March 5, 1952 2 Sheets-Sheet l a Inventor Dec. 30, 1958 A. D. s. CARTER ROTARY BLADED FLUID FLOW MACHINES 2 Sheets-Sheet 2 Filed March 5, 1952 i R 9k 2 Invenfas 2 United States Patent ROTARY BLADED FLUID FLOW MACHINES Alfred Denis Snowdon Carter, Farnb'orough, England, assignor to Power Jets (Research and Development) Limited, London, England, a Britishcompany Application March 5, 1952, Serial No. 274,950

Claims priority, application Great Britain'March 9, 1951 3 Claims. (Cl. 253-65) This invention relates to rotary bladed -fluid flow machines, more particularly to turbines, and has for an object the improvement of the efliciency'thereof by reducing energy losses associated with the passage of ffluid through the machine.

In general, the energy losses associated with a row of cambered blades such as are used in compressors and turbines are attributable to two main causes, namely frictional drag between the fluid and surfaces 'of "the blades and bounding walls, known as primary losses, and dissipation of kinetic energy of the fluid invarious'turbulent flows induced therein, known as secondary. losses, and it is with the reduction of the latter thatthe present invention is concerned.

The invention proposes broadly that the row or last row of blading in a machine of the kind under consideration intended primarily for energy conversion (i. e. the working blades) be immediately succeeded, in the direction of fluid flow, by a row of blades with respect to which this working blade'row is rotatable having loss characteristics which are low in relation to those of'the next preceding row of blades, i. e. as explained in'more detail hereinafter, blades which are at most of very low cambered section compared with the blades of the preceding row.

The invention arises from a consideration of the probable relative magnitudes of primary and secondary losses in various cases and the factors influencing their magnitude as inferred from knowledge at present available, this inferential assessment being obligatory in the absence of any technique for direct but separate measurement of the two forms of loss. I

An indication of the magnitude of these'losses .is afforded by the fact that a typical stage of compressor blading comprising an intermediate stage of a'multi-stage machine would have an efficiency, at best, of approximately 90 percent. Practical experiencerhas led to' the computation of primary and secondary losses as accounting for 6 percent and 4 percent respectively of the 10 percent total loss of efliciency, and in order to' afford a basis for the accurate estimation of secondary losses in such a typical compressor stage the secondary losses have been conveniently expressed as giving rise to a'co-e'fiicient of drag induced thereby, C and the following empirical formula propounded:

where C is the co-efficient of lift of the blades of'the compressor stage.

The turbulence giving rise to the secondary losses has been found in tests on cascades of blades under laboratory conditions to be principally in the'form of vortices extending downstream from the trailing edge-of the-blades, a'pairof oppositely rotating vortices being situated one toward each end of each blade adjacent the bounding walls. Analysis of the circumstances surrounding the formation of this vortex system has led t'o'anacc'eptable explanation and understanding of the phenomena-an'd ice to a theoretical basis for the-assessment of. the secondary losses, again expressedffor convenience as a co-eflicient of drag. induced thereby, .by identifying the phenomena with a vortex 'systemextending from the blade an infinite distance downstream. Thus the-followingexpression has been derived wheres/c is the pitch/chord ratio of the blades in .the cascade,.and h'/h is the ratio of the span of the axes of the pair of vortices extending from each blade to the span of the blades. Typical values for these ratios in a compressor stage would 'be of the order 0.88 and 0.75 respectively so thatin general by substitution in the. above equation;

Cm=0.0532 CLZ Thus both practical and theoretical assessments of 'the secondary loss induced drag co-eflicient are dependent on C the lift co-efiicient; this is to be expected since the turbulence of the vortex 'system'is,-in general, intensified with increasing blade camber, which also gives rise to increased lift co-efficients. In respect of actual magnitude, however, the practical and theoretical estimates do not agree, the latter being approximately 3 times the former, and there is consequently a natural tendency to doubt the theoretical estimate.

