The invention relates to a vertical pump for conveying a process fluid in accordance with the preamble of the independent patent claim.

Vertical pumps are rotary pumps having a pump shaft extending in the direction of gravity, i.e. during operation the pump shaft rotates about a rotation axis, which is oriented vertically. Vertical pumps are used in many different industries, for example in the oil and gas processing industry, the power generation industry, the chemical industry, the clean and waste water industry. Vertical pumps may be designed as single stage pumps or as multistage pumps. Furthermore, vertical pumps can be configured as single phase pumps for conveying a single phase process fluid, such as water, or they can be configured as multiphase pumps for conveying a multiphase process fluid, which comprises a mixture of a plurality of phases, for example a liquid phase and a gaseous phase. An important example is the oil and gas processing industry, where vertical multiphase pumps are used for conveying hydrocarbon fluids, for example for extracting the crude oil from the oil field or for transportation of the oil/gas through pipelines or within refineries.

Fossil fuels are usually not present in pure form in oil fields or gas fields, but as a multiphase mixture which contains liquid components, gas components and possibly also solid components. This multiphase mixture of e.g. crude oil, natural gas, chemicals, seawater and sand has to be pumped from the oil field or gas field. For such a conveying of fossil fuels, vertical multiphase pumps are used which are able to pump a liquid-gas mixture which may also contain solid components, sand for example. Nowadays, vertical multiphase pumps are known that can pump multiphase fluids with strongly varying composition. For example, during exploitation of an oil field the ratio of the gaseous phase (e.g. natural gas) and the liquid phase (e.g. crude oil) is strongly varying. The ratio of the gaseous phase in the multiphase mixture is commonly measured by the dimensionless gas volume fraction (GVF) designating the volume ratio of the gas in the multiphase process fluid. In applications in the oil and gas industry vertical pumps should be able to convey multiphase process fluids having a gas volume fraction of 0% to 100%.

In view of an efficient exploitation of oil and gas fields there is nowadays an increasing demand for pumps that may be installed directly on the sea ground in particular down to a depth of 500 m, down to 1000 m or even down to more than 2000 m beneath the water's surface. Needless to say, that the design of such pumps is challenging, in particular because these pumps shall operate in a difficult subsea environment for a long time period with as little as possible maintenance and service work. This requires specific measures to minimize the amount of equipment involved and to optimize the reliability of the pump.

Vertical pumps are also used at topside locations on or above the water surface. For example, the pump may be arranged ashore or on an oil platform, in particular on an unmanned platform, or on a FPSO (Floating Production Storage and Offloading Unit). In particular regarding a deployment at locations where the available space is strongly restricted, e.g. on a platform or on a FPSO, vertical pumps have the advantage of a small foot print.

If the vertical pump is configured as a subsea pump, the drive unit for driving the pump is usually arranged in the pump housing, meaning that the drive unit and the pump unit are arranged in the same housing. For topside applications it is usually preferred to arrange the drive unit in a separate drive housing which is different from the housing in which the pump unit is arranged.

In terms of rotor dynamics vertical pumps have a considerable drawback as compared to horizontal pumps. Horizontal pumps have a pump shaft extending perpendicular to the vertical direction, namely in horizontal direction.

In horizontal pumps the radial bearings, also referred to as journal bearings, are loaded with a well-defined load, because the gravity pulls the pump shaft with the impellers mounted to it in one constant direction, namely the direction of gravity. Thanks to this static load vector, which consists of the weight of the pump shaft and the impellers, the pump shaft has a preferred position in the bearing clearance of the radial bearing(s) and will move on the fixed locus curve in function of the rotational speed and the load magnitude. The position of the shaft on the locus curve is determined by the Sommerfeld number.

For a vertical pump, there is no force or nearly no force in the bearing plane of the radial bearings, which pulls the shaft in one specific direction. There is nearly no load generated on the radial bearings by the weight of the pump shaft and the impeller(s) and there is no defined direction of the static load. This means that the pump shaft does not have a preferred position in the bearing clearance of the radial bearing(s). Therefore, a small excitation is enough to make the pump shaft move within the bearing clearance. Thus, in a tilting pad journal bearing a small excitation is enough to push the pump shaft from one pad to another pad.

Vibration measurements on operating vertical pumps show a particular shaft movement, where the shaft is observed to be walking from pad to pad. This walking is caused by the lack of a preferred position in the bearing clearance of the radial bearing(s) and it results in a high amplitude shaft movement. This high amplitude shaft movement can even exceed the limits given by industrial norms such as the API vibration criteria. Typically such vibrations are low energy-high amplitude vibrations because it does not require a large force to excite vibrations of large amplitudes.

Not only but particularly vertical multiphase pumps are especially prone to these vibrations for a variety of reasons. A usual single phase centrifugal pump has a significant amount of internal damping due to the single phase leakage over the internal seals or gaps along the rotor of the pump, such as the impeller eye seal, the hub seals, wear rings, throttle bushings and the balance drum. The leakage flow through those seals is counteracting vibrations and is generating damping. This physical phenomenon is called the Lomakin effect. A multiphase pump has by design much less seals as a single phase pump and the Lomakin effect of a multiphase leakage flow in a seal is also significantly smaller. Therefore, a multiphase pump has relatively low internal damping and relatively high hydraulic excitations due to the erratic character of the multiphase flow.

This makes it a challenge to design a vertical multiphase pump in such a manner that it meets the vibration acceptance criteria. To address this problem, EP 3 315 803 A2 proposes to use a magnetic force generated for example by a permanent magnet to pull the shaft of a vertical pump in the radial bearing in a defined direction. The magnet is arranged laterally with respect to the shaft, so that the magnetic force has a comparable effect as the gravity in a radial bearing of a horizontal pump. This would give the shaft thus a preferred position in the bearing clearance.

Starting from this prior art it is an object of the invention to propose another vertical pump, in which the amplitude of said vibrations is considerably reduced, the vibration energy is removed, and the pump shaft has a preferred position in the radial bearing during operation of the pump.

The subject matter of the invention satisfying this object is characterized by the features of the independent claim.

Thus, according to the invention, a vertical pump for conveying a process fluid is proposed, comprising a pump housing with a pump inlet and a pump outlet, a rotor arranged in the pump housing and configured for rotating about an axial direction, and at least one first radial bearing for supporting the rotor with respect to a radial direction perpendicular to the axial direction, wherein the rotor comprises a pump shaft extending in the direction of gravity during operation of the pump, and at least one impeller fixedly mounted on the pump shaft for conveying the process fluid from the pump inlet to the pump outlet, and wherein the first radial bearing is configured as a tilting pad journal bearing comprising a first support carrier and a plurality of pads arranged on the first support carrier for supporting the pump shaft by means of a lubricant, wherein the plurality of pads define a bearing clearance, in which the pump shaft can move with respect to the radial direction. The plurality of pads is configured such that the bearing clearance has a first maximum extension in a first direction perpendicular to the axial direction, and a second maximum extension in a second direction perpendicular to the axial direction, wherein the second direction is different from the first direction, and wherein the first maximum extension is different from the second maximum extension.

