US3495921A

US3495921A - Variable nozzle turbine

Info

Publication number: US3495921A
Application number: US689397A
Authority: US
Inventors: Judson S Swearingen
Original assignee: Individual
Current assignee: Individual
Priority date: 1967-12-11
Filing date: 1967-12-11
Publication date: 1970-02-17
Anticipated expiration: 1987-02-17

Description

3 Sheets-sheaf. 1

FIG. I

62 FIG. 2

IN VENTOR.

ATTORNEYS JUDSON s. SWEAR/NGEN Feb. 17, 1970 I J s. SWEARINGEN vARf1ABLE ifiozzLE TURBINE Filed Dec. 11. 1967 V r o i. vii" Feb. 17, 1970 J- 9 SWEARINGEN 3,495,921

VARIABLE NOZZLE TURBINE Filed Dec. 11, 19s? 3 Sheets- -Shet 2 JUDSON S. SWEARINGEN INVENTOR.

f ,W* g ATTORNEYS Feb. 17, 1970 J. s. SWEARINGEN 3,

. VARIABLE NOZZLE TURBINE Filed Dec. 11. 1967 s Sheets-$heet s V 80 Q7 4 v mnnnnrl v JUDSON s. SWEAR/NGEN I N VEN TOR.

BY M ATTORNEYS United States Patent M 3,495,921 VARIABLE NOZZLE TURBINE Judson S. Swearingen, 500 Bel Air Road, Los Angeles, Calif. 90024 Filed Dec. 11, 1967, Ser. No. 689,397 Int. Cl. F01d 17/08, 17/00 US. Cl. 415163 18 Claims ABSTRACT OF THE DISCLOSURE The turbine disclosed has a variable nozzle assembly which is formed of a plurality of nozzle blades sandwiched between a pair of axially-spaced concentricallymounted rings. The blades are pivotally attached to the two rings. One of the rings is fixed and the other ring, which is an actuator ring, is sealingly mounted on a bearing ring for rotation relative to the fixed ring. Rotation of the actuator ring alters the spacing between complementary portions of adjacent blades. The blades are so spaced that such complementary portions form the passages from the turbine inlet to the turbine wheel.

There is a pressure drop through the nozzles and the outer face of the actuator ring upstream of the seal is subject to high pressure, and downstream of the seal is subject to low pressure. The upstream portion of the blade face of the actuator ring is also subject to high pressure and the downstream side of the blade face of the cam ring is subject to low pressure. However, ordinarily, changes in position of the blades will change the location of respective demarcation line dividing the high and low pressure areas on the blade side of the actuator ring, such line being hereinafter called the pressure-break line. Means are provided in this disclosure to control the pres sure break-line so that between the closed and open position of the blades it will remain in approximately the same position. Such means are illustrated as being a tapered surface on the blade side of the actuator ring so that the actuator ring and the blades are in sealing contact only along a circle adjacent and upstream of the pivot pins mounting the blades. Therefore, a major portion of the upstream side of the actuator ring which is shielded by the blades is relieved so that the area exposed to the high pressure will remain approximately the same regardless of the position of the blades. To assure balanced forces on both sides of the nozzle blades, the fixed ring may also be tapered in a similar fashion. To accomplish full balance, the rings are tapered downstream of the pressure-break line as well as upstream.

Alternate constructions disclosed a nozzle blade which forms an acute angle with the axis of the turbine wheel and a split actuator ring having provisions to prevent over-extension.

BACKGROUND OF THE INVENTION This invention relates to turbines and, more particularly, to turbines having variable nozzle assemblies.

In one type of a variable nozzle turbine there is a plurality of tiltable blades sandwiched between a pair of axially-spaced, concentrically-mounted rings. Complementary portions of adjacent blades form the nozzle openings. Each blade is pivoted on a pin located in a stationary ring. Each blade also has an elongated slot which is engaged with a cam pin located in an opposing, rotatable actuator ring. Rotation of the actuator ring causes the pins in the elongated slots to vary the position of the blades, thereby changing the size of the nozzle openings.

If the actuator ring is axially fixed, there is either an excessive clearance or else a danger of jamming. For example, a clearance of .001 inch will permit about a 3,495,921 Patented Feb. 17, 1970 1% blow-by while at the same time such a clearance is not sufficient to prevent jamming. Therefore, it is preferable that the actuator ring be free to move axially. With such construction, there is no necessity for clearance because, while the actuator ring is urged against the blades by pressure to effect a seal, the actuator ring can move away to prevent jamming.

