Vertical Pump

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 measurements 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 one 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 rotor, which comprises 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 static force or nearly no static 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, for example a small dynamic force, 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 and thus there is not much energy within the movement.

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 first pads arranged on the first support carrier for supporting the pump shaft by means of a lubricant. Each first radial bearing comprises at least one spring element, which is arranged to act in a load direction on the first support carrier or on one of the first pads for placing the pump shaft relative to the first radial bearing into a preferred position, in which the lubricant is squeezed between at least one of the first pads and the pump shaft.

Thus, in a vertical pump a spring force or a spring load is used for the first radial bearing to generate essentially the same effect as the gravity does 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 first pads and the pump shaft. This loading of the first radial bearing by the at least one spring element in a constant and well-defined direction, namely the load direction, ensures that the pump shaft in a vertical pump is no longer centered in the bearing clearance of the first radial bearing. The loading of the first radial bearing in the constant load direction makes sure that the 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.

Consequently the pump shaft will be moving on a locus curve in function of the Sommerfeld number like it is the case for a radial bearing where the loading is due to the gravity, e.g. in a horizontal pump. 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.

According to a preferred embodiment at least one of the first pads is configured as a spring- loaded pad for placing the pump shaft into the preferred position, wherein the spring element is arranged between the first support carrier and the spring-loaded pad for pushing the spring- loaded pad in the load direction. In this embodiment one or more of the first pads are pushed by a particular spring element in the load direction so that the pump shaft is placed in the preferred position, in which the pump shaft has an eccentric position in the bearing clearance, meaning that the radial distance between the pump shaft and the first pad(s) varies when viewed in the circumferential direction of the pump shaft. According to a further preferred embodiment, at least one of the spring-loaded pads may be provided with a pressurized fluid, for example the lubricant, which is fed in between the spring-loaded pad and the pump shaft. To this end the at least one spring-loaded pad comprises a front side facing the pump shaft, a back side facing the first support carrier and a channel for a fluid communication between the front side and the back side, wherein the channel is arranged to receive a fluid for supplying the fluid to the front side of the spring- loaded pad.

Regarding the embodiments with the at least one spring-loaded pad, it is preferred that at least one of the first pads is configured as an unloaded pad which is free of a spring load. Configuring at least one of the first pads as an unloaded pad in combination with the spring- loaded pad(s) is a simple way for loading the first radial bearing in the constant load direction.

As a further preferred measure the first radial bearing comprises a plurality of spring-loaded pads for placing the pump shaft into the preferred position, and a plurality of unloaded pads, wherein all spring-loaded pads are arranged side by side when viewed in a circumferential direction of the pump shaft. By this measure, the pump shaft has a loaded side, namely this part of the radially outer surface of the pump shaft, which faces the spring-loaded pad(s), and an unloaded side, namely this part of the radially outer surface of the pump shaft, which faces the unloaded pads.

According to another preferred embodiment, the first support carrier is arranged to be movable in the radial direction, wherein the at least one spring element is arranged to act on a loaded side of the first support carrier for pushing the first support carrier in the load direction in order to place the pump shaft relative to the first radial bearing into the preferred position, in which the lubricant is squeezed between at least one of the first pads and the pump shaft. According to this embodiment the at least one spring element is arranged to push the entire first support carrier in a radial direction, thus creating a loaded side of the pump shaft so that the pump shaft is placed in an eccentric position with respect to the bearing clearance.

As a variant of this embodiment, the first support carrier has an unloaded side arranged opposite to the loaded side, wherein the unloaded side is free of first pads between the pump shaft and the first support carrier. By this measure the number of first pads may be reduced because there are no first pads on the unloaded side of the pump shaft.

According to still another preferred embodiment, the first support carrier is arranged to be tiltable relative to the pump shaft, and wherein the at least one spring element is arranged to act on the first support carrier in the axial direction for tilting the first support carrier relative to the pump shaft, such that each first pad is extending obliquely relative to the pump shaft when viewed in the axial direction. In this embodiment a kind of wedge is built between each first pad and the pump shaft, meaning that the radial distance between the pump shaft and a particular first pad increases when viewed in the axial direction.

As a preferred measure the vertical pump further comprises at least one second radial bearing for supporting the rotor with respect to the radial direction, wherein the second radial bearing is configured as a tilting pad journal bearing comprising a second support carrier and a plurality of second pads arranged on the second support carrier for supporting the pump shaft by means of the lubricant, and wherein each second pad is configured as an unloaded pad which is free of a spring load. The second radial bearing may be configured for example as a tilting pad journal bearing as it is known in the art.

Preferably the vertical pump comprises one first radial bearing and one second radial bearing, wherein the first radial bearing and the second radial bearing are arranged adjacent to each other regarding the axial direction. By arranging the first radial bearing and the second radial bearing next to each other it becomes possible that both radial bearings are loaded by the configuration of the first radial bearing. In particular, when the first support carrier is arranged to be movable in the radial direction the eccentricity of the pump shaft with respect to the bearing clearance of the first radial bearing will result in the loading of both the first and the second radial bearing. Both radial bearings will be loaded with essentially the same load but in opposite directions.