In classical radial bearings configured as a tilting pad journal bearing the bearing clearance, in which the pump shaft can move with respect to the radial direction, has a circular cross section perpendicular to the axial direction. Thus, in a vertical pump, the pump shaft will be in a centered position between the pads, because with respect to the radial direction there is no preferred position for the pump shaft, which is superior to other radial positions. This results in the effect that even very small excitations are enough to push the pump shaft out of the centered position and to cause a walking of the pump shaft from one pad to another pad. According to the invention, it is proposed to dispose of said circular symmetry of the bearing clearance. The pads of the first radial bearing are arranged such, that the bearing clearance has different maximum extensions in a first direction and in a second direction, wherein both directions are perpendicular to the axial direction. By this configuration at least one preferred position in the bearing clearance is created for the pump shaft, which is an eccentric position with respect to the bearing clearance. Regarding the stability of the position of the pump shaft said preferred position is superior to other positions of the pump shaft in the bearing housing. Creating the preferred position of the pump shaft by means of dispensing with the circular symmetry of the bearing clearance perpendicular to the axial direction results essentially in the same effect as gravity provides in a radial bearing in a horizontal pump, namely to push the pump shaft in a preferred position, in which the lubricant is squeezed between at least one of the pads and the pump shaft. This asymmetry of the bearing clearance with respect to the axial direction ensures that the pump shaft in a vertical pump is no longer centered in the bearing clearance of the first radial bearing. Said asymmetry makes sure that the pump shaft has a preferred position in the bearing clearance and will not walk from one of the first pads to another one of the first pads. In particular, the pump shaft has an eccentric position in the bearing clearance, meaning that the radial distance between the pump shaft and the pad(s) varies when viewed in the circumferential direction of the pump shaft. A small unbalance will result only in a small vibration along the preferred position and therefore will not cause a walking of the pump shaft from one first pad to another first pad.

Preferably, the plurality of pads of the first radial bearing is configured such that the second direction is perpendicular to the first direction, so that the bearing clearance has an oval shape. Thus, the first direction, the second direction and the axial direction can span a Cartesian coordinate system, with an x-axis and an y-axis each directed in the radial direction and a z-axis directed in the axial direction. Due to the oval shape of the bearing clearance, the pump shaft is in a more stable position when located close or at one of the two focuses on the main axis of the ellipse of the bearing clearance as compared to a position in the middle of the ellipse, i.e. at the intersection of the main axis and the minor axis of the ellipse. Thus, there is a preferred position for the pump shaft which is eccentric relative to the bearing clearance of the first radial bearing. In this configuration there are two preferred positions in the bearing clearance for the pump shaft, namely in each case one close or at each focus of the ellipse.

According to a preferred embodiment each pad of the first radial bearing is configured to extend obliquely relative to the pump shaft when viewed in the axial direction. By this measure it is easy to realize in particular an oval cross section of the bearing clearance in a section perpendicular to the axial direction.

Preferably, each pad extends from a first axial end in the axial direction to a second axial end, wherein a first radial distance between the pump shaft and the pad at the first axial end is different from a second radial distance between the pump shaft and the pad at the second axial end. Thus, each pad of the first radial bearing is arranged to be tilted relative to the pump shaft with respect to the axial direction, such that a wedge-like or tapering room is formed between the particular pad and the pump shaft.

Even more preferred, each pad is configured such, that a radial distance between the pump shaft and the pad changes linearly from the first radial distance to the second radial distance.

According to a preferred embodiment the first support carrier has a central axis, wherein the first support carrier is configured with the central axis including an inclination angle with the axial direction, which is different from zero. The first support carrier can be configured for example with an annular shape for surrounding the pump shaft. Then, first support carrier can be arranged such that it is tilted in its entirety relative to the pump shaft, so that the central axis, i.e. the cylinder axis of the annular first support carrier is inclined with respect to the axis of the pump shaft. It is one possibility to realize this configuration that the first support carrier has a radially inner surface supporting the pads, and a radially outer surface, wherein a shim element is provided at the radially outer surface, and wherein the shim element is configured to position the first support carrier in a tilted position, such that the central axis of the first support carrier is inclined with respect to the pump shaft.

Furthermore, it is preferred that the first radial bearing comprises a first side wall and a second side wall, wherein the first support carrier is arranged between the first side wall and the second side wall with respect to the axial direction.

In this configuration at least one lateral shim element can be arranged between one of the side walls and the first support carrier, wherein the lateral shim element is configured to position the first support carrier in a tilted position, such that the central axis of the first support carrier is inclined with respect to the pump shaft.

According to another embodiment the first support carrier has a radially inner surface supporting the pads, and a radially outer surface, wherein the radially inner surface is slanted relative to the radially outer surface when viewed in the axial direction, meaning that the radially inner surface of the first support carrier and the radially outer surface of the first support carrier are not arranged coaxially. This measure renders possible that the radially inner surface supporting the pads extends obliquely relative to the pump shaft, so that the pads arranged on the radially inner surface of the first support carrier are slanted relative to the pump shaft when viewed in the axial direction to form the bearing clearance having no circular symmetry.

The vertical pump may comprise additional radial bearings. In particular, the vertical pump can comprise at least one second radial bearing, which is configured to have a second bearing clearance in which the pump shaft can move with respect to the radial direction, wherein the second bearing clearance has a circular shape or a circular ring shape perpendicular to the axial direction. The second radial bearing may be configured for example as a tilting pad journal bearing as it is known in the art.

According to a preferred configuration the vertical pump has a plurality of impellers, wherein the plurality of impellers comprises at least a first stage impeller and a last stage impeller. Thus, the vertical pump is preferably designed as a multistage pump.

When the vertical pump is designed as a multistage pump, the vertical multistage pump may be configured with an in-line arrangement of all impellers, i.e. all impellers are arranged in series with the suction side facing in the same direction for each impeller. It is also possible to configure the vertical multistage pump with a back-to-back arrangement of the impellers, i.e. the plurality of impellers comprises a first set of impellers and a second set of impellers wherein the first set of impellers and the second set of impellers are arranged in a back-to- back arrangement, so that an axial thrust generated by the first set of impellers is directed opposite to an axial thrust generated by the second set of impellers.

For a back-to-back arrangement of the impellers it is preferred that the vertical pump comprises a center bush, which is fixedly connected to the pump shaft between the first set of impellers and the second set of impellers, wherein a balancing passage is provided between the center bush and a second stationary part configured to be stationary with respect to the pump housing.

In particular for applications in the oil and gas industry it is a preferred design, that the vertical pump is configured as a multiphase pump for conveying a multiphase process fluid having a gas volume fraction of 0% (pure liquid) to 100% (pure gas). The multiphase process fluid is for example a multiphase mixture of crude oil, natural gas, chemicals, seawater and sand.

In view of a preferred application the vertical pump may be configured as a subsea pump, and preferably be configured for installation on a sea ground.

Particularly in view of subsea applications it is preferred that the vertical pump further comprises a drive unit arranged in the pump housing and configured for driving the rotor. The pump housing with the drive unit inside may then be configured as a pressure housing, which is able to withstand the large hydrostatic pressure at a subsea location, e.g. on the sea ground.

Preferably, the drive unit comprises a drive shaft for driving the pump shaft of the rotor, and an electric motor for rotating the drive shaft about the axial direction, wherein a coupling is provided for coupling the drive shaft to the pump shaft.

Furthermore, it is preferred that the drive unit is arranged on top of the pump shaft.

