Rotor for a turbomachine

Abstract

A rotor for a turbomachine is provided. The rotor includes a shaft and a longitudinal bearing disk arranged thereon as an element for an axial bearing for axially supporting the shaft. The longitudinal bearing disk includes first and second contact surfaces. The first contact surface is cylindrical for radial support on the shaft. The second contact surface is conical for self-centering on the shaft. Advantageously, such a rotor may be used for operation with chemically aggressive gases and can be operated at substantially high rotational speeds.

Description

CROSS REFERENCE TO RELATED APPLICATIONS

This application is the US National Stage of International Application No. PCT/EP2010/054210, filed Mar. 30, 2010 and claims the benefit thereof. The International Application claims the benefits of German application No. 10 2009 015 859.6 filed Apr. 1, 2009. All of the applications are incorporated by reference herein in their entirety.

FIELD OF INVENTION

The invention relates to a rotor for a turbomachine, having a shaft and a longitudinal bearing disk, which is arranged on said shaft, as an element for an axial bearing for axially bearing the shaft.

BACKGROUND OF INVENTION

Working elements, for example in the form of paddle wheels, are arranged on a shaft of a turbomachine, such as a gas or steam turbine or a turbocompressor, a stationary pressure difference between an inlet and outlet of the turbomachine being increased or decreased by said working elements. A large force is transmitted onto the shaft in the axial direction of the shaft by operation of the working elements, said force being absorbed by an axial bearing. The axial bearing comprises a longitudinal bearing disk, which is part of the rotor, and bearing elements on the stator, the longitudinal bearing disk being supported on said bearing elements—by magnetic forces or by a lubricated sliding action depending on the type of bearing.

In the case of industrial turbomachines, in particular in the case of those which are used in the chemical industry, rotation speeds in the region of several 10000 revolutions per minute are sometimes required. Rotation speeds of such a high level require enormous strength both of the working elements and of the longitudinal bearing disk which has a relatively large radius for absorbing and passing on a high axial thrust and therefore is subject to large centrifugal forces. In order to ensure the required high strength, high-alloy steels with a yield strength of around 1000 N/mm2 are usually used.

If a turbomachine, in particular in the industrial sector, comes into contact with aggressive gases, for example during compression of hydrogen sulfide-containing gases, the working elements and also the longitudinal bearing disk are chemically corroded and therefore their strength is impaired. However, steels which are not corroded by hydrogen sulfide only have a maximum strength of 700 N/mm2. Therefore, turbomachines which come into contact with hydrogen sulfide or similarly aggressive gases can be operated only at relatively low rotation speeds as machines which are not designed for exposure to such aggressive chemical gases.

SUMMARY OF INVENTION

It is an object of the present invention to specify a rotor for a turbomachine, which rotor can be designed for operation with chemically aggressive gases and can be operated at high rotation speeds in the region of greater than 10000 revolutions per min.

This object is achieved by virtue of a rotor of the type cited in the introductory part, in which, according to the invention, the longitudinal bearing disk has a first, cylindrical contact face for radial mounting on the shaft and a second, conical contact face for automatic centering on the shaft.

In this case, the invention proceeds from the consideration that longitudinal bearing disks of known design are mounted on the shaft of the rotor by means of a hydraulic shrink-fit. In this case, the longitudinal bearing disk is hydraulically widened and pushed onto a slightly conical contact face of the shaft against a stop. A shaft nut holds the shrunk-on longitudinal bearing disk axially in position. In order to prevent the longitudinal bearing disk slipping on the contact face of the shaft, said longitudinal bearing disk still has to be able to retain its stable press-fit on the contact face even when subject to strong centrifugal forces, this requiring the press-fit to be very fine. As a result, the steel of the longitudinal bearing disk is subject to a very high level of stress.

