WO1992019869A1

WO1992019869A1 - Co-planar seal arrangement

Info

Publication number: WO1992019869A1
Application number: PCT/CA1992/000176
Authority: WO
Inventors: Clayton Bear
Original assignee: Nova Corporation Of Alberta
Priority date: 1991-04-30
Filing date: 1992-04-16
Publication date: 1992-11-12
Also published as: CA2041469A1; AU1552692A

Abstract

The axial forces imposed on a rotor of a fluid machine such as a compressor are controlled by modulating the pressure supplied to a fluid motor. The motor includes a pair of radially spaced seal assemblies extending between the end walls of the motor to define an annular cylinder. The nested seal assemblies provide a compact fluid motor.

Description

Co-Planar Seal Arrangement

TECHNICAL FIELD

The present invention relates to rotating fluid machines and in particular to fluid motors suitable for controlling the axial loads imposed on the rotor of the machine during operation.

BACKGROUND ART

Rotating fluid machines are used in a variety of applications to transfer energy between a fluid and a rotating mechanical system. Such machines include compressor which compress a gas in a continuous manner, pumps for pumping liquids and turbines for deriving useful work from a fluid flow. The machines usually have a housing with a fluid duct extending through the housing and one or more rotors rotating within the duct. The rotors rotate at a speed sufficient to cause a pressure differential between the inlet and outlet of the duct.

The rotors include an impeller mounted on a shaft which is, in turn, supported in the housing on bearing assemblies. Because of the high rotational speeds and close tolerances encountered within certain classes of machines, typically compressors, high demands are placed upon the bearing assemblies. Such assemblies tend to be expensive and of course must be designed to withstand the maximum load that may be applied for extended periods. This in turn increases the cost of the bearings.

Conventional hydrodynamic and antifriction bearings incur significant parasitic losses and during start-up the static friction in the bearings may be sufficient to prevent rotation of the rotor assembly subjecting it to adverse conditions.

DISCLOSURE OF INVENTION

Magnetic bearings are utilized in some applications to support the shaft for rotation and also to oppose axial loads on the shaft. Magnetic bearings avoid the limitations encountered in hydrodynamic and antifriction bearings, particularly at high speed, and, through control systems, permit dynamic adjustment of the bearings to maintain the shaft centered. However, the specific load capacity of a magnetic bearing is less than that of a mechanical bearing and so a physically larger bearing is required to withstand the loads typically encountered in a gas compressor. Moreover, where magnetic bearings are used, the typical loads imposed on the bearings result in a relatively large bearing assembly.

The loads imposed on the rotor of the compressor are caused in part by the pressure differential across the machine and also by the mass flow through the machine. Attempts have been made to reduce the axial loads caused by the pressure differential by utilizing a balance piston having one surface exposed to the high discharge pressure and the other surface exposed to the inlet or suction pressure. However, leakage occurs across the balance piston which may represent a substantial loss in machine efficiency. Moreover, the pressure differential and the momentum forces vary with different operating conditions of the machine so that a considerable axial force can still be generated during operation of the machine which must be accommodated by the bearings.

In these prior machines, gas seals were used between the rotor and housing to prevent the egress of gas from the fluid duct. For safety reasons, the seals were typically used in axially spaced pairs so that if one seal failed, the other seal would provide safe operation until the machines could be stopped and a repair could be effected. The seals used usually included components rotating with the rotor and these were kept as small as possible so as not to affect the rotor dynamics.

In U.S. Patent 4,993,917, assigned to the assignee of the present application, there is disclosed an arrangement in which advantage is taken of the seal assemblies to eliminate the balance piston. A fluid motor is provided by a pair of seals of differing effective diameters so that axial forces can be compensated by the action of fluid in the motor. While this arrangement is satisfactory in many applications, in practice the rotating fluid machine experiences a wide range of fluid pressure and flow combinations. To ensure that the axial loads opposed by the bearings are maintained within a given range for all machine operating conditions, the fluid motor must be capable of exerting a wide range of counter thrust on the rotor assembly. The magnitude of the counter thrust applied to the fluid motor equals the difference in pressure across the seal assembly times the effective area of the seal assembly. To accomplish this, the effective diameters of the two seal assemblies must be correspondingly sized, which leads to an increased size of at least one of the seal assemblies. As a result, the rotor dynamics mav be adverselv effected due to the large mass supported on the relatively long shaft to accommodate the fluid motor.

