US20190063450A1

US20190063450A1 - Centrifugal pump with balancing device and a method of balancing axial forces of the centrifugal pump

Info

Publication number: US20190063450A1
Application number: US16/081,471
Authority: US
Inventors: Matti Koivikko; Kalle TIITINEN; Hannu HEISKANEN; Teemu GASMAN
Original assignee: Sulzer Management AG
Current assignee: Sulzer Management AG
Priority date: 2016-03-17
Filing date: 2017-03-06
Publication date: 2019-02-28
Also published as: EP3430270B1; RU2018132444A3; WO2017157702A1; BR112018068267A2; CN108779782A; RU2018132444A; EP3430270A1

Abstract

A single-stage or multi-stage centrifugal pump has a balancing disc for balancing the axial forces of the pump. The balancing disc includes at least one annular groove in at least one of the non-axial counter surfaces and of the balancing disc and the counter member.

Description

CROSS-REFERENCE TO RELATED APPLICATION

This application is a U.S. National Stage application of International Application No. PCT/EP2017/055118, filed Mar. 6, 2017, which claims priority to European Patent Application No. 16160876.5, filed Mar. 17, 2016, the contents of each of which are hereby incorporated herein by reference.

BACKGROUND

Field of the Invention

The present invention relates to a centrifugal pump with a balancing member and a method of balancing axial forces of a centrifugal pump. More specifically the present invention relates to single- or multi-stage centrifugal pumps having a novel disc-type device for balancing the axial forces of the pump.

Background Information

Devices for balancing axial forces of centrifugal pumps are normally in use in multistage pumps, which have a high pressure head, and include several subsequent centrifugal impellers on the same shaft. An axial force is generated while an impeller, or a plurality of impellers, draws liquid axially in the pump and discharges the liquid radially from the pump. The axial force tends to draw the impeller(s) towards the pump inlet, whereby the bearings of the pump are subjected to a considerable axial force when keeping the pack of impellers in place. In order to reduce the axial force subjected to the bearings, and, thus, to make it possible to use smaller or lighter bearings or different types of bearings for balancing the axial force have been developed.
Conventional pumps use two basic types of devices for balancing the axial force. One is a so-called drum-type balancing device, and the other is a disc-type balancing device. Also hybrid balancing devices are known, i.e. one comprising both a balancing drum and a balancing disc. In most cases the balancing devices are positioned on the pump shaft behind the last impeller when viewed from the pump inlet towards the pump outlet. However, it is possible, if desired, to construct a centrifugal pump such that the balancing devices are between the stages of a multi-stage centrifugal pump or in front of the impeller(s) thereof. The disc-type balancing device may be considered as the preferred choice of the two basic balancing devices as it adjusts its operation automatically, i.e. slight wear does not affect the operation of the balancing device at all, whereas even the slightest wear of the drum-type balancing device results in a change in the balancing capability of the balancing device. Furthermore, the disc-type balancing device occupies less space in the axial direction than drum-type balancing device.
DE 539 225 discusses an exemplary hybrid balancing device, i.e. an axial thrust relief by f stepped relief disks and radial and axial throttle gaps. The intention of the axial thrust relief is to reduce the heavy wear, which the graduated relief disk is subjected to on the sealing surface, especially when high pressures and high temperatures occur during the conveying of water, as in the case of newer boiler feed pumps, or when the liquid or its admixtures are aggressive. The cause of the wear of the washers lies in the too high water velocities, which are intended to be reduced by the invention. This is achieved by virtue of the fact that the radial and axial throttle gaps open at their interface into annular calming chambers, which in a manner known per se are connected by conduits with specific pump stages for pressure equalization. In order to compensate for the possible wear on the steps of the relief disks, control elements (valves, cocks, etc.) are connected into the connecting lines of the stabilizing chambers with the pump pressure stages. However, the disclosed balancing device is very complicated of its construction including multiple conduits to various pumping stages and valves arranged to the conduits.
DE 1 082 803 discusses a pair of relief discs for multi-stage centrifugal pumps. The pair of centrifugal discs is an important component of multi-stage centrifugal pumps. It is known that the disk pair causes operating disturbances due to distortion of the pump housing. For example the start-up of the pump easily results in a metallic contact of the flat relief discs. The German document overcomes the described malfunction by applying a hollow ball section and a ball section as a pair of relief discs, the centers of the curvature of which lie in the center of the ball bearing of the shaft bearing. This precludes the possibility of a metallic contact between the disks, thus providing a high degree of operational safety. However, due to the use of a ball-type shaft bearing the clearance between the discs is always constant, whereby the basic object and advantage of the relief discs, i.e. the balancing of the axial thrust, is lost.
US-B1-6,309,174 discusses a pump having a thrust bearing and a casing with an inlet port and a discharge port. A shaft is positioned within the pump. A thrust bearing is coupled to the shaft. A pressure alterable bearing cavity is located within the casing. The pressure alterable cavity allows the axial thrust on the shaft from the impellers to be counteracted. The thrust bearing has an annular seal coupled with respect to the casing disc is coupled to the shaft and positioned adjacent the seal. The seal and the disc have a gap therebetween. A feedback pipe couples the bearing cavity to the inlet port. As the axial thrust acting on the shaft changes, the gap between the seal and the disc also changes. The changing of the gap changes the pressure within the bearing cavity. The disc and thus the shaft are repositioned in response to the change in the bearing cavity pressure change.
DE-867933 discusses a centrifugal pump disposed in the area behind the impeller with a sealing in the form of a thin gap arranged between two radially extending discs, one of the discs being stationary and the other rotary. A further option discussed in the DE-document is the use of the discs for balancing the axial force created by the impeller. The balancing device of the DE-document has in radial surfaces of both of its discs grooves facing one another. The groove in the rotary disc has transverse vanes that together with the groove in the stationary disc act as a side channel pump. The side channel pump raises pressure and thereby is able to pump liquid along a recirculation passage to the pressure space behind the impeller. The recirculation passage has a throttling for adjusting the axial force created in the balancing cavity between the discs.