If a stable connection is also achieved without a shrink-fit, the forces on the longitudinal bearing disk created by the shrinking process are not applicable and said longitudinal bearing disk can use all its strength to withstand the centrifugal forces. By reducing the produced stresses on account of the stresses from the shrink-fit being removed, materials for the longitudinal bearing disk which are suitable for compressing hydrogen sulfide can also be used at high rotation speeds.

The longitudinal bearing disk can be held radially in its position by the first, cylindrical contact face and therefore an unbalance can be avoided. The second, conical contact face serves to transmit the tangential forces from the longitudinal bearing disk to the shaft, so that said longitudinal bearing disk does not rotate relative to the shaft. In order to transmit the forces, the longitudinal bearing disk is expediently pressed against a mating face of the shaft or a component which is attached to said shaft by a shaft nut by way of its conical contact face.

The turbomachine is, in particular, a turbocompressor. The invention can be applied particularly advantageously to a longitudinal bearing disk of the kind which is part of a magnet bearing. On account of the magnet bearing, the longitudinal bearing disk is of relatively large design and is therefore subject to a high level of mechanical stress in the event of fast rotations. Particularly low-friction bearing can be achieved by magnetic bearing.

The conical contact face serves for automatic centering of the longitudinal bearing disk on the shaft. The conical contact face expediently does not directly adjoin the first, cylindrical contact face, in order to avoid a sharp edge which could damage the shaft when the longitudinal bearing disk is pushed onto the shaft. Therefore, the two contact faces are expediently spaced apart from one another, for example by a flat intermediate face, such as a chamfer, or can be connected to one another by a rounded portion.

In an advantageous embodiment of the invention, the second contact face is conical in an offset manner relative to radial compression of the longitudinal bearing disk. As a result, widening of the longitudinal bearing disk is avoided when the contact faces are compressed, and therefore the longitudinal bearing disk is mechanically protected. Compression counteracts the centrifugal forces and therefore has an advantageous effect on the strength of the longitudinal bearing disk.

In order to simultaneously transmit tangential forces, the conical contact face and its mating face can be designed as a Hirth connection. However, this refinement has the disadvantage that it is associated with high outlay on production. Therefore, conicity in the form of a truncated cone face is preferred, in the case of which force is transmitted from the contact face to the mating face by a force-fitting connection, that is to say by adhesion. In this case, the force with which the longitudinal bearing disk, at its conical contact face, is pressed against the mating face is selected to be so large that the resulting force on the contact face can transmit the required starting torques without the second contact face sliding on its mating face by friction moments.

The second contact face is advantageously designed with an angle of inclination relative to the radial direction of the shaft of between 5° and 30°. The greater the angle of inclination, the higher the friction moment between the contact face and mating face—given the same prestressing force—, however, also together with a higher mechanical stressing of the longitudinal bearing disk and the shaft. The forces which occur can be adapted, as is respectively required, by optimizing the angle of inclination.

The longitudinal bearing disk is expediently mounted such that an increasing bearing force acts on the second contact face as the rotation speed increases. This can be achieved, in particular, by the offset conicity of the second contact face. As the rotation speed increases, the increasing radial force leads to an increasing contact pressure by the second contact face against its mating face, as a result of which the torque which can be transmitted increases, said torque being generated by the friction action of the two contact faces on one another.

An increase in the temperature of the longitudinal bearing disk advantageously leads to an increasing bearing force of the second contact face. This can counteract a reduction in frictional forces on the contact faces.

It is also advantageous when the longitudinal bearing disk is mounted on the shaft without stress by way of the first contact face. The longitudinal bearing disk can be mechanically protected and therefore be driven at high rotation speeds. Freedom from stress is provided when the longitudinal bearing disk can be pushed onto the shaft by hand for the purpose of mounting the first contact face on its corresponding mating face.

Secure mounting of the longitudinal bearing disk and reliable operation of the rotor can be achieved when the second contact face rests against a conical mating face which is formed in the shaft. The conical mating face is therefore directly part of the shaft, and therefore a shaft ring or a similar component can be dispensed with.