There is a pressure drop through the nozzles and high pressure acts against the outer face of the actuator ring upstream of the seal diameter and low pressure acts against the face downstream of the seal diameter. Both, together, constitute a force which urges the actuator ring against the nozzle blades. An opposing force prevails on the nozzle blade side of the actuator ring. The force on the blade side of the actuator ring is usually somewhat less than the force against the outer face so that the actuator ring is urged gently against the blades and not away from them. While it is possible to make the seal diameter large enough so that the net force is negative, it is desirable to select a seal diameter which will provide a definite but mild force clamping the nozzle blades. While the magnitude of this force should be sufficient to assure no leakage between the actuator ring and the nozzle blades, it must also be small enough so that the actuator ring will be able to move away from the nozzle blades to let particles pass.

In the case of high-pressure application, the difference in the pressure-break line on the blade side and the seal diameter on the outer face of the actuator ring becornes critical, inasmuch as this difference determines the clamping force on the blades. It has been found that in a structure in which the blades and actuator ring engage each other throughout their opposed surfaces, the position of the nozzle blades normally alters the diameter of the pressure-break lines on the blade side.

Therefore, when a seal diameter is selected which will prevent a negative force with the nozzle blades in the closed position, the pressure-break line diameter on the blade side with the blades in open position will shift, exposing a greater area to high pressure, thereby creating an excessive clamping force. The high clamping force with the blades in the open position makes it difficult if not impossible, to shift the actuator ring and causes the unlubricated sliding surfaces to gall. Accordingly, it is a purpose of the present invention to control the pressure-break line diameter on the blade side of the actuator ring so that the diameter of the pressure-break line will remain fairly constant regardless of the position of the blades, whereby it is possible to locate the seal diameter so that a small but definite and practically constant clamping force will prevail between the rings and the nozzle blades.

It is another object to provide an improved turbine having a variable nozzle assembly formed of spaced rings having a plurality of pivoted nozzle blades therebetween wherein the actuator ring is free to move axially along a low-friction sealing surface whereby a slight clamping pressure may prevail between the rings and the nozzle blades.

It is still another object to provide a turbine with an improved variable nozzle assembly having an axially movable actuator ring which exerts a small but definite and constant clamping force between the rings and the nozzle blades and in which a substantial portion of the area of the blade side of the actuator ring shielded from pressure by the blades is relieved so that regardless of the position of the blades the exposed area will remain virtually constant.

It is a further object to provide an improved variable nozzle turbine having improved nozzle blades to improve flow characteristics.

It is still a further object to provide an improved variable nozzle turbine having an actuator ring having a slot whereby tangential thrust generates an outward force to reduce friction and wherein there is a stop to prevent over-expansion.

SUMMARY OF THE INVENTION The turbine assembly of the present invention is comprised of a housing having a fluid inlet and a fluid discharge. A turbine wheel is rotatabl mounted on the axis of said housing. A variable nozzle assembly formed of a pair of coaxially-spaced, coaxially-mounted rings and a plurality of pivoted blades sandwiched therebetween is positioned between the inlet and the turbine wheel. The spacings between complementary portions of adjacent nozzle blades form the throats for the nozzle passages. One ring is fixedly-mounted and the other ring, which is the actuator ring, is mounted for angular rotative movement relative to the first. The actuator ring is sealinglymounted on a bearing and is capable of axial movement. A plurality of fixedly-positioned pins extend outward from the fixed ring and the individual blades are pivotally-mounted on said pins. An elongated slot is located in each blade and a pin extending from the actuator ring is engaged therewith, whereby relative movement of the actuator ring varies the position of the blades. Means are provided to apply a tangential thrust on the actuator ring to rotate it concentrically with the axis of the turbine and thereby vary the opening of the nozzles.

High pressure acts on the outer face of the actuator ring upstream of the seal and low pressure acts on the outer face downstream of the seal. The blade side of the actuator ring is also subject to opposing pressures. The pressure-break line on the blade side in prior structures normally changes with the relative position of the nozzles. Accordingly, means are provided by this invention to maintain said pressure-break line practically constant regardless of the position of the nozzle blades so that the seal diameter may be so relatively located that a slight but definite pressure is exerted by the actuator ring on the nozzle blades thereby sealing the nozzle blades and directing all of the fluid through the nozzle passages while at the same time permitting easy adjustment of the actuator ring and permitting axial movement of the actuator ring away from the nozzles to prevent jamming.