SUMMARY

The disc-type balancing device the of present invention as discussed in more detail later is formed of a balancing disc fastened on the shaft of the pump and a stationary counter member. In most cases the counter member is arranged to extend from the pump volute or casing radially inwardly between the impeller or one of the impellers and the balancing disc. Often the stationary counter member is the rear wall of the last pumping stage of the centrifugal pump. The balancing disc and the counter member leave a radially extending cavity, so called balancing cavity, therebetween. Either the balancing disc or the counter member or both have an annular axial extension, sometimes a separate circular ring, at the outer periphery of the balancing disc for reducing the axial dimension of the cavity between the balancing disc and the counter member in order to limit the leakage flow of the pressurized liquid from the pump. However, it should be understood that the balancing device, i.e. balancing disc, its counter member and the balancing cavity, may also be located in front of the impeller/s when viewed from the direction of the inlet opening of the pump. In such a case it is required that the pressurized liquid is taken to the balancing cavity along a separate flow passage.
The disc-type balancing device functions such that a part of the liquid pressurized by the impeller or the pack of impellers enters, as is well known in centrifugal pumps, to the cavity behind the impeller of the last pumping stage, and finds its way via the gap between the shaft of the pump or the shaft sleeve of the balancing disc and the stationary counter member to a radially extending balancing cavity between the balancing disc and the stationary counter member. Now that the pressure of the liquid is, in practice, not reduced the full pressure of the pumped liquid affects on the rotary balancing disc pushing the balancing disc away from inlet of the pump, i.e. contrary to the axial force created by the impellers. Thereby, the axial thrust loading the bearings of the pump is the difference of the two axial forces having opposite directions. By properly dimensioning the balancing device the two opposite forces may be equalized resulting in zero thrust, whereby the shaft bearings may be changed into slide bearings that are not able to carry any axial load.
However, while the pressurized liquid flows radially outwardly in the balancing cavity between the balancing disc and its counter member, the liquid reaches the annular extension or ring and enters the annular gap between the annular extension or ring and its counter surface. Now that the annular gap is very thin, i.e. its depth is very small, and the pressure difference radially over the ring is relatively high (depending mostly on the head of the pump), the flow velocity of the liquid in the thin gap is high. Due to the high velocity of the liquid the pressure in the gap between the balancing disc and the counter member is low resulting in that in the area of high flow velocity, i.e. at the ring area, the disc is not able to create any significant axial force. The result, in appropriate conditions, is that a part of the liquid flow entering the gap between the ring and its counter surface evaporates temporarily to vapor. Especially in such a condition that the pressure difference over the balancing device is high compared to how far from the steam pressure the balancing device operates. The temporary evaporation of the liquid in the thin gap results easily in mechanical contact between the ring and its counter surface, which, in the least, increases friction losses, and raises the temperature of the surfaces. Also, sudden evaporation of the liquid may lead to impacts between the counter surfaces as they hit one another. Both the friction and the impacts may, in the long run, cause wear, which may over time lead to need for replacing the balancing device with a new one. In other words, the first problem that may be seen in the operation of the balancing device is high power consumption combined with fluctuations in the power consumption due to the balancing device operating, alternatingly, in both low-friction and high-friction situations.
The above described problems, first of all the mechanical contact between the counter surfaces, have been suggested to be solved by increasing the effective area of the balancing disc by increasing the diameter of the balancing disc. It results in considerable increase in the power consumption of the balancing device without, however, preventing the liquid from boiling in all operating conditions of the pump. In other words, the prior art improvement leads to increased power consumption and occasional wear-related problems.
