The present invention relates to a bearing system for a rotor in rotating machines, as specified in the preamble of patent claim 1 .

In existing rotating machines, the rotor is supported both axially and radially. This may be done using bearings which are lubricated, magnetic, gas-dynamic etc. A common feature of all these bearings is that the length of the rotor shaft increases. Complex and costly support systems are also a necessity, except in the case of gas-dynamic bearings. Gas-dynamic bearings of the foil type do not require such support systems, but their bearing stiffness is presently far less than that required for rotating machines with high power or pressure.

Several bearings such as hydrodynamic, aerodynamic, hydrostatic or aerostatic bearings have been proposed and to some extent tested but have not achieved significant popularity. Typical for hydrodynamic or aerodynamic bearings, for example a foil bearing, is that the shaft rotation generates a lift which gives bearing stiffness. In hydrostatic bearings, external pressurisation is carried out using specially formed recesses in the bearings. Some examples of such bearings are disclosed by EP-A1 1607633 and WO-A1 97/13084. EP-A1 1607633 describes a vacuum pump having two screw rotors which are supported by a pair of bearings installed on a pair of shafts. The invention is distinguished by a pair of shaft seals not being in contact with the pair of shafts are mounted between the screw rotors and pair of bearings. WO-A1 97/13084 discloses an impeller pump, in which the sealing and bearing unit has two rotor-stator pairs of tapered sleeves having a spiral groove which conveys barrier-liquid from an entry-mouth in an entry-chamber to an exit-mouth in an exit-chamber, providing the unit with mechanical stability, thrust support capability, and resistance to vibrations in the unit. This technology provides solutions which are costly, complex and bulky.

Aerodynamic and aerostatic bearings are receiving increased focus in the industry because of the potential to reduce cost and less complicated machinery. Foil bearings and aerostatic/hybrid bearings are the most common concepts under development. Both these technologies have currently strict limitations on achievable stiffness and damping values.

The geometry of a bearing influences the static stiffness and the damping values. There are different types of bearings, depending on the geometry of the bearing clearance arranged between the stator and the rotor. The bearing clearance can have a parallel, convergent or divergent geometry. In the above mentioned documents, the bearing clearance is axially constant throughout the bearing, and the rotor and stator are coaxial or eccentrically arranged.

Another important aspect of a bearing system is cross coupling stiffness. The cross coupling stiffness depends on fluid friction between the stator and the rotor when the bearing is in operation, and the cross coupling stiffness generates instability in the bearing system which is undesirable. Aerodynamic and aerostatic bearings may be provided with different types of seals within the bearing point. The documents EP-A1 1607633 and WO-A1 97/13084 disclose bearing types which require separate seals. Recently developed technology involves the use of aerodynamic combined bearing and seal combination for turbo machinery based on a hole-pattern or honeycomb seal technology. An example of such a bearing and seal combination is disclosed by WO-A1 2009/099334.

Annular seals like sleeve seals and honeycomb seals have a considerable amount of direct stiffness caused by the fluid film between the stator and the rotor, however the static stiffness generated in a bearing and seal is highly influenced by the geometry of the seal annulus.

WO-A1 2009/099334 describes a bearing and seal combination having a sleeve seal with plain or textured static surface. The rotor and stator are coaxially arranged with a convergent geometry of the bearing clearance to achieve high direct static stiffness. The static stiffness provides that rotor weight is carried by the bearing and seal combination.

To compensate for the rotor weight, the rotor in WO-A1 2009/099334 is run eccentrically in relation to the stator. The heavier the rotor, the more eccentric the rotor axis is run. This leads to a decrease in clearance and less favorable dynamic characteristics. Furthermore, also the convergent geometry of the bearing clearance reduces the damping in comparison to a parallel geometry of the bearing clearance.

Moreover, experience has showed that the stiffness forces are, for high pressure machines, sufficient to move the shaft resonance frequency significantly or even cause rotor instability if dominant negative direct stiffness is produced by the seal. In high pressure machines a straight seal may become divergent due to internal pressure distribution forcing the seal or support to deform. The different types of bearings mentioned in the documents above all have in common that the rotor and stator are coaxial or at least arranged in parallel. This means that the existing technology provides solutions which are limited by the tolerable eccentricity.