Regarding topside applications it is preferred that the drive unit is arranged in a separate drive housing which is different from the pump housing.

Further advantageous measures and embodiments of the invention will become apparent from the dependent claims.

The invention will be explained in more detail hereinafter with reference to embodiments of the invention and with reference to the drawings. There are shown in a schematic representation: Fig. 1 : a schematic cross-sectional view of a first embodiment of a vertical pump according to the invention,

Fig. 2: a schematic cross-sectional view of a second embodiment of a vertical pump according to the invention, Fig. 3: a schematic representation of a first embodiment of a first radial bearing according to the invention in a cross-sectional view along the pump shaft,

Fig. 4: a schematic representation of the first embodiment of the first radial bearing in a cross-sectional view perpendicular to the pump shaft along the cutting line IV-IV in Fig. 3, Fig. 5: an even more schematic representation of the first embodiment in a cross- sectional view along the pump shaft (left) and perpendicular to the pump shaft (right), and

Fig. 6: a schematic representation of a second embodiment of a first radial bearing according to the invention in a cross-sectional view along the pump shaft. Fig. 1 shows a schematic cross-sectional view of an embodiment of a vertical pump according to the invention, which is designated in its entity with reference numeral 1 . As commonly understood in the art, a vertical pump has a pump shaft 5 which extends, during operation, in the vertical direction, i.e. the direction of gravity. The vertical pump 1 is designed as a rotary pump for conveying a process fluid. The vertical pump 1 has a pump housing 2, in which a rotor 3 is arranged. The rotor 3 is configured for rotating about an axial direction A. In the vertical pump 1 the axial direction A coincides with the direction of gravity. For rotating the rotor 3 a drive unit 4 is provided. In the embodiment shown in Fig. 1 the drive unit 4 is also arranged inside the pump housing 2. It goes without saying that in other embodiments of the vertical pump - see for example the second embodiment of the vertical pump according to the invention (Fig. 2) - the drive unit 4 is arranged outside the pump housing 2, e.g. in a separate motor housing 40 (Fig. 2).

In the embodiment shown in Fig. 1 both the rotor 3 and the drive unit 4 are arranged within the pump housing 2. The pump housing 2 is designed as a pressure housing, which is configured to withstand the pressure generated by the vertical pump 1 as well as the pressure exerted on the pump 1 by the environment. The pump housing 2 may comprise several housing parts, which are connected to each other to form the pump housing 2 surrounding the rotor 3 and the drive unit 4. It is also possible that a rotor housing and a separate motor housing and/or separate bearing housings are inserted in the pump housing 2. In the embodiment shown in Fig. 1 the pump housing 2 is configured as a hermetically sealed pressure housing preventing any leakage to the external environment.

In the following description of the first embodiment of the vertical pump 1 according to the invention reference is made by way of example to the important application that the vertical pump 1 is designed and adapted for being used as a subsea multiphase pump in the oil and gas industry. Thus, the vertical pump 1 is configured for conveying a multiphase process fluid, for example a process fluid comprising liquid and gaseous components. In particular, the vertical pump 1 is configured for installation on the sea ground, i.e. for use beneath the water- surface, in particular down to a depth of 500 m, down to 1000 m or even down to more than 2000 m beneath the water-surface of the sea. In such applications the multiphase process fluid is typically a mixture containing hydrocarbons that has to be pumped from an oilfield for example to a processing unit beneath or on the water-surface or ashore. The multiphase mixture constituting the multiphase process fluid to be conveyed can include a liquid phase, a gaseous phase and a solid phase, wherein the liquid phase can include crude oil, seawater and chemicals, the gas phase can include methane, natural gas or the like and the solid phase can include sand, sludge and smaller stones without the vertical pump 1 being damaged on the pumping of the multiphase mixture.

It has to be understood that the invention is not restricted to this specific example but is related to vertical pumps in general. The vertical pump 1 may also be configured for topside applications, e.g. for an installation ashore or on an oil platform, in particular on an unmanned platform. In addition, the vertical pump 1 according to the invention may also be used for applications outside the oil and gas industry.

The pump housing 2 of the vertical pump 1 comprises a pump inlet 21 , through which the multiphase process fluid enters the pump 1 , and a pump outlet 22 for discharging the process fluid with an increased pressure as compared to the pressure of the process fluid at the pump inlet 21 . Typically, the pump outlet 22 is connected to a pipe (not shown) for delivering the pressurized process fluid to another location. The pressure of the process fluid at the pump outlet 22 is referred to as ’high pressure’ whereas the pressure of the process fluid at the pump inlet 21 is referred to as ‘low pressure’. A typical value for the difference between the high pressure and the low pressure is for example 100 to 200 bar (10 - 20 MPa).

The rotor 3 of the vertical pump 1 comprises the pump shaft 5 extending from a drive end 51 to a non-drive end 52 of the pump shaft 5. The pump shaft 5 is configured for rotating about the axial direction A, which is defined by the longitudinal axis of the pump shaft 5. The rotor 3 further comprises a plurality of impellers with a first stage impeller 31 , a last stage impeller 32 and optionally a number of intermediate stage impellers 33. In the first embodiment the vertical pump 1 is an eight stage pump having the first stage impeller 31 , the last stage impeller 32 and six intermediate stage impellers 33, which are all arranged one after another on the pump shaft 5. Of course, the number of eight stages is only exemplary. In other embodiments the vertical pump 1 may comprise more than eight stages, e.g. ten or twelve stages, or less than eight stages for example four or two stages, or an odd number of stages, e.g. three stages.

The first stage impeller 31 is the first impeller when viewed in the direction of the streaming process fluid, i.e. the first stage impeller 31 is located next to the pump inlet 21 at the low pressure side. The last stage impeller 32 is the last impeller when viewed in the direction of the streaming process fluid, i.e. the last stage impeller 32 is located next to the pump outlet 22 at the high pressure side of the pump 1 .

Each impeller 31 , 32, 33 is fixedly mounted on the pump shaft 5 in a torque proof manner.

The plurality of impellers 31 , 32, 33 is arranged one after another and configured for increasing the pressure of the fluid from the low pressure to the high pressure.

The drive unit 4 is configured to exert a torque on the drive end 51 of the pump shaft 5 for driving the rotation of the pump shaft 5 and the impellers 31 , 32, 33 about the axial direction A.

A direction perpendicular to the axial direction A is referred to as radial direction. The term ‘axial’ or ‘axially’ is used with the common meaning ‘in axial direction’ or ‘with respect to the axial direction’. In an analogous manner the term ‘radial’ or ‘radially’ is used with the common meaning ‘in radial direction’ or ‘with respect to the radial direction’. Hereinafter relative terms regarding the location like “above” or “below” or “upper” or “lower” or “top” or “bottom” refer to the usual operating position of the pump 1 . Fig. 1 and Fig. 2 show the vertical pump 1 in the usual operating position.

Referring to this usual orientation during operation and as shown in Fig. 1 the drive unit 4 is located above the rotor 3. However, in other embodiments the rotor 3 may be located on top of the drive unit 4.