It is therefore an object of the present invention to provide a fluid motor for use in such machines which obviates or mitigates the above disadvantages.

According to the present invention, there is provided a rotating fluid machine having a housing, a fluid duct extending through the housing, a rotor assembly rotatably supported in the housing to be impinged by fluid flowing through the duct, a fluid motor acting between said rotor assembly and said housing, said motor comprising a pair of radially extending end walls axially spaced along said duct with one wall connected to said rotor and the other connected to said housing, and a pair of seal assemblies extending between said end walls at radially spaced locations to define an annular cylinder therebetween, each of said seal assemblies including a pair of sealing rings, a first of which is secured to said one wall for rotation therewith a second of which is secured to said other wall .

By providing a pair of radially spaced seal assemblies, extending between the end walls, the seal assemblies may be nested permitting a reduction of the axial length of the motor. The annulus formed between the seal assemblies provides an area differential to permit fluid within the cylinder to exert an axial force between the rotor assembly and housing.

According also to the present invention, there is provided a fluid motor for positioning between a pair of relatively rotating members comprising a pair of axially spaced end walls, secured to a respective one of said components for movement therewith, a pair of seal assemblies extending between said end walls at radially spaced locations to define an annular cylinder therebetween, each of said seal assemblies including a pair of sealing rings secured to respective ones of said end walls for movement therewith, whereby upon relative rotation between said components, one of said rings of each seal rotates relative to the other ring of its respective seal.

BRIEF DESCRIPTION OF DRAWINGS

Embodiments of the invention will now be described by way of example only with reference to the accompanying drawings in which

Figure 1 is a sectional view through an overhung compressor: Figure 2 is a view of a portion of Figure 1 on an enlarged scale;

Figure 3 is a schematic representation of the control circuit utilized to control the loads imposed on the rotor of the compressor of Figure 1 ;

Figure 4 is a sectional view similar to Figure 1 of a beam compressor; and

Figure 5 is a view of a portion of the compressor shown in Figure 4 on an enlarged scale.

Referring therefore to Figure 1 , a rotary fluid machine, in this embodiment, a compressor has a housing 10 with a fluid duct indicated generally at 11 extending between an inlet volute 12 and an outlet volute 14. The forward end of inlet volute 12 is defined by a wall 13 (commonly referred to as "a scoop") secured to a door 16 that closes the forward end of the housing 10.

A rotor assembly 17 including a shaft 18 is rotatably supported within the housing 10 by a bearing assembly 20. The bearing assembly 20 includes a bearing housing 22 having a magnetic thrust bearing 24 and a pair of magnetic radial bearings 26, 28 spaced apart on either side of the thrust bearing 24. The magnetic bearings 24, 26, 28 are conventional in nature as described for example in U.S. Patent 3,702,208 and will not be described herein in further detail. Conventional antifriction bearings 29 are also provided at spaced locations on the shaft 18 to provide emergency support for the shaft 18 in the event that the magnetic bearings fail.

The rotor assembly 17 further includes a pair of impellers 30, 32 located at the forward end of shaft 18 that rotate with the shaft 18. Flow from the inlet volute 12 past the impellers 30, 32 to the outlet volute 14 is controlled by a diaphragm assembly 34 comprising an inlet diaphragm 35, an interstage diaphragm 37 and a rear diaphragm 39. The assembly 34 is secured within the casing 10 and inlet vanes 36 direct gas to the first of the rotors 30. An internal passageway 38 directs gas from the discharge of the first impeller 30 to the inlet of the second impeller 32 with labyrinth seals 40 positioned between the rotor assembly 17 to seal between the impeller 30 and the diaphragm assembly 34. A fluid motor 42 is located between the impeller 32 and the bearing housing 22 to seal between the discharge and the shaft 18 and to enable adjustment of the axial loads imposed on the bearing 24 in a manner described below. Fluid motor 42 is shown in further detail in Figure 2 and includes a rotatable carrier 60 secured to shaft 18 and a stationary carrier 62 secured to the bearing housing 22. Carrier 60 includes a radially extending flange 64 with an axial nose 66 to define a pair of annular recesses 68, 70.

