GB2491825A - A permanent magnet combined journal and thrust bearing for a large shaft - Google Patents

Abstract

A pseudo-levitation permanent magnet bearing assembly 14, 15, with thrust and radial load capability, for a tubular rotating shaft 1 is provided. The capability of the bearing assembly 15 can be varied by changing the V angle of its construction. It runs without any lubrication and in a vacuum and this makes it suitable for flywheel and many other applications. Considerable use is made of prime numbers of opposing pairs of permanent magnets in rings, in prime numbers of said rings, to minimise "cogging" and maximise stability. It can be designed as a stand-alone bearing or in a bearing shaft assembly, incorporating a hollow rotating shaft 1 running on a fixed shaft 3 or in any of the usual bearing combinations. The preferred embodiment V shape of the opposing magnet faces and the magnetic repulsion forces resist adequately centrifugal forces due to rotational speed and keep magnets in loci. The bearing assembly incorporates solid lubricated bearing pads in thrust and radial positions which are normally disengaged and only come into action if magnets are in danger of rubbing.

Description

Page 1 Permanent Magnet combined Journal and Thrust bearingjor large shafts Discussion This design concept originated from the need to have a centralising shaft / bearing arrangement for a large flywheel application. The shaft bearing arrangement does not support the flywheel, merely acts as a stabilising unit and supports the hollow output shaft driving the hermetically sealed magnetic gearbox / coupling output shaft.

It can be demonstrated that permanent magnet bearing assemblies can be stabilised by mechanical means and this is termed pseudo-levitation.

The concept shown in the assembly drawing has incorporated within solid lubricated bearing pads which are not engaged unless there is a risk of magnets rubbing. Both axial and radial excessive thrusts are catered for. This is achieved by selecting appropriate dimensions and tolerances.

Instrumentation is also incorporated so that if bearing pads do engage, a shutdown sequence is immediately commenced for investigation as to cause. This instrumentation is in the form of proximity measurement and as a back-up, bearing pad temperature. Excessive movement or temperature rise may involve emergency shutdown using SOLAS techhology.

The magnets for radial thrust may be tapered as shown to assist with axial loading and centralisation or they could be parallel to the shaft and axial magnets take the entire axial load. However the preferred embodiment is the tapered arrays.

The magnets on each array are a prime number, with stationary and rotating members being equal. However in each pair of magnet arrays, each shall have a different prime number from each other.

Whilst the drawing shows only one pair each, on shaft top and bottom, there is a need for a prime number of pairs such as 2, 3, 5 or even 7 or more, each with their own and differerent combinations of prime numbers of magnet assemblies. This is for creating a stable environment with adequate load bearing capability. Each of these magnet ring assembly pairs will have a differing "start" point from the first and "master" pair, such that angularly it is 360 degrees divided by the prime number of pairs, This will minimise cogging and improve stability. The number of pairs of magnet assemblies needs to be matched by an equal number of holes and location for the long stud which binds the stationary and rotating assemblies together. The maximum prime number of pairs of magnet assemblies should be used for best results in terms of stability.

The design is modular and with the exception of differing prime numbers of pairs of magnets, components on top or bottom assemblies are identical.

This shaft and components can be assembled in factory for shipment to site and is designed that if necessary it can be unbolted from the foundation and raised through the opening of the gearbox / coupling.

Both stationary and rotating shafts are one piece in the preferred embodiment, but multiple section components can be used to complete the assembly.

For stifihess the central section of stationary shaft is of larger diameter than the ends.

This assists with raising stationary shaft natural frequency Jevels above maximum running speed frequency. It can be a maximum of the ID of the rotating shaft less running clearances.

Page 2 The rotating shaft can be made from large diameter thick wall tubing, pipe or centrifugally cast steel or iron section. The first natural frequency of this shaft must be above maximum running speed frequency.

The key success factor in this design is to engage all the magnetic restoration forces with the spring forces on the drive spokes and the mechanical out of balance restoration ball assembly to create stable running conditions.