According to another preferred measure the vertical pump comprises two first radial bearings arranged adjacent to each other regarding the axial direction, wherein the load directions of the two first radial bearings are opposite to each other.

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 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 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 an arrangement of a first embodiment of a first radial bearing according to the invention and an embodiment of a second radial bearing 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: a schematic representation of an arrangement of two first radial bearing according to the invention in a cross-sectional view along to the pump shaft,

Fig. 6: a variant of the first embodiment of the first radial bearing according to the invention in a representation as in Fig. 5,

Fig. 7: a schematic representation of the variant shown in Fig. 6 in a cross-sectional view perpendicular to the pump shaft along the cutting line VII-VII in Fig. 6,

Fig. 8: 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. 9: a schematic representation of the second embodiment of the first radial bearing in a cross-sectional view perpendicular to the pump shaft along the cutting line IX-IX in Fig. 8,

Fig. 10: a schematic representation of a third embodiment of a first radial bearing according to the invention in a cross-sectional view along the pump shaft, Fig. 11 : a schematic representation of a fourth embodiment of a first radial bearing according to the invention in a cross-sectional view perpendicular to the pump shaft,

Fig. 12: a schematic representation of an arrangement of a fifth embodiment of a first radial bearing according to the invention and the second radial bearing in a cross-sectional view along the pump shaft,

Fig. 13: a schematic representation of the fifth embodiment of the first radial bearing in a cross-sectional view perpendicular to the pump shaft along the cutting line XIII-XIII in Fig. 12, and Fig. 14: a plan view of the front side of the spring-loaded pad of the fifth embodiment of the first radial bearing.

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 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 and also a second radial bearing 19. In the lower pump bearing unit 20 the first radial bearing 11 and the second radial bearing 19 are arranged adjacent to each other regarding the axial direction A. The first radial bearing 11 is arranged closer to the non-drive end 52 of the pump shaft 5 than the second radial bearing 19. Of course, it is also possible to exchange the position of the first radial bearing 11 and the second radial bearing 19 in the lower pump bearing unit 20, so that the second radial bearing 19 is closer to the non-drive end 52 of the pump shaft 5 than the first radial bearing 11 . 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 bearings 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 and 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.

In other embodiments the upper pump bearing unit 10 comprises a first radial bearing 11 and a second radial bearing 19, whereupon the lower pump bearing unit 20 comprises a second radial bearing 19 but no first radial bearing 11 .

In still other embodiments the upper pump bearing unit 10 comprises a first radial bearing 11 and a second radial bearing 19, and the lower pump bearing unit 20 also comprises a first radial bearing 11 and a second radial bearing 19.

When the lower pump bearing unit 20 comprises a first radial bearing 11 , the lower end of the pump shaft 5, here the non-drive end 52, is stabilized. When the upper pump bearing unit 10 comprises a first radial bearing 11 , the upper end of the pump shaft 5, here the drive end 51 , is stabilized. When both the lower pump bearing unit 20 and the upper pump bearing unit 10 comprise a first radial bearing 11 , the non-drive end 52 and the drive end 51 of the pump shaft 5 are more stabilized and therewith the entire pump shaft 5.

According to the invention the vertical pump 1 comprises at least one first radial bearing 11 for supporting the rotor 3 with respect to a radial direction, wherein each first radial bearing 11 is configured as a tilting pad journal bearing comprising a first support carrier 12 (Fig. 3) and a plurality of first pads 13 arranged on the first support carrier 12 for supporting the pump shaft 5 by means of a lubricant, wherein each first radial bearing 11 comprises at least one spring element 14 which is arranged to act in a load direction L (Fig. 3) on the first support carrier 12 or on one of the first pads 13 for placing the pump shaft 5 relative to the first radial bearing 11 into a preferred position, in which the lubricant is squeezed between at least one of the first pads 13 and the pump shaft 5.

Within the scope of this application the term “first radial bearing” 11 designates a tilting pad journal bearing, which comprises at least one spring element 14 which acts on the first support carrier 12 or on at least one first pad 13. Thus, the term “first radial bearing” 11 designates such a tilting pad journal bearing, in which the first support carrier 12 in its entirety is spring-loaded or at least one of the first pads 13 is spring-loaded to place the pump shaft 5 into the preferred position with respect to the radial direction. The at least one spring element 14 generates a spring load pushing the first support carrier 12 or the first pad(s) 13 in the constant load direction L so that the first radial bearing 11 in the vertical pump 1 is loaded by the at least one spring element 14 in at least a similar manner as a radial bearing is loaded by 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 as will be explained in more detail hereinafter.

Within the scope of this application the term “second radial bearing” 19 designates a radial bearing or a journal bearing having no spring elements for loading the radial bearing. The second radial bearing 19 may be configured for example as a tilting pad journal bearing as it is known in the art. In particular, the term “second radial bearing” 19 designates a tilting pad journal bearing having a second support carrier 16 (Fig. 3) and at least one second pad 17 arranged to be supported by the second support carrier 16, wherein neither the second support carrier 16 nor anyone of the second pad(s) 17 are spring-loaded for generating a load acting on the pump shaft 5 in the radial direction.