One means of maintaining the diameter of the pressurebreak line constant is to taper the surface of the actuator ring outwardly from a line adjacent but upstream of the pivot pins of the fixed ring so that there is a space between the blades and the opposing face of the actuator ring, whereby approximately the same area is subject to high pressure regardless of the position of the blades. If desired, the opposing wall of the fixed ring will also he correspondingly tapered. Although movement of the tip portions of the blades does not make as large a change in area and the pressure in such region is low pressure, further equalization may be accomplished by discontinuous tapering of the downstream region of the blade face of the actuator ring.

The actuator ring may he a solid ring mounted on lowfriction bearing and sealing surface or, if desired, the actuator ring may be a slotted ring having a stop to prevent over-expansion. In the latter case, a tangential thrust is applied to the one side of the slot to provide rotation and thereby alter the spacing of the blades, such thrust generates an outward radial thrust which reduces friction. Also, the faces of the blades opposing the turbine wheel may be at an acute angle to the axis of the wheel to reduce friction of the flow through the nozzle passages.

BRIEF DESCRIE'TION OF THE DRAWINGS FIG. 1 is a cross section through a variable nozzle turbine constructed in accordance with the present invention, the section being take generally along the axis of the turbine;

FIG. 2 is an enlarged cross-sectional view to show the details of construction through a variable nozzle blade;

FIG. 3 is a plan view of a portion of the variable nozzle assembly showing the nozzle blades in the open position in full line and in closed position in dashed line;

FIG. 4 is a diagrammatic view of the nozzle blades illustrating the difference in pressure-break line between the open and closed position in prior art construction;

FIG. 5 is a view similar to FIG. 4 illustrating the difference in pressure-break line when the present invention is utilized;

FIG. 6 is a view similar to FIGS. 4 and 5 illustrating further improvements in retaining the pressure-break line position constant;

FIG. 7 is a cross section view having an alternate nozzle blade construction;

FIG. 8 is an isometric view illustrating the nozzle blade disclosed in H6. 7; and

FIGS. 9 and 10 are plan views illustrating alternate constructions for the cam ring.

DESCRIPTION OF THE PREFERRED EMBODIMENT Referring now to the drawings, and in particular FIG. 1, there is illustrated a variable nozzle turbine having a housing 12 provided with a fluid inlet 14 and an axial fluid discharge 16. Between the inlet and discharge is a turbine wheel compartment 18 in which is located a turbine wheel 20 mounted on a shaft 22 which extends through a sealed opening 24 in a closure member 26 which sealingly closes an opening in housing 12. The turbine wheel 20 is provided with a plurality of radially- .axially extending passages 28 which are designed to receive fluid from inlet 14 and direct it through turbine wheel 20 for discharge through axial discharge 16. A variable nozzle assembly 29 which controls the entry of fluid from the inlet surrounds turbine wheel 20.

The variable nozzle assembly 29 is formed of a pair of

rings

30 and 32 and a plurality of individuallypivotable nozzle blades 34 which are sandwiched between the two rings. The nozzle blades 34 are so mounted that the spacings between complementary portions of adjacent blades define throats forming nozzle passages 33. The ring 30 is fixedly attached to the housing. The actuator ring 32 is axially-spaced from fixed ring 30 and is mounted concentrically about a cover plate 36 attached to the housing. Suitable means for providing rotative movement to actuator ring 32, such as an actuator rod 35, is provided. The cover plate forms a portion of axial discharge 16. A rotating seal 37 is provided between the outlet end of passages 28 and a surrounding circumferential portion 38 of cover plate 36. Another seal 39 is formed on the inner side of turbine wheel 26 between the wheel and a cylindrical portion 41 of closure 26. Accordingly, all fluid which enters through variable nozzles 33 is directed through turbine wheel passages 28 and exits through the axial discharge 16.

Actuator ring 32 is mounted on a cylindrical bearing ring 40 of cover plate 36. The surface of the bearing ring opposing actuator ring 32 is provided with a groove 42 in which is positioned a wafer-type spring expansion ring 44 and a seal ring 46 which may be formed of a low-friction material such as polytetrafluorethylene, com- .monly referred to by its trademark name Teflon or other suitable material. Seal ring 46 bears against inner circumferential surface 48 of actuator ring 32. If desired, inner surface 48 may be of stepped design to provide a narrow metal-to-metal bearin contact between the actuator ring and the bearing surface.