As can be seen in Fig. 1 the plurality of impellers 31 , 32, 33 comprises a first set of impellers 31 , 33 and a second set of impellers 32, 33, wherein the first set of impellers 31 , 33 and the second set of impellers 32, 33 are arranged in a back-to-back arrangement. The first set of impellers 31 , 33 comprises the first stage impeller 31 and the three intermediate impellers 33 of the next three stages and the second set of impellers 32, 33 comprises the last stage impeller 32 and the three intermediate impellers 33 of the three preceding stages. In other embodiments the first set of impellers may comprise a different number of impellers than the second set of impellers.

In a back-to-back arrangement, as it is shown in Fig. 1 , the first set of impellers 31 , 33 and the second set of impellers 32, 33 are arranged such that the axial thrust generated by the action of the rotating first set of impellers 31 , 33 is directed in the opposite direction as the axial thrust generated by the action of the rotating second set of impellers 32, 33. As indicated in Fig. 1 by the dashed arrows without reference numeral, the fluid enters the vertical pump 1 through the pump inlet 21 located at the lower end of the rotor 3, passes the stages one (first stage), two, three and four, is then guided through a crossover line 34 to the suction side of the fifth stage at the upper end of the rotor 3, passes the stages five, six, seven and eight (last stage), and is then discharged through the pump outlet 22, which is arranged between the upper end and the lower end of the rotor 3.

For many applications the back-to-back arrangement is preferred because the axial thrust acting on the pump shaft 5, which is generated by the first set of impellers 31 , 33 counteracts the axial thrust, which is generated by the second set of impellers 32, 33. Thus, said two axial thrusts compensate each other at least partially.

For further reducing the overall axial thrust acting on the pump shaft 5 the pump 1 may further comprise a balance drum 7 and/or a center bush 35.

For supporting the pump shaft 5 the pump 1 further comprises an upper pump bearing unit 10 and a lower pump bearing unit 20. The upper pump bearing unit 10 is arranged adjacent to the drive end 51 of the pump shaft 5 between the impellers 31 , 32, 33 of the rotor 3 on the one side and the drive unit 4 on the other side, more precisely, between the balance drum 7 and the drive unit 4. The lower pump bearing unit 20 is arranged between the first stage impeller 31 and the non-drive end 52 of the pump shaft 5 or at the non-drive end 52. The pump bearing units 10, 20 are configured to support the pump shaft 5 both in axial and radial direction. In the embodiment shown in Fig. 1 the upper pump bearing unit 10 comprises a second radial bearing 19 for supporting the pump shaft 5 with respect to the radial direction, and an axial bearing 15 for supporting the pump shaft 5 with respect to the axial direction A. Within the upper pump bearing unit 10 the second radial bearing 19 and the axial bearing 15 are arranged such that the second radial bearing 19 is closer to the drive unit 4 than the axial bearing 15, and the axial bearing 19 is arranged between the second radial bearing 19 and the balance drum 7. Of course, it is also possible, to exchange the position of the second radial bearing 19 and the axial bearing 15 within the upper pump bearing unit 10, i.e. to arrange the axial bearing 15 between the second radial bearing 19 and the drive unit 4.

The lower pump bearing unit 20 comprises a first radial bearing 11. In the lower pump bearing unit 20 the first radial bearing 11 is arranged adjacent to or at the non-drive end 52 of the pump shaft 5.

In the embodiment shown in Fig. 1 , the lower pump bearing unit 20 does not comprise an axial or thrust bearing. Of course, it is also possible that the lower pump bearing unit 20 comprises an axial bearing for the pump shaft 5. In embodiments, where the lower pump bearing unit 20 at the non-drive end 52 comprises an axial bearing, the upper pump bearing unit 10 at the drive end 51 may be configured without an axial bearing or with an axial bearing.

The first and second radial bearings 11 , 19 are supporting the pump shaft 5 with respect to the radial direction, and the axial bearing 15 is supporting the pump shaft 5 with respect to the axial direction A, i.e. the vertical direction.

A radial bearing, such as the first radial bearing 11 or the second radial bearing 19 is also referred to as a “journal bearing” and an axial bearing, such as the axial bearing 15, is also referred to as an “thrust bearing”.

Preferably all the radial bearings 11 , 19 as well as the axial bearing 15 are configured as hydrodynamic bearings, and even more preferred as tilting pad bearings 11 , 15, 19 respectively. Even more preferred, the first radial bearing 11 and the second radial bearing 19 are each configured as a tilting pad journal bearing. This will be explained in more detail hereinafter.

According to the invention the vertical pump 1 comprises at least one first radial bearing 11 for supporting the rotor 3 with respect to the radial direction, wherein the first radial bearing 11 is configured as a tilting pad journal bearing comprising a first support carrier 12 (Fig. 3) and a plurality of pads 13 arranged on the first support carrier 12 for supporting the pump shaft 5 by means of a lubricant, wherein the plurality of pads 13 define a bearing clearance 136, in which the pump shaft 5 can move with respect to the radial direction, wherein the plurality of pads 13 is configured such that the bearing clearance 136 has a first maximum extension E1

(Fig. 5) in a first direction x perpendicular to the axial direction A, and a second maximum extension E2 in a second direction y perpendicular to the axial direction A, wherein the second direction y is different from the first direction x, and wherein the first maximum extension E1 is different from the second maximum extension E2.

Preferably, the second direction y is perpendicular to the first direction x, so that the bearing clearance 136 has an oval shape as shown in Fig. 5 on the right. Thus, the first direction x, the second direction y and the axial direction A can span a Cartesian coordinate system, with an x-axis and a y-axis each directed in the radial direction and a z-axis directed in the axial direction A.

Within the scope of this application the term “first radial bearing” 11 designates a tilting pad journal bearing, which has said asymmetry in the bearing clearance 136, namely that the bearing clearance 136 has the first maximum extension E1 in the first direction x perpendicular to the axial direction A, and the second maximum extension E2 in the second direction y perpendicular to the axial direction A, wherein the second direction y is different from the first direction x, and wherein the first maximum extension E1 is different from the second maximum extension E2. Thus, the term “first radial bearing” 11 designates such a tilting pad journal bearing, in which the bearing clearance 136 has no circular symmetry regarding the axial direction A, i.e. the shape of the cross section of the bearing clearance 136 perpendicular to the axial direction A is different from a circle and from a circular ring, respectively. Preferably said shape is an ellipse.

Due to the non-circular and preferably oval shape of the bearing clearance 136, the pump shaft 5 is in a more stable position when located in an eccentric position in the bearing clearance 136, e.g. the pump shaft 5 is in a more stable position close or at one of the two focuses on the main axis of the ellipse of the bearing clearance 136 as compared to a position in the middle of the ellipse, i.e. at the intersection of the main axis and the minor axis of the ellipse. Thus, there is a preferred position for the pump shaft 5, which is eccentric relative to the bearing clearance 136 of the first radial bearing 11. Thus, the asymmetry of the bearing clearance 136 with respect to the axial direction A creates essentially the same result as the gravity of the rotor in a horizontal pump.

Of course, the vertical pump 1 according to the invention may comprise more than one first radial bearing 11.