Of course, the vertical pump 1 according to the invention may comprise more than one second radial bearing 19 as will be explained in more detail hereinafter.

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 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 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) will be explained in more detail referring to several embodiments, in particular of the first radial bearing 11 . It has do 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 (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 the lower pump bearing unit 20 in a schematic representation in a cross- sectional view along the pump shaft 5. It goes without saying that also the upper pump bearing unit 10 may be designed in an analogous manner to comprise radial bearings as it will be explained hereinafter with respect to the lower pump bearing unit 20. The lower pump bearing unit 20 comprises an arrangement of a first embodiment of the first radial bearing 11 according to the invention and the second radial bearing 19, wherein the first radial bearing

11 and the second radial bearing 19 are arranged adjacent to each other with respect to the axial direction A.

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 the second support carrier 16 and a plurality of second pads 17 arranged to be supported by the second support carrier 16. The plurality of second pads 17 are supporting the pump shaft 5 by means of the lubricant in a matter known as such. The second pads 17 are arranged on the second support carrier 16 such that the entirety of the second pads 17 surrounds the pump shaft 5. Each second pad 17 is arranged pivotably on the second support carrier 16. To this end the surface of the particular second pad 17 abutting against the second support carrier 16 may be provided with a pivot pin (not shown) or any other pivot member to allow for a pivoting of the second pad 17 relative to the second support carrier 16. 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.

The first radial bearing 11 comprises the first support carrier 12, which is configured and arranged to surround the pump shaft 5. The radially inner surface of the first support carrier

12 faces the pump shaft 5 and is ring-shaped. A plurality of the first pads 13 are arranged on the radially inner surface 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 four first pads 13 arranged around the pump shaft 5 such that the entirety of the first pads 13 completely surrounds the pump shaft 5. It has to be understood that the number of four first pads 13 is only exemplary. In other embodiments of the first radial bearing 11 more or less than four first pads 13 may be provided.

Each first pad 13 has a back side 131 facing the first support carrier 12 and a front side 132 facing the pump shaft 5. Each first pad 13 comprises a pivot 135 (Fig. 4) arranged at the back side 131 of the first pad 13 and supported by the first support carrier 12 such that each first pad 13 can pivot relative to the first support carrier 12. In Fig. 3 the pivots 135 are not shown. Each first pad 13 has a generally arcuate shape and may be configured as a spherical first pad 13 with the back side 131 being spherical or as a cylindrical first pad 13 with the back side 131 having the shape of a cylinder segment. The entirety of the first pads 13 delimits a bearing clearance 136, in which the pump shaft 5 is supported with respect to the radial direction. The bearing clearance 136 is the room which is radially inwardly delimited by the pump shaft 5 and radially outwardly delimited by the front sides 132 of all first pads 13. 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.

In the first embodiment of the first radial bearing 11 the first support carrier 12 is arranged and configured to be movable in the radial direction and to be fixed with respect to the axial direction A, i.e. the first support carrier 12 can move in the radial direction but not in the axial direction A. To this end 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 but allow for a movement of the first support carrier 12 in the radial direction and between the first side wall 121 and the second side wall 122. 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 arranged movably with respect to the radial direction between the first side wall 121 and the second side wall 122, but is prevented from a movement in the axial direction A.

Preferably the first side wall 121 and the second side wall 122 are configured in an identical manner.

Furthermore, the first radial bearing 11 comprises two lateral guide elements 125 for guiding the first support carrier 12 in the radial direction. Each lateral guide element 125 is fixedly mounted to and supported by the pump housing 2 or any other part, which is stationary with respect to the pump housing 2 and fixedly connected to the pump housing 2. The two lateral guide elements 125 are arranged diametrically opposed to each other with respect to the pump shaft 5. Each lateral guide element 125 comprises a planar, i.e. non-curved guide surface 126 facing the first support carrier 12. The two guide surfaces 126 are arranged opposite to each other and parallel to each other for guiding the first support carrier 12 in between the two guide surfaces 126 in the radial direction. The first support carrier 12 is configured with two lateral faces 127 for cooperating with the guide surfaces 126 of the lateral guide elements 125. Each lateral face 127 is configured as a planar, i.e. non-curved surface. The first support carrier 12 has a generally annular shape, wherein the radially outer region of the first support carrier 12 is flattened at two opposing sides to form the two planar lateral faces 127 of the first support carrier 12. Each lateral face 127 faces one of the guide surfaces 126 of the two guide elements 125. During operation of the vertical pump 1 each lateral face 127 may move along one of the guide surfaces 126 in the radial direction.

The first radial bearing 11 further comprises the spring element 14 for loading the first radial bearing 11 in the load direction L. According to the first embodiment of the first radial bearing 11 the entire first support carrier 12 is spring-loaded by the spring element 14 to place the pump shaft 5 into a preferred position. The load direction L of the first radial bearing 11 is a constant direction and defined by the direction of the spring force which is acting on the first support carrier 12. The load direction L is a radial direction, i.e. the spring element 14 exerts a force onto the first support carrier 12, which is directed in the radial direction. This side of the first support carrier 12, on which the spring element 14 acts, is referred to as the loaded side of the first support carrier 12. The opposite side of the first support carrier 12 is referred to as the unloaded side.