As can be seen in FIG. 2, each blade 34 is pivoted on a pivot pin 50 journaled in fixed ring 30. Each individual blade is provided with an elongated slot 52 in which a cam pin 54 extending from actuator ring 32 is engaged. Accordingly, any rotative movement of actuator ring 32 results in a movement of cam pin 54 which, through slots 52, results in a translating movement, varying the position of blades and changing the relative position of cam end 53 of one blade relative to tip end 55 of an adjacent blade whereby the size of the nozzle passage 33 between two adjacent blades is varied.

As previously mentioned, the space between actuator ring 32 and bearing ring 40 is sealed by seal ring 46 and, as can be seen, actuator ring 32 is free to move axially along cylindrical surface 40 of cover plate 36. There is a pressure drop through nozzles 33 and high pressure acts against outer face 56 of actuator ring 32 upstream of seal diameter SD and low pressure acts against that portion of face 56 which is downstream of seal diameter SD. An opposing force prevails on nozzle blade face 58 of actuator ring 32. The force of blade side 58 is usually somewhat less than the force on outer face 56 whereby actuator ring 32 will be urged gently against blades 34 and not away from them. While seal diameter SD may be made so large that the net force clamping the nozzle blades is negative, it is desirable to select a seal diameter which will provide a definite but mild force clamping the nozzle blades. The magnitude of this force must be small so that actuator ring 32 will be able to move away from nozzle blades 34, if necessary, so as to let some particles pass, if any occur in the gas being processed, to keep them from becoming caught between the ring and blades and causing wear or malfunction.

In high-pressure applications, the difference in pressurebreak line on the blade side and seal diameter SD on the opposite side of the actuator ring becomes critical since this difference determines the clamping force on the blades.

It has been found that the position of nozzle blades 34 alters the average pressure-break line on the blade side. This is due to the fact that when there is a positive clamping force the blades will shield some of the area of the blade side of the actuator ring exposed to pressure and the amount shielded in the one position will differ from the amount shielded in the other positions. Thus, the average area of the blade side subject to high pressure when the blades are in the open position is much smaller than when the blades are in the closed position. Therefore, if a seal diameter is selected to prevent a negative force in the closed position an excessively high clamping force will be created when the blades are in the open position. T hehigh clamping force makes it diflicult, if not impossible, to shift the actuator ring and may cause the unlubricated sliding surfaces to gall.

Referring now to FIG. 4, several typical nozzle blades are shown pivoted on pivot pins 50. They are shown in solid outline in the open position and in dashed outline in the closed position. As can be seen, the nozzle blades are moved from one position to the other by the circumferential translating movement of cam pin 54 in slots 52. The solid line PBD shows the estimated average pressure locus or pressure-break line between the high pressure and the low pressure areas on the blade side of the actuator ring for the open position, and dashed line PBC shows the average pressure locus or pressure-break line for the closed position. If the area enclosed by these two lines, as shown in FIG. 4, differs by about /2 sq. in. per blade and if the pressure difference is 1000 p.s.i., the blade in the open position is clamped with a force of 500 lbs. greater than in the closed position. If the coefiicient of friction is of the order of .8 and acts on both sides of the blade, it will take a force of the order of at least 800 lbs. to move one blade. For 16 blades, this is an unacceptable load for a control device and metals this heavily loaded usually will not stand a dry rub between them without damage.

The reason for the difference in the pressure-break lines between the open and the closed position is that as cam end 53 of the blade rises into the upper high pressure zone it brings the average pressure-break line with it, whereas on tip end 55 of the blade the change, although in the opposite direction, is very small so that the net change is substantially outwardly.