Thus an object of the present invention is to design such a novel balancing device for a centrifugal pump that reduces the power consumption of the balancing device.
Another object of the present invention is to design such a novel balancing device for a centrifugal pump that considerably reduces or obviates totally the fluctuations in the power consumption of the balancing device.
A further of the present invention is to design such a novel balancing means for a centrifugal pump that prevents the mechanical contact between the balancing disc and its counter member.
A still further object of the present invention is to develop such a novel balancing device for a centrifugal pump that adjusts automatically its operating clearance.
At least one of the problems is solved and at least one of the objects of the present invention are met with a centrifugal pump comprising a pump casing with an inlet and an outlet, a shaft sealed and mounted with slide bearings to the pump casing, the shaft being movable in axial direction, at least one impeller fastened on the shaft for rotation therewith and a device for balancing axial forces the balancing device comprising a balancing disc fastened on the shaft for rotation therewith and having an outer circumference, and a stationary counter member arranged in connection with the pump casing; the balancing disc and the counter member leaving therebetween a balancing cavity; the balancing disc having a first non-axial surface delimited to the outer circumference and the counter member having a second non-axial surface, the first and the second non-axial surfaces facing one another and leaving a thin gap therebetween, wherein at least one of the first and the second non-axial surfaces including at least one annular groove, and a throttling downstream of the balancing disc, the counter member includes a counter ring having the second non-axial surface and/or the balancing disc includes an annular ring having the first non-axial surface.
Other characteristic features of the present invention become apparent in the appended dependent claims.
The present invention brings about the following advantages over the prior art balancing device

- no or very small fluctuations in the power consumption,
- lower power consumption than traditional disc-type balancing device,
- smaller disc radius than traditional disc-type balancing device,
- continuous adjustment of the gap between the balancing disc and its counter surface,
- the balancing disc is capable of carrying load for the entire radial extension thereof,
- the balancing device may be easily adjusted for different pump properties, especially to varying head, and
- possible wear has no effect on the function of the balancing device contrary to drum-type balancing device.

DESCRIPTION OF THE DRAWINGS

The invention will be explained in more detail hereinafter with reference to the drawings.

FIG. 1 illustrates schematically, and in an axial cross section, a multi-stage centrifugal pump including a disc-type balancing device in accordance with a first preferred embodiment of the present invention;

FIG. 2 illustrates schematically an axial, more detailed cross section of the balancing device in accordance with a first preferred embodiment of the present invention;

FIG. 3 illustrates a detailed axial cross section of the balancing device in accordance with a first variation of a first preferred embodiment of the present invention;

FIG. 4 illustrates a detailed axial cross section of the balancing device in accordance with a second variation of the first preferred embodiment of the present invention;

FIG. 5 illustrates a detailed axial cross section of the balancing device in accordance with a third variation of the first preferred embodiment of the present invention;

FIG. 6 illustrates a detailed axial cross section of the balancing device in accordance with a second preferred embodiment of the present invention;

FIG. 7 illustrates a detailed axial cross section of the balancing device in accordance with a third preferred embodiment of the present invention;

FIG. 8 illustrates a detailed axial cross section of the balancing device in accordance with a fourth preferred embodiment of the present invention;

FIG. 9 illustrates a detailed axial cross section of the balancing device in accordance with a fifth preferred embodiment of the present invention;

FIG. 10 illustrates a detailed axial cross section of the balancing device in accordance with a sixth preferred embodiment of the present invention;

FIG. 11 illustrates a detailed axial cross section of the balancing device in accordance with a seventh preferred embodiment of the present invention; and

FIG. 12 illustrates a detailed axial cross section of the balancing device in accordance with an eighth preferred embodiment of the present invention.