The spring element 14 is arranged between the radially outer surface of the first support carrier 12 and the pump housing 2 or any other part, which is stationary with respect to the pump housing 2 and fixedly connected to the pump housing 2. The spring element 14 rests on the first support carrier 12 on the one side and rests on the pump housing 2 or any other part, which is stationary with respect to the pump housing 2 and fixedly connected to the pump housing 2, on the other side.

The spring element 14 may be configured in any known manner, for example as a coil spring or as a disk spring or as a plate spring or as a spring washer just to list some examples. As it can be best seen in Fig. 4, the spring element 14 is arranged in an angular position, which is in the middle between the two lateral guide elements 125, so that the force exerted by the spring element 14 onto the first support carrier 12 is parallel to the two guide surfaces 126 and pushes the first support carrier 12 in the radial direction, more particular in the load direction L.

Thus, by the action of the spring element 14 the pump shaft 5 is pushed into the preferred position, in which the pump shaft 5 is arranged eccentrically with respect to the bearing clearance136 of the first radial bearing 11. This means, that the radial distance between the pump shaft 5 and the first pad(s) 13 varies, when viewed in a circumferential direction of the pump shaft 5. Referring to the representation in Fig. 3, the radial distance C1 , i.e. the distance in the radial direction, between the pump shaft 5 and the first pads 13 is larger on the left side of the pump shaft 5 than the radial distance C2 on the right side of the pump shaft 5. Concurrently, the radial distance C3 between the pump shaft 5 and the second pad(s) 17 of the second radial bearing on the left side in Fig. 3, i.e. on that side which is opposite to the spring element 14 of the first radial bearing 11 becomes smaller than the radial distance C4 between the pump shaft 5 and the second pad(s) 17 of the second radial bearing 19 on the right side in Fig. 3. By this eccentricity of the pump shaft 5 regarding the bearing clearance 136 of the first radial bearing 11 as well as regarding the bearing clearance 176 of the second radial bearing 19, the lubricant is squeezed both between the pump shaft 5 and the first pad(s) 13 on that side of the pump shaft 5, where the radial distance C2 between the pump shaft 5 and the first pads 13 is smaller, namely on the right side of the pump shaft 5 according to the representation in Fig. 3, and between the pump shaft 5 and the second pad(s) 17 of the second radial bearing 19 on that side of the pump shaft 5, where the radial distance C3 between the pump shaft 5 and the second pads 17 is smaller, namely on the left side of the pump shaft 5 according to the representation in Fig. 3. When the lubricant is squeezed between the pump shaft 5 and the first pad(s) 13 and the second pad(s) 17, respectively, the pressure and therewith the force acting on the pump shaft 5 increases at these locations, where the lubricant is squeezed. This results in an equilibrium position, where the force acting on the pump shaft 5 and generated by the spring element 14 equals and compensates the force acting on the pump shaft 5 by the squeezed lubricant. Said equilibrium position, in which the pump shaft 5 is positioned eccentrically with respect to the bearing clearance 136, is the preferred position of the pump shaft 5. Thus, by the action of the spring element 14 the pump shaft 5 is moved in the preferred position. The spring element 14 creates a loading of the first radial bearing 11 in the vertical pump 1 in an analogues manner as the gravity acting on the rotor 3 in a horizontal pump creates a loading of the radial bearing(s). Concurrently, the spring element 14 also creates a loading of the second radial bearing 19, because the pump shaft 5 is also located eccentrically with respect to the bearing clearance 176 of the second radial bearing 19.

Thus, the action of the spring element 14, which spring-loads the first support carrier 12 of the first radial bearing 11 , results in a loading of both the first radial bearing 11 and the second radial bearing 19. The spring element 14 results in an analogous loading of the first radial bearing 11 and the second radial bearing 19 of the vertical pump 1 as the loading of the radial bearing(s) by gravity in a horizontal pump. The loading of the second radial bearing 19 has the same absolute value as the loading of the first radial bearing 11 , but is directed in the opposite direction. The first radial bearing 11 is loaded in the load direction L, whereas the second radial bearing 17 is loaded in the opposite direction.

Fig. 5 shows another embodiment of the lower pump bearing unit 20 or the upper pump bearing unit 10, respectively, in a representation corresponding to the representation in Fig. 3, i.e. in a cross-sectional view along the pump shaft 5. The lower or upper pump bearing unit 20, 10 comprises an arrangement of two first radial bearings 11 each of which is configured according to the first embodiment of the first radial bearing 11 . The two first radial bearings 11 are arranged adjacent to each other with respect to the axial direction A. The two first radial bearings 11 are configured in the same manner, however the load directions L1 and L2 of the two radial first radial bearings 11 are opposite to each other. In the lower first radial bearing 11 in Fig. 5 the spring element 14 is arranged on the right side of the pump shaft 5 resulting in the load direction L1 from the right to the left in Fig. 5. In the upper first radial bearing 11 in Fig. 5 the spring element 14 is arranged on the left side of the pump shaft 5 resulting in the load direction L2 from the left to the right in Fig. 5.