Within the scope of this application the term “second radial bearing” 19 designates a radial bearing or a journal bearing having a bearing clearance, in which the pump shaft 5 can move with respect to the radial direction, wherein the said bearing clearance has a circular or a circular ring shape perpendicular to the axial direction. To differentiate said bearing clearance of the second radial bearing 19 from the bearing clearance 136 of the first radial bearing 11 the bearing clearance of the second radial bearing 19 is referred to as “second bearing clearance”. Thus, the second bearing clearance is symmetric with respect to the axial direction A, i.e. the second bearing clearance has a circular symmetry. The second radial bearing 19 may be configured for example as a tilting pad journal bearing as it is known in the art.

Of course, the vertical pump 1 according to the invention may comprise more than one second radial bearing 19. In particular and as it is shown in Fig. 1 , the vertical pump 1 may comprise both at least one first radial bearing 11 and at least one second radial bearing 19.

Preferably, as already said, the vertical pump 1 comprises at least one balancing device for at least partially balancing the axial thrust that is generated by the impellers 31 , 32, 33 during operation of the pump 1. The balancing device may comprise the balance drum 7 (also referred to as throttle bush) and/or a center bush 35 (Fig. 1). The first embodiment of the vertical pump 1 comprises the balance drum 7 and the center bush 35 for at least partially balancing the axial thrust that is generated by the impellers 31 , 32, 33.

The balance drum 7 is fixedly connected to the pump shaft 5 in a torque proof manner. The balance drum 7 is arranged above the impellers 31 , 32, 33 of the rotor 3, namely between the impellers 33 of the rotor 3 and the drive end 51 of the pump shaft 5. The balance drum 7 defines a front side 71 and a back side 72. The front side 71 is the side facing the impellers 31 , 32, 33 of the rotor 3. In the first embodiment the front side 71 is facing the intermediate stage impeller 33 of the fifth stage. The back side 72 is the side facing the axial bearing 15 and the drive unit 4. The balance drum 7 is surrounded by a stationary part 26, so that a relief passage 73 is formed between the radially outer surface of the balance drum 7 and the stationary part 26. The stationary part 26 is configured to be stationary with respect to the pump housing 2. The relief passage 73 forms an annular gap between the outer surface of the balance drum 7 and the stationary part 26 and extends from the front side 71 to the back side 72.

A balance line 9 is provided for recirculating the process fluid from the back side 72 of the balance drum 7 to the low pressure side at the pump inlet 21. In particular, the balance line 9 connects the back side 72 with the low pressure side of the pump 1 , where the low pressure, i.e. the pressure at the pump inlet 21 prevails. Thus, a part of the pressurized fluid passes from the front side 71 through the relief passage 73 to the back side 72, enters the balance line 9 and is recirculated to the low pressure side of the pump 1. The balance line 9 constitutes a flow connection between the back side 72 and the low pressure side at the pump inlet 21. The balance line 9 may be arranged - as shown in Fig. 1 - outside the pump housing 2. In other embodiments the balance line 9 may be designed as internal line completely extending within the pump housing 2.

Due to the balance line 9 the pressure prevailing at the back side 72 is essentially the same - apart from a minor pressure drop caused by the balance line 9 - as the low pressure prevailing at the pump inlet 21.

The axial surface of the balance drum 7 facing the front side 71 is exposed to an intermediate pressure between the low pressure and the high pressure. In the first embodiment shown in Fig. 1 said intermediate pressure is the suction pressure of the fifth stage prevailing at the outlet of the crossover line 34 during operation of the vertical pump 1. Of course, due to smaller pressure losses the pressure prevailing at the axial surface of the balance drum 7 facing the front side 71 may be somewhat smaller than said intermediate pressure. However, the considerably larger pressure drop takes place over the balance drum 7. At the back side 72 it is essentially the low pressure that prevails during operation of the vertical pump 1. Thus, the pressure drop over the balance drum 7 is essentially the difference between the intermediate pressure and the low pressure.

The pressure drop over the balance drum 7 results in a force that is directed upwardly in the axial direction A and therewith counteracts the downwardly directed axial thrust generated by the first set of impellers 31 , 33, namely the first stage impeller 31 and the intermediate impellers 33 of the second, third and fourth stage.

As a further balancing device for reducing the overall axial thrust acting on the pump shaft 5, the center bush 35 is arranged between the first set of impellers 31 , 33 and the second set of impellers 33, 32. The center bush 35 is fixedly connected to the pump shaft 5 in a torque proof manner and rotates with the pump shaft 5. The center bush 35 is arranged on the pump shaft 5 between the last stage impeller 32, which is the last impeller of the second set of impellers, and the intermediate impeller 33 of the fourth stage, which is the last impeller of the first set of impellers, when viewed in the direction of increasing pressure. The center bush 35 is surrounded by a second stationary part 36 being stationary with respect to the pump housing 2. An annular balancing passage 37 is formed between the outer surface of the center bush 35 and the second stationary part 36.

The function of the center bush 35 and the balancing passage 37 is in principle the same as the function of the balance drum 7 and the relief passage 73. At the axial surface of the center bush 35 facing the last stage impeller 32 the high pressure prevails, and at the other axial surface facing the intermediate impeller 33 of the fourth stage a lower pressure prevails, which is essentially the same as the intermediate pressure when neglecting the small pressure losses caused by the crossover line 34. Therefore, the fluid may pass from the last stage impeller 32 through the balancing passage 37 to the intermediate impeller 33 of the fourth stage.

The pressure drop over the center bush 35 essentially equals the difference between the high pressure and the intermediate pressure. Said pressure drop over the center bush results in a force that is directed downwardly in the axial direction A and therewith counteracts the upwardly directed axial thrust generated by the second set of impellers 33, 32, namely the intermediate impellers 33 of the fifth, sixth and seventh stage and the last stage impeller 32.

The drive unit 4 comprises an electric motor 41 and a drive shaft 42 extending in the axial direction A. For supporting the drive shaft 42 a first radial drive bearing 43, a second radial drive bearing 44 and an axial drive bearing 45 are provided, wherein the second radial drive bearing 44 and the axial drive bearing 45 are arranged above the electric motor 41 with respect to the axial direction A, and the first radial drive bearing 43 is arranged below the electric motor 41 . The electric motor 41 , which is arranged between the first and the second radial drive bearing 43, 44, is configured for rotating the drive shaft 42 about the axial direction A. The drive shaft 42 is connected to the drive end 51 of the pump shaft 5 by means of a coupling 8 for transferring a torque to the pump shaft 5.

The drive bearings 43, 44 and 45 are configured to support the drive shaft 42 both in radial direction and in the axial direction A. The first and the second radial drive bearing 43, 44 support the drive shaft 42 with respect to the radial direction, and the axial drive bearing 45 supports the drive shaft 42 with respect to the axial direction A. The second radial drive bearing 44 and the axial drive bearing 45 are arranged such that the second radial drive bearing 44 is arranged between the axial drive bearing 45 and the electric motor 41.

Of course, it is also possible, to exchange the position of the second radial drive bearing 44 and the axial drive bearing 45.

The second radial drive bearing 44 and the axial drive bearing 45 may be configured as separate bearings, but it is also possible that the second radial drive bearing 44 and the axial drive bearing 45 are configured as a single combined radial and axial bearing supporting the drive shaft 42 both in radial and in axial direction A. The first radial drive bearing 43 is arranged below the electric motor 41 and supports the drive shaft 42 in radial direction. In the embodiment shown in Fig. 1 , there is no axial bearing arranged below the electric motor 41 . Of course, it is also possible that an axial drive bearing for the drive shaft 42 is - alternatively or additionally - arranged below the electric motor 41 , i.e. between the electric motor 41 and the coupling 8.