DETAILED DESCRIPTION OF THE EMBODIMENTS

FIG. 1 illustrates an axial cross section of a multi-stage centrifugal pump having a casing 10 with an inlet 12 and an outlet 14, the casing 10 housing a plurality of, here four, impellers 16 attached on a shaft 18 for rotation therewith and a balancing device (means) 20.
FIG. 2 illustrates schematically an axial, more detailed cross section of the balancing device 20 and the end part of the centrifugal pump in accordance with a first preferred embodiment of the present invention. Here in this embodiment the balancing device 20 is formed of a balancing disc 22 attached on the shaft 18 for rotation therewith. In connection with the balancing disc 22 there may be a separate sleeve or the balancing disc may include an integrated axial extension, i.e. a cylindrical sleeve 24, either one of the sleeves extending from the disc up to the hub of the impeller 16. The balancing device 20 further comprises a counter member 28 extending from the pump casing 10 radially inwardly between the balancing disc 22 and the impeller 16. The counter member 28 is, in this embodiment, either the rear wall of the centrifugal pump or a specific part attached thereto. In more general terms, the counter member is a part of the casing of the centrifugal pump or a specific part attached thereto. The counter member 28, preferably but not necessarily, includes a counter ring 26 attached to the counter member 28 such that it faces the area of the balancing disc 22 immediately radially inside the outer circumference of the balancing disc 22. The Figure shows by the black thick flow line how the pumped liquid is able to flow from the rear side cavity 30 of the impeller 16 to a radial clearance 32 between the sleeve 24 and the inner circumference 34 of the counter member 28 to an outwardly extending balancing cavity 36 between the balancing disc 22 and the counter member 28. At the radially outer end of the balancing cavity 36 the liquid flows via a thin gap 36′ between the balancing disc 22 and the counter member 28 (here in this embodiment the counter ring 26 is the part of the counter member 28 facing the balancing disc 22) to a space radially outside the balancing disc 22. The space outside the balancing disc 22 is in flow communication with the cavity 38 axially behind the balancing disc 22 when viewed from the direction of the pump inlet 12. The liquid leaked through the balancing device 20 continues towards the slide bearings 40 of the pump shaft 18 such that the liquid is used to lubricate the slide bearings 40 while passing the bearings 40. When having passed the bearings 40 the liquid enters the end cavity 42 of the pump from where it is either introduced via pipeline 44 to the suction of the pump or to the bearings at the opposite end of the pumps shaft 18. The bearings 40 form a fixed throttling in the flow passage of the leaked liquid keeping the amount of leaked liquid the desired one.
FIG. 3 illustrates a detailed axial cross section of the balancing device 20 in accordance with a first variation of a first preferred embodiment of the present invention. The balancing disc 22 has a counter surface 50 facing the counter ring 26. The counter surface 50 and the surface 52 of the counter ring 26 facing thereto are, coated with or otherwise manufactured of a material suitable for working as counter surfaces. The counter ring 26 is fastened on the counter member 28 or in an annular groove in the counter member 28 by screws, adhesives, welding, riveting, just to name a few examples without any intention of limiting the options to the listed ones. The counter ring 26 has in its surface 52 facing the surface 50 of the balancing disc 22 a, preferably but not necessarily, circular groove 54. The groove 54 is, preferably but not necessarily, rectangular of its cross section and has a depth of 1 to tens of millimeters depending on the liquid to be pumped, i.e. the more foreign abrasive material the liquid carries the deeper the grooves should be to allow the depth of the grooves to wear down without losing their ability to work in the desired manner. The basic property of the groove is that it increases the cross sectional flow area in the gap 36′ between the balancing disc 22 and the counter ring 26 by subjecting the leakage flow from the gap 36′ to sudden expansion and contraction type flow resistances. In other words, when entering the groove the liquid loses its flow velocity almost entirely, and when entering the thin gap again the liquid has to be accelerated to the flow velocity corresponding to the thin gap. The thickness dimension of the flow area is increased, when entering the groove, from a micron range to a millimeter range, i.e. to a value from 1 mm to tens of millimeters. In other words, the cross sectional flow area, or in fact the radial depth thereof (as the width, i.e. the circumference remains substantially the same), is increased to at least 10-fold, preferably to more than 50-fold or more 100-fold depending again on the type of liquid to be pumped. And after the groove, the same is decreased back to micron range again.
As to the orientation of the balancing disc 22 and its counter member 28 it is preferably radial, as in such a case, the axial dimension the balancing device requires is the smallest. However, the advantages of the present invention are available as soon as the direction of the counter surfaces 50 and 52 of the thin gap 36′ clearly differs from axial direction. In other words, as soon as the movement of the counter surfaces 50 and 52 relative to one another cause a change in the dimension of the gap 36′ the advantages of the invention are available. Thus, the basic requirement for the direction of the counter surfaces 50 and 52 is that the direction thereof is non-axial. However, it could be assumed that the orientation of the counter surfaces 50 and 52 should be between 30 and 90 degrees from the direction of the axis A (see FIG. 2) of the pump.