The second contact face is advantageously an end face of the bearing disk, with an end face being understood to mean a face with an inclination of less than 45° relative to the radial direction of the shaft.

In order to avoid an unbalance of the longitudinal bearing disk after removal and re-mounting, a rotation-prevention means is expediently provided, said rotation-prevention means prespecifying a fixed tangential position for the longitudinal bearing disk on the shaft. The rotation-prevention means can also absorb tangential forces, which are produced, for example, when the rotor is started or by friction of the longitudinal bearing disk in the bearing, during operation. However, it is expedient for the rotation-prevention means to be designed to be as small as possible in order to not have too great an adverse effect on the stability of the longitudinal bearing disk. Therefore, it is advantageous when the second contact face is intended to transmit at least the major portion of the tangential forces from the shaft to the longitudinal bearing disk. The major portion is more than 50% of the tangential forces.

In a further advantageous embodiment of the invention, the rotor is equipped with a rotation-prevention means which holds the longitudinal bearing disk in a desired tangential position by a positive-locking connection. A change in the balance states after removal and re-mounting can be avoided.

The positive-locking connection can be established directly between the longitudinal bearing disk and the shaft, for example by a tongue-and-groove connection in or in the vicinity of the first contact face. However, the strength of the longitudinal bearing disk is not inconsiderably adversely affected by a recess in a face of the longitudinal bearing disk which faces radially inward. It is therefore advantageous when a positive-locking element of the rotation-prevention means engages in an end face of the longitudinal bearing disk. A recess in the end face leads to a considerably lower mechanical stress on the foot of the longitudinal bearing disk in the event of fast rotations. The end face is expediently arranged opposite the second contact face. The second contact face can be designed to be free of cavities and entirely for transmitting friction.

The rotation-prevention means advantageously comprises a rotation-prevention ring which at least indirectly forms a positive-locking connection with the longitudinal bearing disk and forms at least an indirect positive-locking connection with the shaft in the same direction. Positive transmission of force from the longitudinal bearing disk to the rotation-prevention ring can be continued in a simple manner in the same direction, for example in the tangential direction, as to the shaft. The production of the elements can be kept simple when the two positive-locking connections are each formed by a positive-locking element which engages in the elements which are connected in a positive-locking manner. The positive-locking connection between the rotation-prevention ring and the shaft is expediently achieved by a feather key which engages in a groove in the rotation-prevention ring and the shaft. For balance reasons, two positive-locking elements are expediently provided for each of the two positive-locking connections.

The feather key and the positive-locking element are expediently arranged so as to be tangentially offset to one another; when there are two positive-locking elements and two feather keys, which are advantageously mounted opposite one another in each case, positive-locking elements and feather keys are advantageously arranged at an angle of 90° in relation to one another.

A large increase in temperature may also cause the longitudinal bearing disk to expand to a particularly great extent in the axial direction. Under certain circumstances, the second contact face presses against its mating face with an undesirably great force, as a result of which high mechanical stress is produced. In order to reduce this stress, the rotor expediently comprises a spring means for absorbing bearing forces on the second contact face. The spring means can obtain its spring effect by virtue of a recess which is compressed in the event of a spring-like movement. The recess can be formed in the longitudinal bearing disk or in the component which forms the mating face to the conical contact face of the longitudinal bearing disk, in particular directly in the shaft.

DETAILED DESCRIPTION OF INVENTION

The invention will be explained in greater detail with reference to an exemplary embodiment which is illustrated in a drawing. The single figure in said drawing shows a sectional illustration of a detail of a rotor 2 of a turbomachine in the form of an axial compressor. The rotor 2 comprises a shaft 6 which runs in the axial direction 4 and to which a longitudinal bearing disk 8 is attached. The longitudinal bearing disk 8 is part of an axial bearing 10, two magnet poles 12 of said axial bearing, said magnet poles supporting the longitudinal bearing disk 8 in the axial direction 4, being schematically illustrated for better understanding. The axial bearing 10 is therefore a magnet bearing for holding the shaft 6 in an intended axial position.