- On the other hand, tests have beencarried out, for the purpose of measuring the variation of total losses in a single stage turbine with varying blade tip clearances, from which some indirect support for-the validity of the theoretical estimate of secondary loss induced drag may be deduced. Since a variation in tip clearance'would seemingly not affect appreciably the frictional losses of blades and bounding walls, the variation in total losses would be expected to give a direct indication of the variation of secondary losses due to re-adjustment of the span of the vortex system upon variation of the tip clearance. On this basis, it is found-that the variation in'measured losses is, in fact, closely represented by the variationin C as derived from the general theoretical equation given above, which would appear in consequence to be valid, at least in the case ofa single stage turbine.

In explanation of the disparity of the-secondary losses estimated on a practical basis in a typical compressor stage and on a practical basis in a turbine stage, or on a theoretical basis in either case, it is concluded that these losses are least in the case of the 'typical' intermediate stage of a multi-sta'ge compressor because in that case only is the blade stage followed immediately by a further relatively moving blade stage. It is thought that the turbulence giving rise to secondary loss is curtailed by that further blade-row instead of extending some distance (in theory an infinite distance) downstream. From this conclusion one would expect the secondary losses in the final blade row of a multi-stage compressor to be higher than in a typical intermediate row; i. e. if in the latter' case the ratio of primary to secondary losses is 6z'4,-inthe former case the ratio would be 624 3 or 6:12, and consequently the ratio of total (primary plus secondary) losses of the final to any other row should be approximately 6+12/6-l-j4, i. e. 1.8. This figure is effectively confirmed by the only published measurements of losses through the various blade rows of'a compressor, where the correspondingratiois approximately 2. 4

-'It may, therefore, be assumed with confidence that while the primary losses of any given row" of blades are substantially unaffected-by external influences, the secondary losses are considerably reduced by "the'mere presenceof a succeeding relatively"mov ing blade row, {the latter actingas aloss inhibitor to the "said row -of blades't Accordingly the invention is based on the proposition, arising out of that assumption, that the total losses in the said blade row when not succeeded by another row will be greater than the combined total losses of the said row and a succeeding (loss inhibiting) row provided thatthe blades of the succeeding row have of themselves sufficiently low combined primary and secondary losses compared with the consequent reduction in secondary losses of the blades of the first mentioned row.

A quantitative illustration of this proposition is alforded by considering by way of example a particular working blade under certain operating conditions with or without an inhibitor blade and using the comparative loss relationships discussed in the foregoing. If it is supposed that the uninhibited secondary loss of the working blade would be X then the inhibited secondary loss of the working blade would be X/3, and if it is supposed that the uninhibited secondary loss of the inhibitor blade would be y then the primary loss of the inhibitor blade would be 6y/12 or y/Z. If, therefore, the total inhibitor blade losses are to be smaller than the difference between inhibited and uninhibited working blade secondary losses then;

It would appear, therefore, that a net reduction in loss is possible wherever a row of working blades, not being followed by a further row of working blades, can be followed immediately by a row of inhibitor blades whose losses are somewhat less than half as great in comparison with those of the working blade. These inhibitor blades can moreover be of simple aerodynamic form with little or no camber so that their losses are of a very low order, there being no other limitations affecting their design.

The invention offers the possibility of improving the efiiciency of turbines in particular. It is well known that conventional turbines are generally of lower efficiency than compressors. The reason would appear at least partly to be that a typical compressor has several blade stages of which only the last stage, usually a row of stator blades suffers uninhibited secondary losses which are of a relatively low order as the blades of the last row are not usually highly cambered; thus the uninhibited secondary losses are only a small proportion of the overall losses. On the other hand, the typical turbine comprises fewer stages of which the last stage, usually a rotor stage, has highly cambered blades, whose uninhibited secondary losses comprise a considerable proportion of the overall losses. The use of a row of inhibitor blades after the last turbine rotor stage will, therefore, substantially reduce the overall losses, giving a correspondingly greater efiiciency. The inhibitor blades are conveniently stationary and are so cambered, if at all, as to tend to reduce any residual whirl component of velocity in the fluid leaving the last rotor stage.