FIG. 4 illustrates a detailed axial cross section of the balancing device in accordance with a second variation of the first preferred embodiment of the present invention. Here, in the second variation the counter ring 26′ is not radial but somewhat inclined as was discussed already above. Naturally, also the counter surface in the balancing disc 22′ is inclined in a corresponding manner. In the variation having inclined non-axial counter surfaces the same advantage as in the radial variation is gained, i.e. the gap between the surfaces adjusts automatically such that the gap remains the same for the entire inclined length thereof. A further advantage gained with the inclined counter surfaces is the fact that for a certain radial dimension a higher number of grooves may be fitted in an inclined counter surface than in a radial counter surface.
FIG. 5 illustrates a detailed axial cross section of the balancing device in accordance with a third variation of the first preferred embodiment of the present invention. Here, in the third variation the counter ring 26″ is not radial but somewhat inclined as was discussed already above. Naturally, also the counter surface in the balancing disc 22″ is inclined in a corresponding manner. Furthermore, the entire balancing disc 22″ is also inclined. [0039] The inclination of the machine elements discussed above may be needed for some constructional reasons. For instance, with the structure of FIG. 5 the shaft bearing may be brought somewhat closer to the impeller.
FIG. 6 illustrates a detailed axial cross section of the balancing device in accordance with a second preferred embodiment of the present invention. Here, in the second embodiment, the groove 64, similar to that in the first embodiment is arranged in the balancing disc 22.
FIG. 7 illustrates a detailed axial cross section of the balancing device in accordance with a third preferred embodiment of the present invention. Here, in the third embodiment, there are several grooves 74, similar to that in the first embodiment arranged in either the counter ring 26 (as shown) or in the balancing disc 22 (not shown).
FIG. 8 illustrates a detailed axial cross section of the balancing device in accordance with a fourth preferred embodiment of the present invention. Here, in the fourth embodiment, there are grooves 84′ and 84″ in both the balancing disc 22 and in the counter ring 26, the grooves being similar to that in the first embodiment.
FIG. 9 illustrates a detailed axial cross section of the balancing device in accordance with a fifth preferred embodiment of the present invention. Here, in the fifth embodiment, there are grooves 84′ and 84″ in both the balancing disc 22 and in the counter ring 26, the grooves being similar to that in the first embodiment. Additionally, the balancing disc 22 here includes an annular ring 86 having a radial surface 50′. The annular ring 86 is, preferably but not necessarily, made of the same material as the stationary counter ring 26. Naturally, the annular ring 86 may be used in connection with any one of the earlier or following embodiments, i.e. the surface 50′ thereof includes grooves or not. The annular ring 86 may be fastened on the balancing disc 22 or in a groove arranged in the balancing disc. The annular ring 86 may be fastened to the balancing disc 22 by screws, adhesives, welding, riveting, just to name a few examples without any intention of limiting the options to the listed ones.
FIG. 10 illustrates a detailed axial cross section of the balancing device in accordance with a sixth preferred embodiment of the present invention. Here, in the sixth embodiment, the balancing disc 22 is disposed near its outer circumference with a number of circular rings 94 extending axially outwardly from the surface 50 of the balancing disc 22 and leaving annular grooves 94′ therebetween. Such circular rings may be disposed in at least one of the balancing disc 22, the counter member 28, the counter ring 26 and the annular ring 86 in the balancing disc 22. The grooves left between the circular rings are similar to that in the first embodiment. The circular rings 94 are, preferably but not necessarily, made of the same material as the stationary counter ring 26 and fastened to the balancing disc and the annular ring by screws, adhesives, welding, riveting, just to name a few examples without any intention of limiting the options to the listed ones.
FIG. 11 illustrates a detailed axial cross section of the balancing device in accordance with a seventh preferred embodiment of the present invention. Here, in this embodiment, the balancing disc includes circular rings and the counter ring with annular depressions into which the circular rings are extended. The circular rings and the walls of the depressions define therebetween the grooves of the present invention.
FIG. 12 illustrates a detailed axial cross section of the balancing device in accordance with an eighth preferred embodiment of the present invention. Here, in this embodiment, the balancing disc includes an annular ring and the counter member with the counter ring. Both the annular ring and the counter ring have annular grooves. What makes a difference in this embodiment is that at least one groove in the annular ring is facing at least one groove in the counter ring, whereby the cavity to which the liquid flows from the thin gap expands in both axial directions, and not only in one axial direction as shown in the other embodiments.
The operation principle of the above described balancing device will be explained in the following by referring mainly to FIGS. 2 and 3.