Fig. 6 shows another embodiment of the lower pump bearing unit 20 or the upper pump bearing unit 10, respectively, in a representation corresponding to the representation in Fig. 5. The lower or upper pump bearing unit 20, 10 comprises an arrangement of two first radial bearings 11 as in the arrangement in Fig. 5. In the arrangement shown in Fig. 6 each of the first radial bearings 11 is configured according to a variant of the first embodiment of the first radial bearing 11 .

For a better understanding Fig. 7 shows said variant of the first embodiment of the first radial bearing 11 in a cross-sectional view perpendicular to the pump shaft 5 along the cutting line VII-VII in Fig. 6. The representation in Fig. 7 corresponds to the representation in Fig. 4. The variant of the first embodiment of the first radial bearing 11 has no first pads 13 on the unloaded side of the radial bearing 11 , i.e. the unloaded side of the first support carrier is free of first pads between the pump shaft 5 and the first support carrier 12. Only on the loaded side, i.e. on that side where the spring element 14 acts on the first support carrier 12, a plurality of first pads 13, here three first pads 13, is provided. This variant has the advantage that the number of first pads 13 is reduced which results in lower costs for the first radial bearing 11 .

As shown in Fig. 7 the entirety of the first pads 13 does not completely surrounds the pump shaft, but the three first pads 13 extend roughly speaking only around half of the circumference of the pump shaft 5. Also in Fig. 7 the pump shaft 5 is shown in the preferred position. It can be seen that the pump shaft 5 is arranged eccentrically with respect to the bearing clearance 136. Due to the lack of first pads on the unloaded side of the first support carrier 12 the bearing clearance 136 does not completely surround the pump shaft but is restricted to the room between the pump shaft 5 and the first pads 13.

Fig. 8 shows a schematic representation of another embodiment of the lower pump bearing unit 20 or the upper pump bearing unit 10, respectively, in a cross-sectional view along the pump shaft 5. According to this embodiment the lower or upper pump bearing unit 20, 10 comprises exactly one first radial bearing 11 and no second radial bearing. Said first radial bearing 11 is configured according to a second embodiment of the first radial bearing 11 according to the invention. Variants are possible, where either the lower pump bearing unit 20 or the upper pump bearing unit 10 comprises exactly one first radial bearing 11 and no second radial bearing, wherein said first radial bearing 11 is configured according to the second embodiment of the first radial bearing 11 . According to other variants each of the lower pump bearing unit 20 and the upper pump bearing unit 10 comprises exactly one first radial bearing 11 and no second radial bearing, wherein said first radial bearing 11 is in each case configured according to the second embodiment of the first radial bearing 11 .

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 or functionally equivalent features.

For a better understanding Fig. 9 shows the second embodiment of the first radial bearing 11 in a cross sectional view perpendicular to the pump shaft 5 along the cutting line IX-IX in Fig. 8.

In the second embodiment of the first radial bearing 11 the first support carrier 12 is arranged to be tiltable relative to the pump shaft 5 wherein the at least one spring element 14 is arranged to act on the first support carrier 12 in the axial direction A for tilting the first support carrier relative to the pump shaft 5, such that each first pad 13 is extending obliquely relative to the pump shaft 5 when viewed in the axial direction A.

With exemplary character Fig. 9 shows a configuration with three spring elements 14 which are all arranged on the same side of the pump shaft 5 such that the three spring elements 14 generate a torque acting on the first support carrier 12 for tilting the first support carrier 12 relative to the pump shaft 5. Different from the first embodiment, where the spring force is directed in the radial direction, in the second embodiment the spring force is directed in the axial direction A.

As it is shown in Fig. 8, in the second embodiment of the first radial bearing 11 the first support carrier 12 comprises a back part 128 on which the first support carrier 12 is supported by the pump housing 2 or, alternatively by any other part, which is stationary with respect to the pump housing 2 and fixedly connected to the pump housing 2. On the side of the back part 128, which contacts the pump housing 2, the back part 128 is chamfered so that the first support carrier 12 may smoothly roll or move on the supporting surface.

Different to the first embodiment, in the second embodiment of the first radial bearing 11 there is no need for guiding the first support carrier 12 for a movement of the entire first support carrier 12 in the radial direction. Therefore, in the second embodiment the side walls 121 , 122 (Fig. 3) as well as the lateral guide elements 125 (Fig. 4) are not required.