The invention is considered to be of greatest importance in its application to turbines having a single stage of rotor blading as are commonly used in aircraft and other propulsion plants comprising typically a row of nozzle blades followed by a row of very heavily cambered rotor blades with, in some cases, a small number of struts affording support for a bearing or exhaust cone usually at some remote point downstream of the rotor blades. The secondary losses are particularly heavy in such a turbine and are not sensibly inhibited by the struts due to their remoteness or small number or both. By providing, in accordance with the invention, a row of inhibitor blades of a numerical order comparable to or a large percentage of that of the rotor blade row and immedia't'elfadjacent thereto, a significant reduction in chord leiigth of the blades of the first mentioned row.

in order that the, invention may be fully understood it will now.v be described by way of example only in its embodiment in-a turbo-compressor jet propulsion plant with reference to the accompanying drawings in which:

Figure 1 represents a half cross section in a longi tudinal plane through the jet propulsion plant;

Figures 2 and 3 represent fragmentary longitudinal and transverse sections through the turbine blading of the plant of Figure l to an enlarged scale, Figure 3 being a section on the line III.III of Figure 2;

Figure 4 represents a cross-section through the successive blade rows as taken along section line lVIV of Figure 2;

Figures 5(a) and (b) represent velocity triangles relating to the blade sections of Figure 4;

Figure 6 represents a cross-section through the successive blade rows of an alternative form of blading to that of Figure 4;

Figures 7(a) and (b) represent velocity triangles relating to the blade sections of Figure 6.

In Figure 1 a turbine rotor 1 having a single row of blades 2 attached thereto drives through a shaft 3 a com pressor rotor 4 carrying several rows of blades 5 arranged in interdigital relationship with several rows of stationary blades 6 attached to the compressor stator 7. Compressed air delivered at the annular compressor outlet 8 is supplied to an annular combustion system 9 wherein fuel is injected by the injector 10 and the combustion gases are delivered through the flame tube 11 into the turbine annulus 12. This annulus contains three rows of circumferentially evenly spaced blading disposed succesively and immediately adjacent one another in the direction of fiow of the combustion gases (shown enlarged in Figure 2) comprising firstly a row of nozzle blades 13 attached to the turbine stator 14; secondly, the rotor blades 2; and finally, in accordance with the invention, a row of loss inhibiting blades 15 also attached to the turbine stator 14. A transverse cross-section through this last row of blades 15 (at IlI-III in Figure 2) is shown in Figure 3 from which it will be seen that the blades are similarly closely spaced to those of the usual nozzle or rotor blade row. After traversing the blade rows 13, 2 and 15, the gases pass through the exhaust duct 16 to the jet orifice 17 where they issue as a high velocity propulsive jet.

The cross-sections of the blade rows 13, 2 and 15, at their mean radius (IV-IV in Figure 2) are shown in the fragmentary view of Figure 4, and the corresponding velocity triangles for inlet to and outlet from the rotor blades at the designed operating conditions are shown in Figures 5(a) and 5 (b) respectively. The gases approach the nozzle blades 13 in a generally axial direction (indicated by the arrow A) and are delivered therefrom with a velocity V having a high circumferential component V the nozzle blades being highly cambered to effect this. Due to the fact that the circumferential velocity, U, of the rotor blades 2 is in the same general direction as V the velocity of the fluid relative to the rotor at inlets is reduced to V After traversing the rotor blades 2, which are necessarily highly cambered to derive sufficient energy from the fluid to drive the compressor, the gases leave with a relative velocity V having in this case a circumferential component equal and opposite to the blade velocity U so that the absolute velocity of the fluid at outlet from the rotor, V,,, is purely axial in direction, being equal in magnitude to the axial component of the absolute velocity at inlet to the rotor.