When any single suction centrifugal pump having one or more impellers is pumping liquid, the suction created by the impeller/s draws the impeller/s towards the pump inlet, i.e. creates thrust. The thus created thrust requires the use of thrust bearings that prevent the impeller(s) from making mechanical contact with the volute of the pump. Another way to prevent the mechanical contact is to arrange balancing device on the shaft of the impeller(s). The balancing device discussed in the present invention are mainly formed of a balancing disc fastened on the shaft for rotation therewith. The balancing device operate such that the pumped liquid is guided to the balancing cavity between the balancing disc and its stationary counter member, whereby the liquid pressure acting on the balancing disc tends to move the shaft away from the inlet of the pump, i.e. in a direction contrary to the thrust created by the impeller(s). The force the balancing disc is capable of creating is proportional to the radius of the balancing disc. In other words, the stronger force is needed the bigger the radius of the balancing disc should be. However, the power consumption of the balancing disc is also proportional to the radius of the balancing disc.
If the power consumption of the disc is to be reduced, the only way, in practice, is to decrease the diameter of the balancing disc. But, when the diameter of the balancing disc 22 (naturally together with the counter ring 26) is reduced without any other measures the pressure difference radially across the thin gap 36′, i.e. from the inner diameter of the counter ring 26 to the outer diameter thereof, is very high in relation to the cross sectional flow area in the gap 36′ between the counter ring 26 and the balancing disc 22. The high pressure difference across the thin gap results in the very high flow velocity at the entrance to the thin gap, whereby occasional local low pressure zones are formed in the gap so that the liquid is able to evaporate into vapor. While the liquid in the gap is evaporated, the disc loses its load carrying capability at least partially at the area of the counter ring 26 as the vapor escapes from the gap 36′ very quickly. The loss of load carrying capability allows the shaft 18 to move towards the pump inlet, whereby the non-axial counter surfaces 50 and 52 may end up into contact in substantially dry conditions. When the non-axial counter surfaces contact the power consumption increases rapidly, the friction heats the non-axial counter surfaces, and may at its worst damage the surfaces.
In other words, the reason for the loss of load carrying capability of the balancing disc is the combination of too high a pressure difference radially across the thin gap 36′ and too little resistance to flow in the gap area. When the behavior of the liquid in the thin gap 36′ between the counter ring 26 and the balancing disc 22 was considered in more detail, it was learned that the problem could be solved by adding resistance to flow in the gap 36′ at the counter ring area by dividing the total pressure difference into two or more partial pressure differences by arranging one or more annular grooves either in the balancing disc or in the annular ring, or in the counter member or in the counter ring, or in both between the inner and outer circumferences of the counter ring area. By arranging one or more annular grooves in the surfaces 50 and/or 52 the resistance to flow between the inner and outer circumference of the counter ring 26 is increased such that the liquid entering the thin gap 36′ between the counter ring 26 and the balancing disc 22 sees only the pressure difference between the entrance to the thin gap 36′ and the first groove 54, whereby the velocity of the liquid induced by that particular pressure difference is smaller than in such a case that the total pressure difference would act in the liquid. Thereby the local pressure in the thin gap 36′ is high enough for not allowing the liquid to evaporate. Furthermore, the flow velocity of the liquid in each groove is reduced close to nil, as the height or thickness of the flow cross section is suddenly changed from that of the thin gap 36′, i.e. from a micron range, to that in the groove, i.e. to a millimeter range. For the above reason the liquid pressure everywhere in the thin gap 36′ is able to carry some load, and as the pressure in the groove area is able to carry a substantial load, abrupt mechanical impacts of the balancing disc 22 to the counter ring 26 or counter member are prevented.
On the one hand, the fluctuation in the power consumption is a clear indication of occasional evaporation, whereas, on the other hand, the lack of fluctuations indicates the lack of evaporation.
In view of the above discussed various embodiments it should be understood that the details, i.e. grooves, circular rings, counter rings etc. thereof are interchangeable with those of any other embodiment whenever applicable. The same applies to the variations discussed in FIGS. 4 and 5. In other words, in the variations having inclined counter surfaces in both the radial or non-radial balancing disc and the counter member, the counter surfaces may have any imaginable combination of grooves or circular rings discussed in FIGS. 3, 6-12.
A feature that is in most cases, but not always, necessary for the working of the present invention is a throttling downstream of the balancing me device ans, i.e. a device that controls the liquid flow from the cavity downstream of the balancing disc further. Such a device may be a manual or otherwise controlled valve or a pipeline having a suitable cross sectional flow area for the throttling purpose. However, a preferable throttling device is the slide bearing shown in FIG. 2, for instance. In other words, as slide bearings, which now may be used, as the thrust of the impeller is balanced and there is thus no need for axial bearings, need liquid for lubrication purposes, it is quite practical to lead the liquid leaked via the balancing device to one or both shaft bearings of the centrifugal pump.
While the invention has been described herein by way of examples in connection with what are, at present, considered to be the most preferred embodiments, it is to be understood that the invention is not limited to the disclosed embodiments, but is intended to cover various combinations or modifications of its features, and several other applications included within the scope of the invention, as defined in the appended claims. The details mentioned in connection with any embodiment above may be used in connection with another embodiment when such combination is technically feasible.