The electric motor 41 of the drive unit 4 may be configured as a cable wound motor. In a cable wound motor the individual wires of the motor stator (not shown), which form the coils for generating the electromagnetic field(s) for driving the motor rotor (not shown), are each insulated, so that the motor stator may be flooded for example with a barrier fluid. Alternatively, the electric motor 41 may be configured as a canned motor. When the electric drive 41 is configured as a canned motor, the annular gap between the motor rotor and the motor stator of the electric motor 41 is radially outwardly delimited by a can (not shown) that seals the motor stator hermetically with respect to the motor rotor and the annular gap. Thus, any fluid flowing through the gap between the motor rotor and the motor stator cannot enter the motor stator. When the electric motor 41 is designed as a canned motor a dielectric cooling fluid may be circulated through the hermetically sealed motor stator for cooling the motor stator.

Preferably, the electric motor 41 is configured as a permanent magnet motor or as an induction motor. To supply the electric motor 41 with energy, a power penetrator (not shown) is provided at the common housing 2 for receiving a power cable (not shown) that supplies the electric motor 41 with power.

The electric motor 41 may be designed to operate with a variable frequency drive (VFD), in which the speed of the motor 41 , i.e. the frequency of the rotation, is adjustable by varying the frequency and/or the voltage supplied to the electric motor 41. However, it is also possible that the electric motor 41 is configured differently, for example as a single speed or single frequency drive.

The drive shaft 42 is connected to the drive end 51 of the pump shaft 5 by means of the coupling 8 for transferring a torque to the pump shaft 5. Preferably the coupling 8 is configured as a flexible coupling 8, which connects the drive shaft 42 to the pump shaft 5 in a torque proof manner but allows for a relative lateral (radial) and/or axial movement between the drive shaft 42 and the pump shaft 5. Thus, the flexible coupling 8 transfers the torque but no or nearly no lateral vibrations. Preferably, the flexible coupling 8 is configured as a mechanical coupling 8. In other embodiments the flexible coupling may be designed as a magnetic coupling, a hydrodynamic coupling or any other coupling that is suited to transfer a torque from the drive shaft 42 to the pump shaft 5.

The vertical pump 1 further comprises two sealing units 50 for sealing the pump shaft 5 against a leakage of the fluid along the pump shaft 5. By the sealing units 50 the process fluid is prevented from entering the drive unit 4 as well as the pump bearing units 10, 20. One of the sealing units 50 is arranged between the balance drum 7 and the upper pump bearing unit 10 and the other sealing unit 50 is arranged between the first stage impeller 31 and the lower pump bearing unit 20. Preferably each sealing unit 50 comprises a mechanical seal. Mechanical seals are well-known in the art in many different embodiments and therefore require no detailed explanation. In principle, a mechanical seal is a seal for a rotating shaft and comprises a rotor fixed to the pump shaft 5 and rotating with the pump shaft 5, as well as a stationary stator fixed with respect to the pump housing 2. During operation the rotor and the stator are sliding along each other - usually with a liquid there between - for providing a sealing action to prevent the fluid from escaping to the environment or entering the drive unit 4 of the pump 1 or the pump bearings.

For the lubrication and the cooling of the sealing units 50 and the pump bearing units 10, 20 as well as for the cooling of the drive unit 4 the lubricant is provided, which also acts as barrier fluid of a barrier fluid system. Barrier fluid systems as such are well-known in the art since many years and therefore do not require a detailed explanation. A barrier fluid system comprises a reservoir for a barrier fluid, i.e. the lubricant as well as a circuit through which the barrier fluid is moved.

Fig. 2 shows a schematic cross-sectional view of a second embodiment of a vertical pump 1 according to the invention.

In the following description of the second embodiment of the vertical pump 1 only the differences to the first embodiment are explained in more detail. The explanations with respect to the first embodiment and variants thereof are also valid in the same way or in analogously the same way for the second embodiment. Same reference numerals designate the same features that have been explained with reference to the first embodiment or functionally equivalent features.

Compared to the first embodiment, it is the main difference, that in the second embodiment of the vertical pump 1 the drive unit 4 is arranged outside the pump housing 2, e.g. in the separate motor housing 40. Thus, the drive unit 4 and the vertical pump 1 have separate housings, namely the motor housing 40 and the pump housing 2. Preferably, the motor housing 40 is arranged on top of the pump housing 2 as it is shown in Fig. 2. The second embodiment is particularly suited for topside applications. For example, the vertical pump may be arranged ashore or on an oil platform, in particular on an unmanned platform, or one a FPSO. In particular regarding a deployment at locations where the available space is strongly restricted, e.g. on a platform or on a FPSO the vertical pump 1 has the advantage of a small foot print.

In many configurations the pump bearing units 10, 20 are arranged in separate bearing housings, which are fixedly connected to the pump housing 2. For example, the lower pump bearing unit 20 may be arranged in a lower bearing housing which is fixedly mounted to the pump housing 2, and the upper bearing unit 10 may be arranged in an upper bearing housing which is fixedly mounted to the pump housing 2.

The drive unit 4 for driving the pump shaft 5 of the vertical pump 1 comprises the electric motor 41 , the drive shaft 42 extending in the axial direction A and the motor bearings 43, 44, 45 which are not shown in detail in Fig. 2. The electric motor 41 , which is arranged inside the motor housing 40 is configured for rotating the drive shaft 42 about the axial direction A. The drive shaft 42 has an end 421 , which is arranged outside the motor housing 40. The drive end 51 of the pump shaft 5 is arranged outside the pump housing 2. The end 421 of the drive shaft 42 is connected to the drive end 51 of the pump shaft 5 by means of the coupling 8 for transferring a torque to the pump shaft 5. Preferably the coupling 8 is configured as a flexible coupling 8, which connects the drive shaft 42 to the pump shaft 5 in a torque proof manner, but allows for a relative movement between the drive shaft 42 and the pump shaft 5, e.g. lateral movements. Thus, the flexible coupling 8 transfers the torque but no or nearly no lateral vibrations. The flexible coupling 8 may be configured as a mechanical coupling, a magnetic coupling, a hydrodynamic coupling or any other coupling that is suited to transfer a torque from the drive shaft 42 to the pump shaft 5.

Preferably, the electric motor 41 is configured as a direct drive motor 41 , meaning that there is no gear between the electric motor 41 and the pump shaft 5. The rotor of the electric motor 41 and the drive shaft 42 and the pump shaft 5 rotate all with the same rotational speed.

The electric motor 41 is configured to operate with a variable frequency drive (VFD), in which the speed of the drive, i.e. the frequency of the rotation is adjustable by varying the frequency and/or the voltage supplied to the electric motor 41 . Preferably, the electric motor 41 is designed as a high speed drive for generating the required rotational speed of the pump shaft 5. In the second embodiment of the vertical pump 1 the vertical pump 1 is designed with an inline arrangement of all impellers 31 , 32, 33. In an inline arrangement all impellers 31 , 32, 33 are configured such that the axial thrusts generated by the individual rotating impellers 31 , 32, 33 are all directed in the same direction, namely downwards in the axial direction A .