As it is best seen in Fig. 8 in the second embodiment the first radial bearing 11 comprises a spring support element 221 and an first upper limitation 222 which are arranged on that side of the pump shaft 5, where the spring elements 14 are provided, i.e. the left side in Fig. 8. On the opposite side of the pump shaft 5, i.e. the right side in Fig. 8, a lower limitation 321 and a second upper limitation 322 are provided. Each of the spring support element 221 , the first upper limitation 222, the lower limitation 321 and the second upper limitation 322 is fixedly mounted to the pump housing 2 or to any other part that is stationary with respect to and fixedly connected to the pump housing 2. Furthermore, each of the spring support element 221 , the first upper limitation 222, the lower limitation 321 and the second upper limitation 322 extends in the radial direction, wherein none of them juts out beyond anyone of the first pads 13 with respect to the radial direction. The spring support member 221 and the first upper limitation 222 are arranged on different heights with respect to the axial direction A, so that there is a first distance D1 between the spring support member 221 and the first upper limitation 222 with respect to the axial direction A. The lower limitation 321 and the second upper limitation 322 are arranged on different heights with respect to the axial direction A, so that there is a second distance D2 between the lower limitation 321 and the second upper limitation 322 with respect to the axial direction A.

Both the first distance D1 and the second distance D2 are larger than the extension of the first support carrier 12 in the axial direction A. With respect to the axial direction A the first support carrier 12 is arranged between the spring support member 221 and the first upper limitation 222 on the one side of the pump shaft 5, namely the left side in Fig. 8, and between the lower limitation 321 and the second upper limitation 322 on the other side of the pump shaft 5, namely the right side in Fig. 8.

The three spring elements 14 are arranged between the spring support member 221 and the first support carrier 12. Thus, the load direction L, i.e. the direction in which the spring elements 14 act on the first support carrier 12, is the axial direction A. The first distance D1 is configured such, that the spring elements 14 and the first support carrier 12 can be arranged between the spring support member 221 and the first upper limitation 222, and that the tilting movement of the first support carrier 12 is not hindered by the spring support member 221 or by the first upper limitation 222. Thus, the first support carrier 12 can be tilted by the action of the spring elements 14 without getting in direct physical contact with the spring support member 221 or the first upper limitation 222.

The second distance D2 is smaller than the first distance D1 but larger than the extension of the first support carrier 12 in the axial direction A such that the first support carrier 12 can be tilted between the lower limitation 321 and the second upper limitation 322. When the first support carrier 12 is tilted such that the pump shaft 5 is in the preferred position as it is shown in Fig. 8 the axially upper end of the first support carrier 12 abuts against the second upper limitation 322. Thus, the second upper limitation 322 prevents a movement of the first support carrier 12 in its entirety in the axial direction A. The second upper limitation 322 has an extension in the radial direction which is smaller than the extension of the lower limitation 321. The extension of the second upper limitation 322 in the radial direction is such, that the second upper limitation 322 on the one hand prevents an movement of the entire first support carrier 12 in the axial direction A, an on the other hand does not hinder the tilting of the first support carrier 12.

When the pump shaft 5 is in the preferred position relative to the first radial bearing 11 (see Fig. 8) a wedge is built between each first pad 13 and the pump shaft 5, so that for each particular first pad 13 the radial distance between the pump shaft 5 and the particular first pad 13 increases or decreases when viewed in the axial direction A, i.e. the bearing clearance 136 varies with respect to the axial direction

Fig. 10 shows a schematic representation of a third embodiment of a first radial bearing 11 according to the invention in a representation corresponding to Fig. 8, namely in a cross- sectional view along the pump shaft 5.

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

Similar as in the second embodiment also in the third embodiment of the first radial bearing 11 the first support carrier 12 is arranged to be tiltable relative to the pump shaft 5, wherein the at least one spring element 14 is arranged to act on the first support carrier 12 in the axial direction A for tilting the first support carrier 12 relative to the pump shaft 5, such that each first pad 13 is extending obliquely relative to the pump shaft 5 when viewed in the axial direction A.

However, the third embodiment of the first radial bearing 11 is configured more symmetrically than the second embodiment. As it is shown in Fig. 10, in the third embodiment of the first radial bearing 11 the first support carrier 12 is spring-loaded on both sides of the pump shaft 5 to cause the tilting of the first support carrier 12 relative to the pump shaft 5. On the one side of the pump shaft 5, namely the left side in Fig. 10, the spring element(s) 14 act(s) on the first support carrier 12 to push it upwardly regarding the axial direction A, and on the other side of the pump shaft 5, namely the right side in Fig. 10, the spring element(s) 14 act(s) on the first support carrier 12 to push it downwardly regarding the axial direction A, Thus, the spring force generated by the spring element(s) 14 and acting on the first support carrier 12 on the one side of the pump shaft 5 is oppositely directed to the spring force acting on the first support carrier 12 on the other side of the pump shaft 5. The two spring forces build a couple of forces generating a torque acting on the first support carrier 12 which tilts the first support carrier 12 relative to the pump shaft 5. Preferably, the spring elements 14 are configured such that the absolute value of the oppositely directed spring forces is the same.

The third embodiment of the first radial bearing further comprises a lower support element 223 for supporting the spring element(s) 14 on the one side of the pump shaft 5, and an upper support element 224 for supporting the spring elements 14 arranged on the other side of the pump shaft 5. The first support carrier 12 and all spring elements 14 are arranged with respect to the axial direction A between the lower support element 223 and the upper support element 224. Both the lower and the upper support element 223, 224 are 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 of the lower and the upper support member 223 and 224 has a generally annular shape and is arranged to surround the pump shaft 5. Furthermore, each of the lower and the upper support member 223 and 224 extends in the radial direction and has an inner diameter which is configured such that the lower and the upper support 223, 224 do not jut out beyond anyone of the first pads 13 but allow for a tilting movement of the first support carrier 12 relative to the pump shaft 5 and between the lower and the upper support element 223, 224. The lower support element 223 and the upper support element 224 are arranged parallel to each other.