The carrier 62 also includes a radial flange 72 that is axially spaced from the flange 64 and has a pair of annular recesses 74, 76 directed toward the recesses 68, 70. An annular disc 78 is secured to the forward edge of the carrier 62 and carries a labyrinth seal 80 at its inner edge which engages the outer surface of the carrier 60. A labyrinth seal 81 is also located at the opposite end of carrier 62 to engage the outer surface of carrier 60.

A pair of dry gas seal assemblies 82, 84 respectively are located between the flanges 64 and 72. Each seal assembly 82, 84 is similar in construction and includes a pair of sealing rings 88, 90. The seal assembly 82 is of greater diameter than seal assembly 84 so that an area differential is provided between the two seal assemblies. Moreover, it will be noted that the circumferential wall at the inner diameter of the seal assembly 82 is greater than the circumferential wall at the outer diameter of seal assembly 84 so that the assemblies may be nested and define an annular cylinder 86 between them.

Each of the sealing rings 88 is located in a respective one of the recesses 68, 70 and secured by a dowl 92. An 0-ring 94 is located between the rings 88 and the carrier 62 to prevent gas passing behind the ring 88.

Each of the rings 90 is located in a respective one of the recesses 74, 76 and is axially moveable relative to the carrier 62 by virtue of splines 96.

A spring 98 acts through a thrust washer 100 to bias each of the rings 90 toward the ring 88 so that radial faces 102, 104 of the rings 88, 90 abut. The faces 102, 104 are configured to provide a pumping action for gas from a high pressure zone to a lower pressure zone and so provide a controlled leakage of gas across the seal upon relative rotation between the rings. The configuration of the faces is well known in the dry gas seal art and so will not be described in further detail. It will be noted that the abutting faces 104, 106 of each seal assembly 82, 84 lie in a common radial plane, perpendicular to the rotational axis of the shaft. A vent line 106 is provided in the carrier 62 and is used in a manner described below to control the pressure within the cylinder 86. A vent line 108 is also provided to evacuate gas passing through the seal assembly 84 and contained by labyrinth seal 81.

To control the axial forces imposed on the rotor, the pressure in the fluid motor 42 is regulated by the control scheme shown in Figure 3.

As can best be seen in Figure 3, the control line 106 is connected to a pressure control valve 110 that vents gas flowing through the control line 106 to a suitable vent. The pressure control valve 110 is controlled by a pilot pressure line 112 so that the pressure maintained in line 106 of the valve 110 is set by the pressure in the line 112. The pressure in line 112 is derived from a signal fed to a current-to-pressure converter 114 through signal line 116 that is itself connected to a ratio bias module 118. The ratio bias module receives a control signal from a tachometer 120 that senses the rotational speed of the shaft 18 in conventional manner. The tachometer 120 is also used to operate the speed control system indicated at 122.

It has now been recognized that the net axial force imposed on the shaft 18 varies with output speed.

To reduce the net axial forces, the control arrangement shown in Figure 3 is used to vary the pressure in chamber 86 as the rotational speed of the shaft 18 varies. The difference in effective diameter of seal assemblies 82, 84 provides an annular area that may be used to generate an axial force along the shaft 18. By varying the pressure of gas in the cylinder 86, the axial force exerted on the shaft 18 may also be varied. By correlating the pressure in the cylinder 86 to the rotational speed of that shaft 18, an appropriate axial force may be imposed on the shaft 18 to counteract the inherent axial forces generated by operation of the machine. This maintains the net axial force on the shaft 18 within a predetermined range over the range of normal operating speeds.