Now, the pump bearing units 10, 20 with the first radial bearing(s) 11 and/or the second radial bearing(s) 19 will be explained in more detail referring to several embodiments, in particular of the first radial bearing 11 . It has to be noted that the following explanation refers both to the first embodiment of the vertical pump 1 (Fig. 1) and to the second embodiment of the vertical pump 1 (Fig. 2), i.e. all the embodiments and variants explained hereinafter may be used for the first and the second embodiment of the vertical pump 1 .

Fig. 3 shows a schematic representation of a first embodiment of the first radial bearing 11 according to the invention in a cross-sectional view along the pump shaft 5. The first radial bearing 11 belongs to the the lower pump bearing unit 20 (Fig. 1). It goes without saying that also the upper pump bearing unit 10 may be designed to comprise a first radial bearing 11 in addition to or as an alternative for the second radial bearing 19.

In the embodiment shown in Fig. 1 the upper pump bearing unit 10 comprises the second radial bearing 19. As already mentioned, the second radial bearing 19 may be configured for example as a tilting pad journal bearing as it is known in the art having a second support carrier (not shown in detail) and a plurality of pads (not shown in detail) arranged to be supported by the second support carrier. The plurality of the pads are supporting the pump shaft 5 by means of the lubricant in a matter known as such. The pads of the second radial bearing 19 are arranged on the second support carrier such that the entirety of said pads surrounds the pump shaft 5. Each pad is arranged pivotably on the second support carrier. To this end the surface of the particular pad abutting against the second support carrier may be provided with a pivot pin (not shown) or any other pivot member to allow for a pivoting of the particular pad relative to the second support carrier. Since such classical tilting pad journal bearings are well known to the person skilled in the art, it is not necessary to explain the second radial bearing 19 in more detail.

For a better understanding Fig. 4 shows a schematic representation of the first embodiment of the radial bearing 11 in a cross-sectional view perpendicular to the pump shaft 5 along the cutting line IV-IV in Fig. 3. Fig. 5 shows an even more schematic representation of the first embodiment in a cross-sectional view along the pump shaft 5 on the left side of Fig. 5 and a representation of the bearing clearance 136 in a cross-sectional view perpendicular to the pump shaft 5 on the right side. Fig. 5 more clearly shows different extensions and distances. The first radial bearing 11 comprises the first support carrier 12, which is configured and arranged to surround the pump shaft 5. The first support carrier 11 has a radially inner surface 125 and a radially outer surface 126. The radially inner surface 125 of the first support carrier 12 faces the pump shaft 5 and is ring-shaped. The pads 13 are arranged on the radially inner surface 125 of the first support carrier 12 for supporting the pump shaft 5 with respect to the radial direction by means of the lubricant. In the first embodiment of the first radial bearing 11 there are provided five pads 13 arranged around the pump shaft 5 such that the entirety of the pads 13 completely surrounds the pump shaft 5. It has to be understood that the number of five pads 13 is only exemplary. In other embodiments of the first radial bearing 11 more or less than five pads 13 may be provided.

Each pad 13 has a back side 131 (Fig.4) facing the first support carrier 12 and a front side 132 facing the pump shaft 5. Each pad 13 comprises a pivot 135 (Fig. 4) arranged at the back side 131 of the pad 13 and supported by the first support carrier 12 such that each pad 13 can pivot relative to the first support carrier 12. In Fig. 3 the pivots 135 are not shown. Each pad 13 has a generally arcuate shape and may be configured as a spherical pad 13 with the back side 131 being spherical or as a cylindrical pad 13 with the back side 131 having the shape of a cylinder segment.

The entirety of the pads 13 surrounds the bearing clearance 136 of the first radial bearing 11 , in which the pump shaft 5 is supported with respect to the radial direction. The bearing clearance 136 is the room, in which the pump shaft 5 can move between the pads 13 in the radial direction. The bearing clearance 136 is filled with the lubricant for forming a lubricant film or a lubricant layer between the rotating pump shaft 5 and the non-rotating first pads 13 during operation of the vertical pump 1.

According to the first embodiment of the first radial bearing 11 the first support carrier 12 is ring-shaped or has a cylindrical shape, respectively, and arranged to surround the pump shaft 5. The first support carrier 12 is arranged in a tilted position relative to the pump shaft 5, such that each pad 13 is extending obliquely relative to the pump shaft 5 when viewed in the axial direction A as can be best seen in Fig. 3. The ring-shaped first support carrier 12 has a central axis C (Fig. 5) and is arranged in the tilted position such that the central axis C includes an inclination angle a with the axial direction A, wherein the inclination angle a is different from zero.

Due to the tilted position of the first support carrier 12 with respect to the pump shaft 5 each of the pads 13, which are arranged on the radially inner surface 125 of the first support carrier! 2, extends obliquely relative to the pump shaft 5, when viewed in the axial direction A. Thus, a wedge is formed between each pad 13 and the pump shaft 5, so that for each particular pad 13 the radial distance between the pump shaft 5 and the particular pad 13 increases or decreases when viewed in the axial direction A.

Regarding the axial direction A each pad 13 extends from a first axial end 138 in the axial direction A up to a second axial end 139, wherein a first radial distance R1 between the pump shaft 5 and the pad 13 at the first axial end 138 is different from a second radial distance R2 between the pump shaft 5 and the pad 13 at the second axial end 139.

The radial distance between the pump shaft 5 and the pad 13 changes linearly from the first radial distance R1 to the second radial distance R2.

By this arrangement of the first support carrier 12 and the pads 13, respectively, the usual circular symmetry of the bearing clearance about the axial direction as it is known from a conventional tilting pad journal bearing is deliberately avoided. As it can be best seen in Fig. 5, in the first embodiment of the first radial bearing 11 according to the invention the bearing clearance 136 has an oval shape in a cross section perpendicular to the axial direction A. The first maximum extension E1 of the bearing clearance 136, which is the length of the minor axis of the ellipse, is smaller than the second maximum extension E2 of the bearing clearance 136, which is the length of the main axis of the ellipse.

For comparison, the representation on the right of Fig. 5 also shows a circle B, which would delimit the bearing clearance radially outwardly i.e. the space in which the pump shaft 5 may move with respect to the radial direction, if the pads 13 were arranged parallel to or coaxially with the pump shaft 5. In this case the bearing clearance had a circular radially outer boundary or a circular ring shape in a cross section perpendicular to the axial direction as it is indicated by the circle B, meaning that the bearing clearance had a circular symmetry with respect to the axial direction. The diameter of the circle B is denoted with D. For example, the circle B in Fig. 5 may represent the second bearing clearance of the second radial bearing 19 in a cross sectional view perpendicular to the axial direction A, with the second bearing clearance having the outer diameter D.

Due to the oblique arrangement of the pads 13 with respect to the axial direction A, the space, within which the pump shaft 5 can radially move between the pads 13 becomes smaller with respect to the first direction x and remains - at least approximately - the same with respect to second direction y (as compared to a parallel arrangement of the pads relative to the pump shaft 5). This results in the oval shape of the bearing clearance 136 having the first maximal extension E1 in the first direction x, which is smaller than the second maximal extension E2 in the second direction y.