On the one side of the pump shaft 5, namely the left side in Fig. 10, the spring elements 14 are all arranged between the lower support element 223 and the first support carrier 12. On the other side of the pump shaft 5, namely the right side in Fig. 10, the spring elements 14 are all arranged between the upper support element 224 and the first support carrier 12.

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

In the following description of the fourth embodiment of the first radial bearing 11 only the differences to the first, the second and the third embodiments are explained in more detail. The explanations with respect to the first, the second and the third 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 fourth embodiment. Same reference numerals designate the same features that have been explained with reference to the first, the second and third embodiment or functionally equivalent features.

The main difference between the fourth embodiment of the first radial bearing 11 and the first three embodiments is that according to the fourth embodiment the at least one spring element 14 directly acts on at least one of the first pads 13.

Thus, at least one of the first pads 13 is configured as a spring-loaded pad 13a for placing the pump shaft into the preferred position, wherein the spring element 14 is arranged between the first support carrier 12 and the spring-loaded pad 13a for pushing the spring-loaded pad 13a in the load direction L.

Regarding the fourth embodiment of the first radial bearing 11 it will be differentiated between first pads 13 which are directly loaded by a spring element 14 and first pads 13 which are not directly loaded by a spring element 14. Those first pads 13 which are directly loaded by one of the spring elements 14, meaning that one of the spring elements 14 rests against said first pad 13, are referred to as “spring-loaded pads” and designated with the reference numeral 13a. Those first pads 13, which are not directly loaded by one of the spring elements, meaning that none of the spring elements 14 rests against said first pad 13, are referred to as “unloaded pads” and designated with the reference numeral 13b.

In the fourth embodiment of the first radial bearing shown in Fig. 11 , there are two spring loaded pads 13a which are arranged next to each other with respect to the circumferential direction of the pump shaft 5 and four unloaded pads 13b. There are two spring elements 14, namely one spring element 14 for each spring loaded pad 13a. Each spring element 14 is arranged between the first support carrier 12 and one of the spring-loaded pads 13a with respect to the radial direction.

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 the pump housing 2. The radially inner surface of the first support carrier 12 faces the pump shaft 5 and is ring-shaped. The unloaded pads 13b are arranged on the radially inner surface of the first support carrier 12 for supporting the pump shaft 5 with respect to the radial direction by means of the lubricant. Each unloaded pad 13b comprises the pivot 135 arranged at the back side 131 of the unloaded pad 13b and supported by the first support carrier 12 such that each unloaded pad 13b can pivot relative to the first support carrier 12. Between each spring-loaded pad 13a and the radially inner surface of the first support carrier 12 one of the spring elements 14 is arranged, such that the spring element 14 rests on the one side against the back side 131 of the spring-loaded pad 13a, and on the other side against the radially inner surface of the first support carrier 12. Thus, each spring-loaded pad 13a is pushed by the spring element 14 acting on its back side 131 in radial direction towards the pump shaft 5.

The entirety of the unloaded pads 13b and the spring-loaded pads 13a delimits the bearing clearance 136, in which the pump shaft 5 is supported with respect to 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 on the one side and the unloaded and spring-loaded pads 13b, 13a on the other side.

By the action of the spring elements 14 pushing the spring-loaded pads 13a towards the pump shaft 5, the pump shaft 5 is placed into the preferred position, in which the pump shaft 5 is arranged eccentrically with respect to the bearing clearance136. By this eccentricity of the pump shaft 5 regarding the bearing clearance 136 of the first radial bearing 11 the lubricant is squeezed between the pump shaft 5 and the spring-loaded pads 13a. Since the lubricant is squeezed between the pump shaft 5 and the spring-loaded pads 13a the pressure and therewith the force acting on the pump shaft 5 increases at these locations, where the lubricant is squeezed. This results in an equilibrium position, where the force acting on the pump shaft 5 by the spring-loaded pads 13a equals and compensates the force acting on the pump shaft 5 by the squeezed lubricant. Said equilibrium position, in which the pump shaft 5 is positioned eccentrically with respect to the bearing clearance 136, is the preferred position of the pump shaft 5.

It has to be noted, that the number of two spring-loaded pads 13a and four unloaded pads 13b has an exemplary character. There are many different combinations of spring-loaded pads 13a and unloaded pads 13b which result in the effect of placing the pump shaft 5 relative to the first radial bearing 11 into the preferred position, in which the lubricant is squeezed between the spring-loaded pads 13a and the pump shaft 5.

It is preferred that there is at least one of the first pads 13 which is configured as an unloaded pad 13b.