In operation therefore, the rotor assembly 18 is rotated by a suitable drive means and gas supplied to the inlet 12 is compressed and discharged through the outlet volute^' 14. A small flow of the high pressure discharge gas passes by labyrinth seal 80. The dry gas seal assembly 82 functions by permitting a controlled but very small amount of gas to flow between the relatively moving surfaces of the rings 88, 90. Thus a small amount of gas flows into the cylinder 86 where its pressure is applied between the end walls defined by flanges 64, 72 of carriers 60, 62 respectively. The pressure in cylinder 86 is controlled by the valve 110 to be maintained at the required level as determined by the tachometer signal.

Rotation of the rotor assembly 17 also generates a signal from the tachometer 120 which is applied to the ratio bias module 118. The ratio bias module may provide varying gains and varying offsets so that the desired output relationship to the input may be obtained. The input signal to the module 118 therefore produces the desired output signal in line 116 and sets the converter 114 at the required control pressure in line 112 to produce the desired pressure in control line 106.

As the speed of the compressor increases, the discharge pressure in volute 14 and the mass flow acting on the impellers 30, 32 increase. The mass flow may also vary depending upon the inlet and outlet conditions. The net effect typically is an increase in the axial thrust in the direction of the inlet volute due to increased pressure at the discharge volute 14. This may be offset in part by an increase in momentum forces toward the discharge volute 14. The pressure in the cylinder 86 is modulated so that the force acting through flange 64 of the carrier 60 away from the discharge volute is decreased. In this way, the net axial forces imposed on the thrust bearing assembly 24 are reduced, allowing for a smaller bearing assembly.

It will be seen, therefore, that by monitoring the speed of the compressor shaft 18 and utilizing that signal as an indication of end thrust, it is possible to reduce the variations in thrust forces imposed on the shaft 18 in a progressive and controlled manner. Moreover, the compact configuration of fluid motor 42 by virtue of the nested seal assemblies 82, 84 permits installation of the motor within the constraints of existing machines while providing the necessary control function.

An alternative form of compressor known as a beam type is shown in Figures 4 and 5 in which the shaft 18a is supported at laterally spaced locations. The operation of the compressor shown in Figures 4 and 5 is substantially similar in many respects to that of the overhung compressor shown in Figures 1 and 2 and therefore Iike reference numerals will be utilized to describe like components with a suffix "a" added for clarity. In the compressor shown in Figures 4 and 5, gas from the inlet volute 12a passes through rotor assembly 17a and into the discharge duct 14a. Of course, additional impellers 30a may be mounted upon the shaft 18a to provide multiple stages of compression if desired.

The shaft 18a is supported at spaced locations by radial magnetic bearings 26a and 28a respectively and axial forces are accommodated by a magnetic thrust bearing 24a at the forward end of the compressor. The bearings 24a and 26a are mounted outboard of an end 16a that closes the inlet volute 12a and utilizes a dry gas seal assembly 100 to prevent the flow of gas between the door 16a and the shaft 18a.

Control over axial loading of the shaft 18a is provided by a fluid motor 42a shown in more detail in Figure 5. Fluid motor 42a is substantially the same as motor 42, having a pair of radially spaced seal assemblies 82a, 84a defining an annular cylinder 86a between them. However, an alternative embodiment of the motor 42 is shown in Figure 5 where the cylinder 42a, the rings 88a of both seal assemblies are formed as an integral unit identified as 130 that extends radially along the flange 64a and is secured by dowl pin 92a. An 0-ring 94a seals between the unit 130 and flange 64a. This arrangement simplifies the construction of motor 42a by reducing the individual components. It will be appreciated that this arrangement could be used in the overhung compressor shown in Figures 1 to 3 and that separate sealing rings could be used in the beam type compressor of Figures 4 and 5.

Operation of the motor 42a to control the axial forces is similar to that described above with respect to Figures 1 to 3. It will be appreciated that alternative control strategies may be used, such as monitoring pressure differential. In each case, however, the compact configuration of the motor 42a permits incorporation in an installation that would otherwise be extremely difficult due to mechanical constraints.