The main object of the present invention is therefore to provide an improved bearing system for the rotor in rotating machines.

This object is achieved by the bearing system disclosed in independent claim 1 . Preferred embodiments of the invention will be understood from the dependent claims and the following description of preferred embodiments.

In particular, the present invention decreases the need for an eccentric mode of operation of the bearing system. Among the advantages of a bearing system according to the invention is that positive static stiffness can be achieved without running the bearing eccentrically and without sacrificing much of the damping.

The nominal clearance of the bearing may be reduced which leads to a decrease in bearing and seal leakage flow. Another advantage is that the dimension of the bearing components may be decreased, thereby extending the field of use of this type of bearing. Furthermore, the design may be compact, allowing the rotor to be made shorter and more rigid for enhanced rotor-dynamic performance, or alternatively shorter and thinner for weight reduction. Another advantage is that the sealing aspect has much less importance than before, and that the costs are cut substantially as a result of the reliability or even the practicability of using the rotating machine in a subsea environment. The actual working medium in the machine may also be used during operation of the system so as to further reduce the complexity compared with known solutions. It should be noted that the fluid creating the bearing properties is not limited to gas and liquid only but can also consist of a mixture of gas and liquids.

The present invention will now be discussed in more detail with the aid of preferred illustrative embodiments shown in the drawings, in which: Brief Description of Drawings

Fig. 1 shows a schematic sectional view of the basic structure of a rotor machine having a combined bearing and seal according to prior art; Fig. 2 shows a schematic sectional view of the basic structure of a bearing system according to one embodiment of the present invention having convergent clearance geometry.

Fig. 3 shows a schematic sectional view of the basic structure of a bearing system according to another embodiment of the present invention, showing means providing for angular flexibility in the axial direction.

Fig. 4 shows a schematic sectional view of the basic structure of a bearing system according to another embodiment of the present invention, having parallel clearance geometry Fig, 5A to 5C shows schematically different embodiments of a stator provided with among others guide blades in perspective and section view, respectively, and seen projected into a horizontal plane.

Detailed description of the invention

With reference to the figures, the present invention shall be explained in more detail in connection with rotating machines, for example such as a compressor for use in subsea environments and which has a motor-powered rotor. However, this must not be understood as meaning that the invention relates solely to the illustrated compressor, as it is of course suitable for other rotating machine types and environments of use. Other examples are rotating machines such as pumps, turbines and expanders. Furthermore, it should be noted that the figures only show details which are important for the understanding of the invention. Moreover, same features are indicated by the same reference number throughout the specification.

Fig. 1 shows a schematic sectional view of the basic structure of a rotor machine 100 having a combined bearing and seal according to prior art. The prior art combined bearing and seal has an approximately cylindrical stator 1 10, i.e., the stationary part surrounding the rotor 120, is formed with a bore 1 1 1 , whereby an annular clearance 1 12 is formed between these. The stator 1 10 thus constitutes a "bearing point" for the rotor 120. Furthermore, a pressure difference between the inlet and the outlet of the bearing, i.e., the pressure drop across the clearance 1 12, is used to obtain the required stiffness and damping. In Fig. 1 this is symbolised by means of P2 and P1 , that is to say the inlet pressure and the outlet pressure of the bore 1 1 1 .