For magnetic restoration forces to be adequate the clearances between the magnets cannot be small, for example 1mm or less, as the net repulsive forces are a combination of increased force where clearance is getting smaller and lesser force where clearance is getting larger. At small air gaps since the repulsive force to air gap is a reverse exponential function and the mass of the flywheel is so large then larger air gaps of circa 4 -5mm arc more likely for stable operation.

This design caters for the larger shaft outputs associated with much larger MW output for a shorter period of time, on flywheel energy storage devices.

For smaller outputs ceramic bearings should provide a stable environment.

Page 3

Description

See Drawing 1 -Assembly Drawing.

The tubular rotating shaft is represented by #1. At top of this shaft is the hermetically sealed magnetic coupling / gearbox, #9, driving the output shaft, #11, to which the electrical power system is connected, either variable speed for synchronous generation or direct drive for power electronics conversion.

The rotating shaft, #1, rotates around the fixed shaft, #3, by means of magnetic bearing assemblies, #14 and #15. There are two types of bearing illustrated, #15 has angularly offset juxtaposed rings of magnets, which will take both radial and axial loads. Alternatively #14 has as shown, two juxtaposed rings of magnets, parallel to the axis of rotation, which only provides radial load capability and must be augmented by #13, in multiple locations for axial load capability.

The main shaft assembly without the coupling / gearbox component, #9, which is a separate component, can be assembled on site but preferably in a factory environment and shipped to site. The complete assembly can be inserted or withdrawn through the opening in the lid #10. The bottom of the shaft is fastened to the base flange #5, which is bolted to the foundation via bolt holes #21. The containment / vacuum vessel top, #10, makes a solid fixing for the shaft #3 at the top, so that top and bottom are fixed. The fasteners at #22 and or offset spigot and recess gives some adjustment to top shaft fixing to locate precisely the centre of rotation in a vertical position, or slightly offset to compensate for precession angle, relative to the earth's rotation.

Alternatively the bottom fixing of #3 to #5, by the addition of an intermediate flange with top fasteners, can be made adjustable, making #5 separate from the assembly.

The drive spokes are fastened to the flanges #15 on the rotating shaft #1.

The bearings are assembled as follows: Two steel drilled and tapped flanges, #4, are shrink fined to fixed shaft #3.

Two rings with compression double tapered split ring, #23, are fitted in their machined bore locations inside rotating shaft #1, by the recessed bolts in the tapered assembly and locked in position.

Fixed shaft, #3, is placed inside rotating shaft #1.

Assembly tooling #25 is threaded into D&T holes in flange #4.

On fixed shaft #3, flange #24, with minimal shaft clearance is fitted abutting flange #4, feeding it over the assembly tooling #25.

Assembly tooling #28 is fined to #23, tapered clamp flange D & T holes.

Flange #27, with minimal rotating shaft #1, ID clearance, is threaded onto tooling #28, until it abuts locating or tapered clamp flange ring #23.

Magnet assembly # 15, also with minimal shaft ID clearance is then threaded onto tooling #28 until it abuts #27.

Fixed double magnet assembly, #20, is threaded on to tooling #25, having also minimal shaft OD clearance, until it abuts flange #24.

Magnet assembly half, #16, with minimal ID clearance, is now fed on to tooling #28.

Final rotating flange, #29, is now fitted to tooling #28. This assembly is now subjected to the repulsive forces axially and needs clamping together.

Final clamping flange, #7, is now fitted to tooling #28.

Fixed shaft flange, #30, is now fitted to tooling #25.

Final clamping flange #6 is now fined to tooling #25.

Page 4 Tooling #28 and #25 are long threaded on one end and nuts are tightened in a manner to thaw the total assembly together uniformly.

On torquing up the tooling on both fixed and rotating assemblies is replaced, one at a time, with correct long, short threaded on each end, rods, again to specific torque levels,#17& #18.

Assembly of the parallel with axis magnet assemblies is somewhat similar in sequence and methodology.

Axial thrust magnetic bearing is located at #31.