In order to minimize the difference in the clamping load between the open and the closed position, the present invention brings the two pressure-break lines closer together. This is accomplished by relieving certain portions of the mating faces of actuator ring 32 and blades 34. As can be seen in FIG. 2, surface'60 of actuator ring 32 and surface 62 of fixed ring 30 are tapered outward from a circle adjacent the upstream side of pivot pins 50. Therefore, the blade side of actuator ring 32 is subject to the high pressure outside of the line PB at all times independent of the position of the nozzle blades. At the same time, a seal is still established bet-ween the two rings and the blades in the area'around the pivot pins which surfaces are not relieved. Accordingly, as can be seen in FIG. 5, the average diameter of the pressurebreak line PBD for the open position and the pressurebreak line PBC for the closed position are almost the same. The cam end 53 of end blade 34 still rises into the high pressure area as the nozzles open, but does not bring the pressure break-line with :it because the taper on the actuator ring exposes it to the high pressure down to a line almost through, but just outside of, pivot pins 50. As can be seen in FIG. 5, the area between the lines P'BD' and PBC' is much less than that between PBD and PEG in FIG. 4 and this degree of balance is usually sufficiently close to prevent galling and maintain the force required to rotate the actuator ring at a satisfactory level. Although it is not necessary to have a corresponding taper on fixed ring 30, it is desirable to do so to assure balanced forces on both sides of the nozzle blades. Such tapering makes the system symmetrical and eliminates any tendency for the nozzle blade to be pressed against the fixed ring by some unusual pressure distribution.

If complete balance is desired, it can be accomplished by further relief, such as by tapering surface 64 of actuator ring 32 and surface 66 of fixed ring 30 downstream of pivot pins 50. This additional relief as shown in FIGS. 2 and 6 is discontinuous covering only the area defined by X, Y and Z, that is to say, the taper surface is discontinuous between lines X and Z. If the actuator ring had a fully tapered surface, some of the effectiveness would be lost since there would be blow-by under the tip of the nozzle blade which would be undesirable. As can be seen in FIG. 6, the estimated pressure-break line P"B"C" for the closed position is less than the pressurebreak line P"BD" for the open position. This is the reverse of that depicted in FIG. 4 and demonstrates the possibility of reaching a point of equality between the two pressure-break lines.

Accordingly, it can be seen that by relieving the influence of the cam ends of the nozzle blades when they are moved from one position to another the pressurebreak line remains fairly constant; therefore seal diameter SD on the outer side of actuator ring 32 can be accurately determined to provide the desired clamping force. While tapered surfaces are shown as preferred means for relieving the pressure caused by shifting of the nozzle blades from one position to another, other means such as a circumferential groove may also be utilized. Any means capable of equalizing the shielding of the blade side of the actuator ring from pressure regardless of position of the nozzle blades may be utilized.

Although the tapered surfaces provide a passage which permits the upstream high pressure to be in constant communication with the blade side of the actuator ring and the opposing faces of the blades, such tapered surfaces terminate upstream of the circle formed by the fixed pivot pins. Therefore, provided a positive clamping force exists, the actuator ring will be forced against the blades and the blades against the fixed ring and a seal will be established in the area adjacent the pivot pins and there will be no blow-by.

The nozzle blade illustrated in FIGS. 1 through 5 has its end portion 55 parallel to the axis of turbine wheel 20, Le, the surface of the blade closest to the axis of the turbine wheel is generated by a line which moves parallel to the axis of the turbine wheel. In nozzle blade 80, illustrated in FIGS. 7 and 8, there is an end portion which is at an acute angle to the axis of turbine wheel 86 and which, if projected, will intersect the axis of the turbine wheel. The surface 84 of blade 80, which surface corresponds to the underneath surface of the forms of FIGS. 1-6, is generated by a line which is positioned at an angle to the axis of the turbine wheel. The angle at the end 88 is an acute angle. The underneath surface may be generated by such an angularly-positioned line whose angularity from end 88 to the heel of the blade is either constant or decreases in angularity as it moves from end 88 to the heel of blade 80.

In the form shown in FIG. 8, the surface is generated by a generating line of changing angularity as described above. The angularly-positioned surface 84 of the blade has an important advantage. The fluid discharging from the nozzle into passageways 90 of wheel 86 is not subjected to any material friction loss since it is not subjected to flow over frictional surfaces between the surface 84 and the passageway entrances. The shroud 92 and the disc 94 of turbine wheel 86 acting together with the vanes from passageways 90 through turbine wheel 86 whose inlet at the nozzle has a direction with a substantial axial component.

The construction described above minimizes the area of surface exposed to the high-velocity stream from the nozzle to the entrance of turbine wheel 86.