Consequently the gases enter the loss inhibiting blades 15 with a purely axial velocity, sothat these blades are accordingly of very simple symmetrical streamlined section. It follows that the losses associated with these blades are small as compared with those of the highly cambered rotor blades which are inhibited by the presence of the former. Furthermore, although the rotor blades will normally be twisted throughout their length to accommodate relative fluid velocity conditions at other radii ditfering from those at their mean radius, this need not be the case with the loss inhibiting blades since at all radii the fluid outlet velocity from the rotor blades can be arranged to be entirely axial in direction. Accordingly the inhibiting blades 15 are conveniently of untwisted form and bi-convex in cross-section. By the term untwisted, it is meant that the inlet and outlet blade angles are constant along the blade lengm.

In the alternative form of blading of Figure 6, the nozzle blades 13 are the same as in Figure 4 and the inlet velocity triangle to the rotor, Figure 7 (a), is also the same. In this case, however, the rotor blades 2 are arranged so that the relative circumferential velocity component of the fluid at their outlet, V in the outlet velocity triangle of Figure 7 (b), is greater than the blade speed U so that the absolute fluid velocity at outlet from the rotor, V,,, has a small circumferential component, the axial component, 1,, being unchanged. In this case the loss inhibiting blades 15 are conveniently slightly cambered to eliminate the circumferential component of V so that the fluid velocity at exit from the loss inhibiting blades is entirely axial in direction. It will be apparent that a small degree of camber in the loss inhibiting blades 15 will not significantly detract from their loss inhibiting characteristics in association with the highly cambered rotor blades 2.

Owing to the very small camber, if any, of the loss inhibiting blades, they are not subjected to any large circumferential loads and the drag loads of the passing fluid are also necessarily small. Consequently, the attachment of these blades does not present any difiiculties. As shown in Figures 2 and 3 they are attached only at the outer end, being positioned and secured by lands 18 formed on their roots 19 which are engaged by complementary grooves in the associated abutting stator sections 14 and 20.

Although the exhaust cone 21 in Figure 1 is supported from the stator section 19 by a small number of arms 22, this function could, of course, be performed by the loss inhibiting blade row.

What I claim is:

1. A turbine comprising a rotor partly defining a passage for the flow of working fluid, structure independent of said rotor also partly defining said passage and with respect to which said rotor is rotatable, and a number of blade rows each having circumferentially evenly spaced elongate blades of streamlined profile in transverse section extending longitudinally across said flow passage which blade rows are arranged successively in the direction of flow of the working fluid, wherein the last of said rows having blades shaped primarily for energy conversion and substantially cambered in section to derive considerable energy from the traversing fluid is on said rotor and is succeeded by .an immediately adjacent row of blades attached to said independent structure and being in number a substantial proportion of the number of the blades of the preceding row and whose chord length is smaller than the chord length of the blades of the preceding row, said last mentioned blades which are attached to said independent structure being uncambered in section.

2. A turbine according to claim 1, wherein the blades of said last mentioned row attached to said independent structure have blade inlet and outlet angles which are constant along the blade length.

3. A turbine comprising a rotor partly defining a passage for the flow of working fluid, stator structure also partly defining said passage, and a number of blade rows each having circumferentially evenly spaced elongate blades of streamlined profile in transverse section extending longitudinally across said flow passage which blade rows are arranged successively and immediately adjacent one to another in the direction of flow of the working fluid, of which the penultimate row traversed by the working fluid is attached to said rotor and is the last of said rows having blades shaped primarily for energy conversion and substantially cambered in section to derive considerable energy from the traversing fluid, wherein the blades of the last row traversed by the working fluid, being in number a substantial proportion of the number of the blades of said penultimate row, are attached to said stator structure and have a substantially symmetrical biconvex section, are untwisted throughout their length and having their chords lying in radial planes containing the axis of said rotor.

References Cited in the file of this patent UNITED STATES PATENTS 1,688,808 Gill Oct. 23, 1928 2,426,270 Howell Aug. 26, 1947 2,468,461 Price Apr. 26, 1949 2,488,783 Stalker Nov. 22, 1949 2,613,869 Anxionnaz Oct. 14, 1952 2,648,492 Stalker Aug. 11, 1953 FOREIGN PATENTS 116,512 Germany Dec. 31, 1900

Images