By this asymmetry of the bearing clearance 136 the pump shaft 5 has a preferred position in the bearing clearance 136, which is an eccentric position. For the avoidance of doubt, the centric position of the pump shaft 5 in the bearing clearance 136 is the position, where the center of the pump shaft 5 is located on the intersection of the main axis and the minor axis of the ellipse representing the bearing clearance 136 on the right side in Fig. 5.

Due to the asymmetry of the bearing clearance 136 the preferred position of the pump shaft 5 in the bearing clearance 136 is an eccentric position as it is shown on the right side of Fig. 5. The pump shaft 5 has its preferred position when the pump shaft 5 is located close to or at one of the two focuses on the main axis of the ellipse of the bearing clearance 136. This position is preferred, because the pump shaft 5 is in a more stable position as compared for example with the centric position.

The first support carrier 12, which is configured and arranged to surround the pump shaft 5, is fixedly connected to and supported by the pump housing 2 or any other part that is stationary with respect to the pump housing 2 and fixed relative to the pump housing 2. The first support carrier 12 is arranged with respect to the axial direction A between a first side wall 121 and a second side wall 122. Each side wall 121 , 122 is fixedly mounted to the pump housing 2 or to any other part, which is stationary with respect to the pump housing 2 and fixedly connected to the pump housing 2. Each side wall 121 , 122 has a generally annular shape and is arranged to surround the pump shaft 5. Furthermore, each side wall 121 , 122 extends in the radial direction and has an inner diameter which is configured such that the side walls 121 , 122 do not jut out beyond anyone of the first pads 13. The side walls 121 and 122 are arranged parallel to each other, wherein the distance between the first side wall 121 and the second side wall 122 with respect to the axial direction A is larger as the extension of the first support carrier 12 in the axial direction A, so that the first support carrier 12 can be arranged in a tilted position between the first side wall 121 and the second side wall 122 (Fig. 3). The first support carrier 12 is fixedly mounted in the tilted position between the first side wall 121 and the second side wall 122, so that the first support carrier 12 cannot move. Preferably the first side wall 121 and the second side wall 122 are configured in an identical manner.

For fixing the first support carrier 12 in the tilted position with respect to the pump shaft 5, so that the central axis C of the first support carrier 12 includes the inclination angle a with the axial direction A, a shim element 16 is provided at the radially outer surface 126 between the first support carrier 12 and the pump housing 2, or any other part that is provided for supporting the first support carrier 12. The shim element 16 is configured to position the first support carrier 12 in the tilted position, such that the central axis C of the first support carrier 12 is inclined with respect to the pump shaft 5. As shown in Fig. 3 the shim element 16 is configured in the form of a wedge. It has to be noted that the shim element 16 does not necessarily extend along the entire outer circumference of the first support carrier 12. Furthermore, there may be a plurality of shim elements 16 arranged between the first support carrier 12 and the pump housing 2. For example, the shim element 16 shown on the left side of the pump shaft 5 in Fig. 3 may be a different, i.e. a separate element than the shim element 16 shown in Fig. 3 on the right side of the pump shaft 5.

To further secure the tilted position of the first support carrier 12 it may be advantageous to arrange at least one lateral shim element 17 between one of the side walls 121 , 122 and the first support carrier 12, wherein the at least one lateral shim element 17 is configured to position the first support carrier 12 in a tilted position, such that the central axis C of the first support carrier 12 is inclined with respect to the pump shaft 5.

In the embodiment shown in Fig. 3 there are lateral shim elements 17 arranged both between the first support carrier 12 and the first side wall 121 and between the first support carrier 12 and the second side wall 122. It has to be noted that each lateral shim element 17 does not necessarily extend along the entire outer circumference of the first support carrier 12. Furthermore, there may be a plurality of lateral shim elements 17 arranged between the first support carrier 12 and the first side wall 121 and/or a plurality of lateral shim elements 17 arranged between the first support carrier 12 and the second side wall 122. For example, the lateral shim elements 17 shown on the left side of the pump shaft 5 in Fig. 3 may be different, i.e. separate elements than the lateral shim element 17 shown in Fig. 3 on the right side of the pump shaft 5.

In other embodiments all the lateral shim elements 17 and the shim element(s) 16 are integrally formed as a single piece, which is arranged between the first support carrier 12 and the pump housing 2 and the side wall(s) 121 , 122. In case the entirety of the shim elements 16 and the lateral shim elements 17 forms a ring-shaped structure for completely surrounding the pump shaft 5, said ring-shaped structure may be manufactured e.g. as two half shells or four quarter shells to ease the mounting.

In other embodiments there is only one shim element 16 or a plurality of shim elements 16 but no lateral shim element 17. The shape of the first support carrier 12 may be adapted in order to fix the first support carrier 12 between the first side wall 121 and the second side wall 122. In other embodiments there is only one lateral shim element 17 or a plurality of lateral shim elements 17 but no shim element 16. The shape of the first support carrier 12 may be adapted in order to fix the first support carrier 12 between the first side wall 121 and the second side wall 122. In such embodiments it is possible that lateral shim elements 17 are arranged only between the first side wall 121 and the first support carrier 12, or that lateral shim elements 17 are arranged only between the second side wall 122 and the first support carrier 12, or that lateral shim elements 17 are arranged both between the first side wall 121 and the first support carrier 12 and between the second side wall 122 and the first support carrier 12.

The shim elements 16 and the lateral shim elements 17 are preferably made from a metallic materiel, for example a metal sheet, wherein the metallic material is machined or cut to the appropriate shape, so that the first support carrier 12 is fixed in the tilted position relative to the pump shaft 5.

Fig. 6 shows a schematic representation of a second embodiment of a first radial bearing 11 according to the invention in a cross-sectional view along the pump shaft 5.

In the following description of the second embodiment of the first radial bearing 11 only the differences to the first embodiment are explained in more detail. The explanations with respect to the first embodiment of the first radial bearing 11 and variants thereof are also valid in the same way or in analogously the same way for the second embodiment. Same reference numerals designate the same features that have been explained with reference to the first embodiment of the first radial bearing 11 or functionally equivalent features.

The main difference between the second embodiment of the first radial bearing 11 and the first embodiments is that according to the second embodiment the first support carrier 12 is shaped in such a manner that on the one side the first support carrier 12 is fixed between the first and the second side wall 121 , 122 and on the other side the radially inner surface 125 of the first support carrier 12 is slanted relative to the pump shaft 5, when viewed in the axial direction A. To this end, the radially inner surface 125 of the first support carrier 12 is slanted relative to the radially outer surface 126 of the first support carrier 12 when vied in the axial direction A.

Both the radially inner surface 125 and the radially outer surface 126 of the first support carrier 12 are configured as a cylindrical surface. The cylinder axis C1 of the radially outer surface 126 extends in the axial direction A, wherein the cylinder axis C2 of the radially inner surface 125 is inclined relative to the axial direction A, so that the two cylinder axes C1 , C2 include the inclination angle a. The radially inner surface 125 and the radially outer surface 126 are not arranged coaxially.

The side walls 121 and 122 are arranged parallel to each other, wherein the distance between the first side wall 121 and the second side wall 122 with respect to the axial direction A is essentially the same as the extension of the first support carrier 12 in the axial direction A, so that the first support carrier 12 is fixed between the first side wall 121 and the second side wall 122.