The actuator ring 32 illustrated in FIGS. 1 through 6 may be a solid ring riding on the low-friction Teflon seal ring 46. Inasmuch as the seal ring is formed of lowfriction material, the frictional drag during rotative movement of the ring is kept to a minimum. However, if desired, the actuator ring may be provided with a slot and the tangential thrust made at one side of the slot to rotate the actuator ring. The tangential thrust generates an outward radial thrust on the actuator ring and causes the friction between the actuator ring and the bearing ring to be reduced resulting in easy rotative movement of the actuator ring. The slot may be merely a slit in the ring, as for example a saw cut, or it may take the forms shown in FIGS. 9 and 10, in which case there is a stop to prevent overstrain of the ring when the tangential pressure is applied. This prevents any overstrain which could possibly cause the ring to fail.

As shown in FIG. 9, the slot 100 is formed to produce an internal stop. The slot is formed with a bore 101 to receive an actuator member 102. The slot 100 cuts the ring across the bore and extends circumferentially at 104, makes a reverse U-bend at 105 and extends radially across ring 106; however, being stepped at 107 and 108 to form an offset labyrinth out which minimizes any fluid leakage through the slot. The reverse U-bend at 105 acts as an internal stop to limit expansion of ring 106.

In the form shown in FIG. 10, a slot 110 extends radially across ring 111, being stepped at 112 and 113 to form the labyrinth cut which minimizes fluid leakage. The slot has a bore 114 in which an actuator member 115 is located. The ring is provided with an externallypositioned dog 116 and stop 117 positioned on each side of a slot 110 forming an effective stop against undue expansion of the ring. Therefore, as can be seen, the slots are so designed that the actuator ring cannot be unduly expanded, therefore preventing any overstrain of the actuator ring.

From the foregoing it will be seen that this invention is one well adapted to attain all of the ends and objects hereinabove set forth, together with other advantages which are obvious and which are inherent to the apparatus.

It will be understood that certain features and sub-' combinations are of utility and may be employed without reference to other features and subcombinations. This is contemplated by and is within the scope of the claims.

As many possible embodiments may be made of the invention without departing from the scope thereof, it is to be understood that all metter herein set for.h or shown in the accompanying drawings is to be interpreted as illustrative and not in a limiting sense.

I claim:

1. A turbine assembly comprising a housing, a fluid inlet and a fluid outlet in said housing, a turbine wheel rotatably mounted on an axis in said housing, a pair of rings coaxially mounted about said axis in said housing, the first of said rings being fixedly-mounted in said housing and the second of said rings mounted on a bearing for angular rotative movement relative to the first of said rings and movable axially along the bearing, a seal between said second ring, which is an actuator ring, and its bearing, a plurality of adjacent nozzle blades mounted between said rings at spaced intervals about said rings, the spacing between said complementary portions of adjacent nozzle blades forming a throat for a nozzle passage, means including said actuator ring for altering the spacing between said complementary portions of said nozzle blades, there being an area exposed to high pressure on the outer side of said actuator ring upstream of said seal, and an area exposed to low pressure downstream of said seal and an area exposed to upstream high pressure and an area exposed to downstream low pressure on the blade side of said actuator ring, a potion of the blades moving into the high-pressure area in their nozzle open position, and said blades and actuator being formed to provide spaces therebetween and expose to high pressure the areas of said actuator ring in the high pressure area into which said blades so move for maintaining the area on the blade side of the actuator ring subject to high pressure constant within predetermined limits regardless of the position of the blades.

2. The turbine assembly specified in claim 1, wherein there is a passage between the blade side of the actuator ring and the opposing face of the blades which is subject to upstream high pressure, such passage maintaining the area of the blade side of the actuator ring subject to high pressure constant within predetermined limits regardless of the position of the blades.

3. The turbine assembly specified in claim 1, wherein the blades are pivotally mounted on pivot pins on the fixed ring, the surface of the blade side of the actuator ring is tapered outward of a seal area circumscribed by a circle adjacent the upstream side of the pivot pins whereby pivotal movement of the blades will not materially vary the effective area subject to high pressure.

4. The turbine assembly specified in claim 3, wherein the surface of the fixed ring corresponding to the tapered surface of the actuator ring is correspondingly tapered.

5. The turbine assembly specified in claim 3, wherein the surface of the blade side of the actuator ring downstream of the seal line is provided with outwardly tapered portions so that movement of the blades in o and out of such portions will not materially vary the effective area subject to low pressure.

6. The turbine assembly specified in claim 5, wherein the surfaces of the blade side of the fixed ring upstream and downstream of the pivot pins are provided with tapered portions corresponding to the tapered portions on the actuator ring.