The longitudinal bearing disk 8 is provided with a first, hollow-cylindrical contact face 14 which rests on a likewise cylindrical mating face 16 of the shaft 6. The bearing which is formed by the contact face 14 and the mating face 16 holds the longitudinal bearing disk on the shaft 6 without play in the radial direction 18. The longitudinal bearing disk 8 has a first end face 20 and a second end face which is opposite said first end face and is designed as a second and conical contact face 22. The contact face 22 rests on a mating face 24 which is formed directly in the material of the shaft 6.

The contact face 22 is provided with an angle of inclination α of 25°, with the contact face 22 being designed to be offset, that is to say such that the mating face 24 engages somewhat beyond the longitudinal bearing disk 8 in the region of the contact face 22. As a result, the longitudinal bearing disk 8 is compressed when the longitudinal bearing disk 8 is axially braced, and pressed onto the shaft 6 against a centrifugal force effect. The conicity of the contact face 22 has the effect of the longitudinal bearing disk 8 being automatically centered on the shaft 6, this assisting in centering the longitudinal bearing disk 8 by contact between the faces 14, 16.

In order to avoid an edge of the longitudinal bearing disk 8 butting against the shaft material, a chamfer 26 is formed between the two contact faces 14, 22, with a rounded portion between the contact faces 14, 22 also being expedient. A cavity 28 in the shaft material leads to a reduction in the stress peaks in the material of the shaft 6 in the event of a high contact-pressure force of the longitudinal bearing disk 8 against the shaft 6 in the axial direction 4. For the same purpose, cavities 30 are made in both sides of the longitudinal bearing disk 8, with the cavities 28, 30 advantageously being formed by calculation using the finite element method in order to subject the cavities 28, 30 to as uniform a material stress as possible along the wall.

A shaft nut 32, which is connected to the shaft 6 by a thread 34 in a positive-locking manner, is braced against a rotation-prevention ring 36 in order to mount the longitudinal bearing disk 8, said rotation-prevention ring pressing the longitudinal bearing disk against the corresponding mating face 24 by way of its contact face 22. A prestressing force FV which is applied by the shaft nut 32 is selected to be of such a magnitude that the normal force FN of the contact face 22 on the mating face 24, said normal force resulting from an axial force FAX and the angle of inclination α, produces such a great friction action that the longitudinal bearing disk 8 does not slip on the shaft 6 in the tangential direction when the rotor 2 is started.

As the rotation speed increases, the increasing radial force FR leads to an even greater normal force FN. As a result, the transmissible torque is automatically increased, this being generated by the friction action of the longitudinal bearing disk 8 on the mating face 24. An increase in temperature in the longitudinal bearing disk 8 also produces a higher normal force FN, and therefore a higher transmissible torque, on account of the expansion of the longitudinal bearing disk 8 in the axial direction 4. The resulting axial force FAX is absorbed by the shaft nut 32.

In order to avoid an excessively high normal force FN or axial force FAX in the event of a very high temperature of the longitudinal bearing disk 8, and therefore damage to the material of the longitudinal bearing disk 8 or shaft 6, a recess 38 is made in the shaft 8. This recess runs around the shaft 6 and thereby forms a slightly elastic web 40 which limits the contact-pressure force of the two contact faces 22, 24 on one another to a level which is safe for the materials.

Mounting or removal of the longitudinal bearing disk 8 can be performed in a simple manner by the longitudinal bearing disk 8 being pushed into the shaft 6 or being removed from said shaft by hand. In order to avoid an unbalance after removal, the longitudinal bearing disk 8 should be returned to its original tangential position on the shaft 6 when it is re-mounted. In order to ensure this, the rotation-prevention ring 36 is arranged between the shaft nut 32 and the longitudinal bearing disk 8, said rotation-prevention ring being connected to the longitudinal bearing disk 8 in a positive-locking manner by means of two positive-locking elements 42 in the form of bolts in the tangential direction, that is to say in the circumferential direction.