The precondition for a successful result is that the annular clearance 1 12 has a geometric configuration that gives sufficient stiffness and damping in the relevant frequency ranges, as symbolised by K and C in Fig. 1 . The stiffness can be provided by allowing the annular clearance to converge towards the lower pressure, so that the inlet clearance is greater than the outlet clearance. Positive direct stiffness is thus obtained in the bearing. Positive direct damping may be provided by means of the characteristics of the surface 1 13 of the stator 1 10 facing the rotor 120, e.g., by means of for example a honeycomb structure or other type of roughness in the surface. The stator is, for example, mounted in a T-shaped groove (not shown) with loose fit in the rotor machine housing, or provided with a tilt pad bearing seal arrangement. The location of the pivot, the surface grooves, the length and width of the pad is determined by the bearing properties. Positive direct stiffness is a known concept in the field of rotor dynamics and entails the countering of radial motion of the rotor by the bearing, so that the same holds the rotor centred in the clearance for correct positioning in relation to the stator. Positive direct damping means that the rotor vibration is damped by the bearing. The main principle of the static stiffness ability according to this concept is that on the side of the seal which shows a reduction of the clearance, the relative amount of the pressure drop in axial direction is increasing towards the seal exit. This means that the average pressure increases on this side of the seal. On the opposite side of the seal, the effect reduces the average pressure. This change in pressure distribution with seal eccentricity is the major mechanism producing the required static stiffness for a convergent and/or a tilted seal.

Fig. 2 shows a schematic sectional view of the basic structure of a bearing system 1 for a rotor 2 in a rotating machine according to one embodiment of the present invention having a rotor 2 being provided with at least one bearing 3. The bearing point for the rotor 2 is formed of a cylindrical stator 4 located within a rotating machine housing 5 and surrounding the rotor 2. The stator 4 is formed with a bore 6. It should be noted that the illustrated bearing system is very exaggerated for explanatory purposes.

The rotor 2 has a rotor axis 7 and the stator 4 has a stator axis 8. In the figure the dotted line indicates the rotor axis T without angular deviation before tilting the rotor. The rotor axis 7 ^bls and stator axis 8 are in this figure arranged coaxially.

Fig, 2 shows an annular clearance 9 formed between the rotor 2 and the stator 4. The annular clearance may be varied both in the axial direction and in the transverse direction depending on circumstances, pressure difference between the inlet and outlet side of the bearing, provided that the required stiffness and damping of the bearing is achieved. The bearing 3 is adapted to have a fluid 10 flowing in an axial direction between the inlet and the outlet of the bearing in the annular clearance by means of a an axial pressure gradient. According to this embodiment of the invention the fluid 10 is a gas, however the fluid can alternatively be a liquid or a multiphase fluid. A fluid film is thereby formed which provides stiffness and damping according to the same principle as in a radial bearing having desired dynamic stiffness and damping. The pressure gradient across the clearance is used to obtain the required stiffness. In Fig. 2 this is symbolised by means of P2 indicating the inlet pressure and P1 indicating the outlet pressure of the bore 6. In this embodiment the pressure P2 is higher than pressure P1 .

According to the present invention, the rotor 2 and stator 4 are arranged such that there is an angular deviation a between the rotor axis 7 and the stator axis 8. Thus, the rotor axis is tilted an angle a in relation to the stator axis. It should be noted that in Fig. 2 the illustrated angular deviation is very exaggerated for explanatory purposes.

The introduction of a small angular deviation a between the stator 4 and the rotor 2, has the advantageous effect that the required positive static stiffness can be achieved without the need to run the rotor 4 eccentrically and additionally, calculations shows that there is only a limited effect on the damping in the bearing.

By introducing a small angular deviation a, the rotor 2 is tilted within the bearing 3.

The rotor tilt influences the pressure in the gas layer of the aerodynamic bearing.

In dependence of the rotor tilt the pressure increases or decreases in different parts of the gas layer within the bearing. The pressure differences gives rise to corresponding bearing forces acting upon the rotor axis.

The angular deviation a is achieved by tilting the rotor about a predetermined tilting point TP located on the stator axis. As shown in Fig. 2, a first end 2a of the rotor is thereby moved in a positive direction along a y-axis, and a second end 2b of the rotor is moved in a negative direction along a y-axis.

For example, in one mode of operation the first end or the rotor 2a, at the bearing inlet, is moved in a positive direction along the y-axis such that the clearance increases by 10% on the negative side of the y-axis. In case the tilting point TP is located in the center of the bearing, the second end of the rotor 2b, at the bearing outlet, is moved in the negative direction along the y-axis such that the clearance increases by 10% on the positive side of the y-axis. This means that the rotor axis and the stator axis intersect in a tilting point which is an angular deviation neutral point on the stator axis. The tilting point TP is located at any preferred point on the stator axis between the end points of the bearing.