Section #19 is shown to illustrate where the changes in overall length are accommodated.

Solid lubricated pads to stop excessive movement could be located in the area #12, both radial and axial as well as the instrumentation probes.

Some of these components will be steel, but ones containing magnets may be of insulating material at least partially to eliminate and or minimise eddy currents.

Drawing 2 illustrates that each magnet, #58, has to be inserted into a tray, #52, made of electrical laminated steel or other magnetic material, such as iron powder or soft magnetic manganese zinc or nickel zinc powder, resin and fibre reinforcement in a cast pocket or from an open top box made of laminated magnetic slot wedge material.

This simply provides the flux path, #53 a return route so that magnet operates at optimum performance. The size and cross section of #52 will have to be below a chosen flux density for this purpose.

Claims (19)

1. Claims 1. A pseudo-levitation, using multiple small permanent magnets, each set in a magnetic pocket, making a combined axial and radial thrust magnetic bearing utilising repulsion, for a vertical or horizontal axis shaft, wherein the outer rotating shaft is hollow running centralised on a fixed solid or hollow shaft, supported at both ends to a fixed base or alternatively the fixed one piece shaft may be interchanged to two stub shafts mounted in cantilever fashion to fixed ends, as the radial effect of centrifugal forces on the magnets is alleviated by the magnetic repulsive forces in the preferred embodiment of V formations.
2. 2. As in Claim I the preferred embodiment for the magnet installation is in a V formation at right angles to axis of rotation and this provides both axial and radial constraints.
3. 3. As per Claim 2, the angle of the V formation is a function of the combined axial and radial thrust design needs. A small included angle would have large axial properties and small radial properties, while a large included angle would be the reverse in properties.
4. 4. As in Claim 2, the magnet assemblies on one side of the V formation may be more than one long rectangular magnet and could be formed from multiple concentric rows of opposing magnets, with differing prime numbers of magnets in each opposing pair to assist stabilisation.
5. 5. As in Claim 4, the multiple rows of opposing concentric magnets should be a prime number.
6. 6. As in Claim 5, each magnetic opposing ring pairs shall have a differing "start" position as determined by 360 degrees divided by the prime number of magriet rings in the design.
7. 7. As in Claim 6, there will be an optimum solution as to number of rows of opposing prime numbers of magnets giving more stability as the higher prime number of pairs of V magnetic ring assemblies reduces the "cogging" effect.
8. 8. As in Claim 2, additional axial thrust can be provided by axially repelling magnet assemblies at both ends of the bearing assembly.
9. 9. As in Claim 2, the magnets for radial loading may be parallel to axis of rotation with constraint in the axis performed by axially repelling magnet assemblies as in Claim 7.
10. 10. As in Claim 8, a higher prime number of concentric rows of axial magnetic ring assemblies reduces the "cogging" effect.
11. 11. As in Claim 1, in the event of excessive movement, the bearing incorporates solid lubricated bearing pads to prevent magnets actually touching by selection of design running clearances.
12. 12. As in Claim 11, a proximity probe is incorporated to detect such an event and to shut the machine down for investigation.
13. 13. As per Claim 12, for back-up, the static bearing pads will incorporate a temperature measuring device to also shut the machine down for investigation.
14. 14. As per Claim 12, continuous vibration monitoring will take place to assist in analysing and predicting bearing behaviour.
15. 15. As per Claim 1, the magnets could be round to reduce cogging.
16. 16. As in Claim 1, the magnet assembly carrier can be wholly or partially manufactured from a non conducting material around the magnet / tray assemblies to eliminate eddy currents and improve efficiency.
17. 17. As in Claim 16, these trays can be made from an electrical steel laminated assembly, a moulded / resin component with high iron powder content or from other soft magnetic powders such as manganese zinc ferrite or nickel zinc ferrite or alternatively laminated magnetic slot wedge material.
18. 18. As per Claim 1, a bearing design which does not need lubrication and can run in a vacuum.
19. 19. As per Claim 1, a bearing which could be designed in a wide format for greater load bearing and stability properties.