The positive-locking elements 42, of which only one is illustrated for the sake of clarity, the other being considered to be offset through 180°, that is to say in relation to the rotation axis 44 of the shaft 16, engage in the end face 20 of the longitudinal bearing disk 8, so that the contact face 14 of the longitudinal bearing disk 8 continues to not be adversely affected. Both the rotation-prevention ring 36 and the shaft 6 are provided with a groove 46 and, respectively, 48 which are held in position relative to one another in the tangential direction by means of a feather key 50 which is designed as a rectangular steel bar. In this way, an indirect positive-locking connection is established between the shaft 6 and the longitudinal bearing disk 8 in such a way that the longitudinal bearing disk 8 can be mounted only in a preset position in the tangential direction and remains in this position.

The feather key 50 is also provided in duplicate, with the two feather keys 50 likewise being arranged opposite one another in relation to the rotation axis 44. The feather keys 50 are arranged in a manner offset through 90° in relation to the positive-locking elements 42 in the tangential direction. Solely for the sake of simplifying the figures, said feather keys are shown in the same sectional plane, in order to be able to dispense with the need for a further sectional illustration through the rotor 2 rotated through 90° about the rotation axis 44.

The positive-locking element 42 and the feather key 50 can be designed as additional security means against tangential rotation of the longitudinal bearing disk 8 on the shaft 6. However, in order to keep the stability and thus the dimensions of these two elements low, the interaction between the prestress FV, contact face 22 and angle of inclination α is selected such that the longitudinal bearing disk 8 would not slip on the shaft 6 in the tangential direction even without these two elements. Said elements therefore serve mainly to aid mounting and to ensure absorption of tangential forces in undesirable or unforeseen states.

Claims

1-14. (canceled)

15. A rotor for a turbomachine, comprising:

a shaft, and

a longitudinal bearing disk arranged on said shaft as an element for an axial bearing for axially bearing the shaft, the longitudinal bearing disk comprising: a first, cylindrical contact face for radial mounting on the shaft, and a second, conical contact face for automatic centering on the shaft.

16. The rotor as claimed in claim 15, wherein the second contact face is conical in an offset manner relative to radial compression of the longitudinal bearing disk.

17. The rotor as claimed in claim 15, wherein the second bearing face has an angle of inclination relative to the radial direction of the shaft of between 5° and 30°.

18. The rotor as claimed in claim 15, wherein the longitudinal bearing disk is mounted such that an increasing bearing force acts on the second contact face as the rotation speed increases.

19. The rotor as claimed in claim 15, wherein the longitudinal bearing disk is mounted on the shaft without stress by way of the first contact face.

20. The rotor as claimed in claim 15, wherein the second contact face rests on a conical mating face which is formed in the shaft.

21. The rotor as claimed in claim 15, wherein the second contact face is an end face of the longitudinal bearing disk.

22. The rotor as claimed in claim 15, wherein the second contact face is provided for transferring at least the major portion of the tangential forces from the shaft to the longitudinal bearing disk.

23. The rotor as claimed in claim 15, further comprising a rotation-prevention device for holding the longitudinal bearing disk in a desired tangential position by a positive-locking connection.

24. The rotor as claimed in claim 23, wherein a positive-locking element of the rotation-prevention device engages in an end face of the longitudinal bearing disk.

25. The rotor as claimed in claim 24, wherein the end face is arranged opposite the second contact face.

26. The rotor as claimed in claim 23, wherein the rotation-prevention device has a rotation-prevention ring which at least indirectly forms a positive-locking connection with the longitudinal bearing disk along a direction, and forms at least an indirect positive-locking connection with the shaft in the same direction.

27. The rotor as claimed in claim 15, further comprising a spring element for absorbing bearing forces on the second contact face.

28. The rotor as claimed in claim 27, wherein the spring element obtains its spring action by a recess which is compressed in the event of a springing operation.