In one variant of the invention, the tilting point TP is located outside an end point of the bearing, but in close vicinity of the end point of the bearing.

In one variant of the invention the tilting point TP, the angular deviation neutral point, is located in the centre point, in the middle, of the bearing. In another variant the tilting point TP, the angular deviation neutral point, is located at one endpoint of the bearing. Another advantageous effect of operating the bearing 3 with the rotor 2 tilted in relation to the stator axis 8, is that it allows for a reduction of the nominal clearances in the bearing 3, hence, a seal and bearing leakage flow is reduced.

In Fig. 2 the bearing 3 is provided with a convergent geometry of the bearing clearance 9. However, a rotor 2 having an angular deviation a, a rotor tilt, may be provided in any type of bearing clearance 9 geometry such as parallel, convergent and divergent geometry as well as other geometries whereby the beneficial advantage of increasing the positive static stiffness of the bearing is achieved. Fig. 4 shows a bearing system 1 ' according to another embodiment of the invention provided with a parallel geometry of the bearing clearance 9 where the rotor 2 is tilted in relation to the stator axis 8.

The axis of rotation of the tilted rotor, when the rotor axis 7 is angularly deviated to the stator axis 8, may be orientated in any plane that intersects a plane orientated in alignment with the stator axis 8.

Preferably, the angular deviation a of the rotor axis 7 is selected in dependence of the magnitude of external radial forces to be counteracted. Such external radial forces are for example rotor gravitation or any other radial forces acting on the rotor.

The radial component of the bearing force generated by the angular deviation a of the rotor axis within the bearing is used to counteract the external radial forces acting upon the rotor. Thus there is a resulting radial force corresponding to the sum of the radial component of the bearing force and the external radial forces.

Preferably the angular deviation a of the rotor axis 7 is arranged in a plane such that the radial component of the bearing force acting on the rotor axis resulting from the angular deviation a counteracts the external radial forces of different sources acting on the rotor 2.

The angular deviation a may be varied within the limitations of the cross section of the annular clearance 9 of the bearing 3. The limitation is typically corresponding to a minimum clearance D _min between the rotor and the stator which is at least larger than the sum of the maximum allowable vibration amplitude of the rotor and a suitable margin, i.e;

D _min ≥ maximum allowable rotor vibration amplitude + margin

By providing the rotor with an angular deviation a, a restoring moment M about the tilting point TP is produced in the rotor 2 in addition to the resulting radial force on the rotor 2. The restoring moment M about the tilting point TP in the rotor provides a self-aligning effect between stator 4 and rotor 2, which advantageously can operate as a self-aligning mechanism.

The bore 6 of the stator 4 may have a varying cross-sectional area. In one embodiment, the bore 6 has a gradually increasing cross-sectional area in the direction of higher pressure P2 within the rotating machine, as shown in Fig. 2. Alternatively, the bore 6 may have a gradually decreasing cross-sectional area in the direction of higher pressure P2 within the rotating machine. In another embodiment the bore 6 has a constant cross-sectional area in the direction of higher pressure P2 within the rotating machine, see for example Fig. 4.

With reference to Fig. 3, one embodiment of the bearing system is shown which is provided with means for allowing angular flexibility of the bearing in the axial direction. Such means are for example a radial tilt pad 13 which is provided between the stator 4 and the rotating machine housing 5. This has the advantageous effect that the bearing is rigid in the radial direction and angularly flexible in the axial direction to allow for the self-alignment of the bearing system in response to the restoring moment M.

With reference to Figure 4, the bearing system shows a bearing having parallel clearance 9 geometry. The rotor axis 7 intersect the stator axis in a tilting point TP, an angular deviation neutral point, located in between the centre of the bearing and the outlet of the bearing in direction of the fluid flow. In this embodiment, the rotor axis 7 is moved in the negative direction of the y-axis on the inlet side of the bearing.