Particularly preferred are configurations comprising a plurality of spring-loaded pads 13a for placing the pump shaft 5 into the preferred position, and a plurality of unloaded pads 13b, wherein all spring-loaded pads 13b are arranged side by side when viewed in a circumferential direction of the pump shaft 5, meaning that the sequence of spring-loaded pads 13a arranged next to each other with respect to the circumferential direction is not interrupted by an unloaded pad 13b arranged between two adjacent spring-loaded pads 13a.

Fig. 12 shows still another embodiment of the lower pump bearing unit 20 or the upper pump bearing unit 10, respectively, in a cross-sectional view along the pump shaft 5. The lower or upper pump bearing unit 20, 10 comprises an arrangement of a fifth embodiment of a first radial bearing 11 according to the invention and one second radial bearing19. Fig. 13 shows a schematic representation of the fifth embodiment of the first radial bearing 11 in a cross- sectional view perpendicular to the pump shaft 5 along the cutting line XIII-XIII in Fig. 12.

Fig. 14 shows a plan view of the front side 132 of the spring loaded pad 13a of the fifth embodiment of the first radial bearing 11 .

In the following description of the fifth embodiment of the first radial bearing 11 only the differences to the first, the second, the third and the fourth embodiments are explained in more detail. The explanations with respect to the first, the second, the third and the fourth 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 fifth embodiment. Same reference numerals designate the same features that have been explained with reference to the first, the second, the third and fourth embodiment or functionally equivalent features. Similar to the fourth embodiment of the first radial bearing 11 the fifth embodiment of the radial bearing comprises at least one first pad 13 which is configured as a spring-loaded pad 13a.

As it is shown in Fig. 13 the fifth embodiment of the first radial bearing 11 comprises only one spring-loaded pad 13a and no unloaded pad 13b. It has to be understood that the fifth embodiment may also comprise more than one spring-loaded pad 13a and/or at least one unloaded pad 13b. Since it is sufficient for the understanding reference is made to the configuration with only one spring-loaded pad 13a and no unloaded pad 13b as it is shown in Fig. 13. The spring-loaded pad 13a is configured to extend approximately around half of the outer circumference of the pump shaft 5. Furthermore, the spring-loaded pad 13a is configured for supplying a pressurized fluid through the interior of the spring-loaded pad 13a to the bearing clearance 136. In the fifth embodiment of the first radial bearing the loading of the first radial bearing is generated by the combination of a spring force and a hydraulic force generated with the pressurized fluid. Preferably, the pressurized fluid is the same as the lubricant.

The spring-loaded pad 13a comprises a channel 138 for a fluid communication between the front side 132 and the back side 131 of the spring-loaded pad 13a. The channel 138 extends generally in the radial direction from the back side 131 to the front side 132.

Fig. 14 shows a plan view of the front side 132 of the spring-loaded pad 13a. As it can be seen in Fig. 14 the front side 132 is provided with an elongated circumferentially extending cavity 139 in which the channel 138 ends. At the back side 131 of the spring-loaded pad 13a the channel 138 is in fluid communication with an supply port 80 which extends in radial direction through the first support carrier 12 from the radially inner surface of the first support carrier 12 to the radially outer surface of the first support carrier 12, where the supply port 80 is in fluid communication with an annular chamber 85. The annular chamber 85 receives the pressurized lubricant from a source for the lubricant.

The back side 131 of the spring-loaded pad 13a is provided with a stub 81 fixedly connected to the back side 131 , extending in the radial direction and configured to tightly fit into the support port 80 of the first support carrier 12, such that the stub 81 is still movable in the radial direction relative to the support port 80. Furthermore, the stub 81 is arranged to surround the orifice of the channel 138 in the back side 131 of the spring-loaded pad 13a, so that the pressurized lubricant is guided from the annular chamber 85 through the stub 81 arranged in the supply port 80 of the first support carrier 12 and through the channel 138 into the cavity 139 in the front side 132 of the spring-loaded pad 13a. The spring element 14 for spring-loading the spring-loaded pad 13a is arranged in the supply port 80 of the first support carrier 12 between a shoulder 82 provided in the supply port 80. and the radially outer end of the stub 81 . At the shoulder 82 the diameter of the supply port 80 becomes smaller, so that the spring element 14 can rest against the shoulder 82 and is supported by the shoulder 82. At the other side the spring element 14 rests against the radially outer end of the stub 81 , so that the stub 81 is spring-loaded by the spring element 14 in the radial direction towards the pump shaft 5. Since the stub 81 is fixedly connected to the back side 131 , the spring-loaded pad 13a is spring-loaded by the spring element 14.

A sealing element 83 such as an O-ring may be provided and arranged around the stub 81 to prevent a leakage of the lubricant along the outer side of the stub 81 between the stub 81 and the supply port 80.

During operation of the vertical pump 1 the combined action of the spring element 14 and the pressurized lubricant in the cavity 139 places the pump shaft 5 relative to the first radial bearing 11 into the preferred position, in which the lubricant is squeezed between the spring- load pad 13a and the pump shaft 5. In this preferred position the pump shaft 5 has an eccentric position in the bearing clearance 136, meaning that the radial distance between the pump shaft 5 and the spring-loaded pad 13a varies when viewed in the circumferential direction of the pump shaft 5.