7. The turbine assembly specified in claim 1, wherein the actuator ring has a radially-extending slot and the blade space altering means applies a tangeniial thrust at one side of said slot.

8. The turbine assembly specified in claim 7, wherein said radial slot is in the form of a labyrinth passage to minimize flow leakage therethrough.

9. The turbine assembly specified in claim 8, wherein said actuator ring has stop means cooperating with the slot to prevent undue expansion of said actuator ring by the application of the tangential thrust.

10. The turbine assembly specified in claim 9, wherein said stop means is formed of an externally-positioned first stop member on the outer periphery of the actuator ring on one side of the slot and a second member on the outer periphery of the actuator ring on the other side of the slot having a recess cooperating with the first stop member.

11. The turbine assembly specified in claim 9, wherein said stop means is formed by the slot extending radially inward from the outer edge, then circumferentially, and then having an inverted U-bend, one leg of which extends toward the inner edge.

12. The turbine assembly specified in claim 1 wherein a plurality of pivot pins extend from the fixed ring and the individual nozzle blades are pivoted on said pins, the individual nozzle blades and the actuator ring having inter-engaging, elongated slots and cam pins which engage with said slots respectively whereby rotation of the actuator ring causes the nozzle blades to pivot about their pivot pins thereby changing the spacing between complementary portions adjacent nozzle blades, the surface of the nozzle blades adjacent to the turbine wheel being parallel to the axis of the turbine wheel.

13. The turbine assembly specified in claim 1, wherein a plurality of pivot pins extends from the fixed ring and the individual nozzle blades are pivoted on said pins, the individual nozzle blades and the actuator ring having inter-engaging, elongated slots and cam pins which engage with said slots respectively whereby rotation of the actuator ring causes the nozzle blades to pivot about their pivot pins thereby changing the spacing between complementary portions of adjacent nozzle blades, the surface of the nozzle blade adjacent to the turbine wheel being at an acute angle to the axis of the turbine wheel.

14. A turbine assembly comprising a housing having a high pressure fluid inlet, and a low-pressure fluid discharge, a turbine wheel rotatably mounted on an axis in said housing, a pair of rings coaxially mounted about said axis, one of said rings being fixedly mounted in said housing and the other of said rings being an actuator ring mounted for rotative movement relative to the fixed ring, a plurality of adjacent nozzle blades mounted on and between said rings and movable to provide larger or smaller nozzle openings, said actuator ring having an outer circumferential surface in fluid communication with high pressure from said inlet and an opposite inner ring surface, a cylindrically outwardly facing bearing surface in said housing, the inner surface of said actuator ring opposing said bearing surface, an annular recess in said bearing surface, a low-friction sealing ring in said recess sealingly engaging said inner ring surface of said actuator ring, the outer face surface of said actuator ring being in fluid communication with said inlet upstream of said seal and the outer surface of said actuator ring downstream of said seal being in fluid communication with the pressure downstream of the nozzle blades, the blade side of said actuator ring also having an area exposed to said pressure from said inlet whereby there are opposing axial pressures on said actuator ring, and the blade side of said actuator ring and the opposing face of the blades being exposed to said high pressure whereby a change in position of the nozzle blades will not materially affect the area on the blade side of the actuator ring subject to the high pressure.

15. The turbine assembly specified in claim 14 wherein the blades are mounted on pivot pins and the upstream surface of the blade side of the actuator ring is tapered outward of a seal line extending through the pivot pins to vent said area so that movement of the blades will not materially vary the effective area subject to high pressure.

16. The turbine assembly specified in claim 15, wherein the upstream surface of the fixed ring opposing the tapered surface of the actuator ring is correspondingly tapered.

17. The turbine assembly specified in claim 14 wherein the downstream surface of the blade side of the actuator ring is provided with tapered portions so that movement of the blade into and out of such portions will not materially vary the effective area subject to low pressure.

18. The turbine assembly specified in claim 17, wherein the upstream and downstream surfaces of the blade side of the fixed ring opposing the tapered surfaces of the actuator ring are provided with correspondingly tapered portions.

References Cited UNITED STATES PATENTS 1,197,761 9/1916 Pfau 253-122 3,232,581 7/1963 Swearingen 253122 3,243,159 3/1966 Hefler et al 253-122 EDWARD L. MICHAEL, Primary Examiner