Preferably, the bearing is combined with a seal 1 1 , typically a shaft seal, thereby forming a combined bearing and seal. Such a combined bearing and seal is showed in Fig. 3 and Fig. 4. The shaft seal 1 1 will naturally have an axial pressure gradient. This is advantageous since the fluid, being a liquid, gas or a multiphase fluid, generate stiffness and damping properties between the rotor and the seal surface.

Preferably the shaft seal 1 1 is attached to the stator 4 by a seal retainer 12. The seal retainer 12 is connected to the stator and preferably arranged such that the angular deviation a of the rotor axis 7 is adjustable. Preferably the angular deviation a of the rotor axis 7 can be measured and adjusted during installation or based on rotor response after testing.

In one embodiment of the invention, the seal retainer is arranged like a tilt pad bearing (not shown in the figures). Tilt pad bearings are known in the field of bearings and are therefore not further explained here. The seal retainer/tilt pads are located on the surface of the bore 6. However, such a seal retainer/tilt pads should be optimized so that the leakage between the stator and seal retainer/tilt pads is low. This is an advantageous arrangement since it allows the angular deviation of the rotor axis to be adjusted. In case the bearing is formed like an aerodynamic bearing the seal retainer can alternatively be arranged like a foil bearing (not shown in the figures). Foil bearings are also known in the field of bearings. This also provides the advantageous effect that the angular deviation of the rotor axis is adjustable. As shown in Fig. 1 , damping can be increased with the aid of an alternative configuration of the surface of the stator facing the rotor. In this case, the bore 6 is formed with rough surface structure like a hole pattern, honeycomb or similar. This has a dampening effect in that the rotation of the fluid decreases which improves the damping of the bearing. Fig 5A-5C shows a stator according to the invention which is provided with a shaft seal. Such a seal is for example shown in WO-A1 2009/099334. Now, reference is made to Fig. 5A to 5C for further explanation of advantageous

embodiments of the invention having alternatives to reduce gas rotation. The alternatives may be seen isolated or in combinations with one another, and in the drawings it is illustrated with reference indicators that "a" specifies one example of an axial rib, "b" specifies one example of axial brush, "c" specifies one example of inclined hole-pattern, and "d" specifies one example of a guide apparatus, where the latter may be guide blades. Even if Fig. 5A to 5C all show such guide blades, it should be understood that the guide blade may be separately used or in combination with one or more of the other variants. It shall particularly be noted that a separate honeycomb structure disclosed for example in WO-A1 2009/099334 may possibly be used along with the other mentioned above. The examples above and their respective uses will be explained below. Typical radial clearance is approximately 0,2-0,6 mm is utilized.

Longitudinal, axial ribs "a "means a raised portion in the surface which is situated in the longitudinal direction of the stator. The rib clearance is constant and corresponds to a minimum clearance along the entire axial length of the stator.

Brushes "b" correspond partly to the longitudinal, axial ribs mentioned above, but instead of being of "solid wood" these are made from brushes having minimal radial or inclined threads.

Configuration of the hole-pattern "c" is by means of small holes in the surface arranged in a pattern which is inclined to the axial direction, whereby the flow restriction of gas is directionally orientated. As an example, inclined is here meant relative to the axial direction of the stator, as can be seen in figure 5C.

In one variant of the invention the stator is formed in such a manner that the gas rotation in the clearance providing for sealing and bearing is minimal, to achieve sufficiently effective damping at low frequencies. For instance, this may be achieved having longitudinal, axial ribs, brushes or by forming a hole-pattern, whereby the gas flows more easily against the rotating direction of the shaft. Similarly, one or more segmented annuluses may be situated in the stator, with or without injection/extraction of gas to achieve the same effect.

These embodiments involve that the stator is formed in such a manner that the gas rotation within the sealing and bearing clearance is minimal, to achieve sufficiently effective damping at low frequencies. As mentioned above, this may be achieved generally and as examples having longitudinal ribs, brushes or by forming a hole-pattern, whereby the gas flows more easily to the rotating direction. Similarly, one or more segmented annuluses may be situated in the stator, with or without injection/extraction of gas to achieve the same effect.

The above-mentioned embodiments, variants and examples of the present invention may be freely combined within the limits of the following claims without departing from the invention.