The embodiments of Figures 4 and 5 illustrate the fluid motor at the discharge end of the compressor 10a. It will be appreciated, however, that the motor 42a could be incorporated in place of seal assembly 100 at the inlet end to control the net axial forces on the rotor assembly 17a with a conventional dry seal assembly utilized at the discharge end of the compressor.

As will be appreciated from considering the arrangement of the seals, the range of trust forces that can be generated is a function of the area differential between the seals and the maximum pressure that can be imposed in the cylinder 86. By selecting the seals so that their physical dimensions permit nesting, the effective area differential will be determined and thereafter the pressure range can be determined. Moreover, the nesting of the assemblies reduces the overall shaft length and enhances the rotor dynamics.

INDUSTRIAL APPLICABILITY

The present invention is suitable for use in rotary fluid machines, especially gas compressors. This invention assists with the control the axial loads placed upon such machines during their operation, and is capable of the assistance over a wide range of pressure and force conditions.

Claims

1. A fluid motor for positioning between a pair of relatively rotating components comprising a pair of axially spaced end walls, each of said end walls being secured to a respective one of said components for movement therewith, said fluid motor comprising a pair of seal assemblies extending between said end walls at radially spaced locations to define an annular cylinder therebetween, each of said seal assemblies including a pair of sealing rings secured to respective ones of said end walls for movement therewith, whereby upon relative rotation between said components, one of said rings of each seal rotates relative to the other ring of its respective seal.

2. The motor according to claim 1 wherein said rings of each seal abut on a radial plane.

3. The motor according to claim 2 wherein one of said rings of each seal is adjustable in an axial direction relative to the other of said rings.

4. The motor according to claim 3 wherein biasing means act on said one ring to bias it toward the other ring of the respective seal.

5. The motor according to claim 2 wherein said rings of both seals abut on a common radial plane.

6. The motor according to claim 4 wherein each seal permits a controlled leakage between said rings upon relative rotation between said rings.

7. The rotary fluid machine having a housing, a fluid duct extending through the housing, a rotor assembly rotatably supported in the housing and a fluid motor acting between said rotor assembly and said housing, said motor comprising a pair of radially extending end walls axially spaced along said duct with one wall connected to said rotor and the other connected to said housing, and a pair of seal assemblies extending between said end walls at radially spaced locations to define an annular cylinder therebetween, each of said seal assemblies including a pair of sealing rings, a first of which is secured to said one wall for rotation therewith and a second of which is secured to said other wall.

8. The machine according to claim 7 wherein said rings of each seal abut on a radial plane.

9. The machine according to claim 8 wherein one of said rings of each seal is adjustable in an axial direction relative to the other of said rings.

10. The machine according to claim 9 wherein biasing means act on said one ring to bias it toward the Other ring of the respective seal.

11. The machine according to claim 8 wherein said rings of both seals abut on a common radial plane.

12. The machine according to claim 10 wherein each seal permits a controlled leakage between said rings upon relative rotation between said rings.

13. The machine according to claim 10 wherein said one ring of each seal is connected to said one end wall for rotation with said rotor.

14. The machine according to claim 13 wherein a seal is located between the other of said rings and the end walls secured to said housing to inhibit the flow of fluid between the duct and said cylinder.

15. The machine according to claim 13 wherein both of said other rings are formed as a common component.

16. The machine according to claim 7 including a vent to permit fluid to pass into and out of said cylinder.

17. The machine according to claim 16 including control means operable on said vent to control the pressure of fluid in said cylinder.

18. The machine according to claim 17 wherein said control means is responsive to changes in axial loads imposed on said rotor to maintain said loads within a predetermined range.

19. The rotary fluid machine according to claim 1 wherein said rotor assembly is supported in said housing on axially spaced bearings, said rotor assembly including an impeller located between said bearings.

20. The rotary fluid machine according to claim 1 wherein said rotor assembly is supported in said housing on axially spaced bearings, said rotor assembly including an impeller located in advance of each of said bearings.