Claims

1. A centrifugal pump, comprising:

a pump casing with an inlet and an outlet;

a shaft sealed and mounted with slide bearings to the pump casing, the shaft being movable in axial direction;

at least one impeller fastened on the shaft for rotation therewith and a balancing device configured to balance axial forces, the balancing device comprising a balancing disc fastened on the shaft for rotation therewith and having an outer circumference; and

a stationary counter member arranged in connection with the pump casing,

the balancing disc and the counter member forming therebetween a balancing cavity, the balancing disc having a first non-axial surface delimited to the outer circumference and the counter member having a second non-axial surface, the first and the second non-axial surfaces facing one another and forming a thin gap therebetween, at least one of the first and the second non-axial surfaces including at least one annular groove, a throttling being arranged downstream of the balancing disc and the counter member includes a counter ring having the second non-axial surface or the balancing disc includes an annular ring having the first non-axial surface.

2. (canceled)

3. (canceled)

4. The centrifugal pump in accordance with claim 1, wherein the at least one annular groove is a plurality of annular grooves, and at least one of the first and the second non-axial surfaces includes circular rings forming the plurality of annular grooves therebetween.

5. The centrifugal pump in accordance with claim 1, wherein the annular groove has a depth of at least 10-fold to that of the thin gap.

6. The centrifugal pump in accordance with claim 1, wherein the annular groove has a dimension of 1 to tens of millimeters in a direction at right angles to the first and the second non-axial surfaces.

7. The centrifugal pump in accordance with claim 1, wherein the annular groove has a dimension of 1 to tens of millimeters in a direction parallel with the first and the second non-axial surfaces.

8. The centrifugal pump in accordance with claim 1, wherein the counter member is a rear wall of the centrifugal pump.

9. The centrifugal pump in accordance with claim 1, wherein the at least one annular groove is in the balancing disc, and the at least one annular groove faces is facing a groove in the counter member.

10. The centrifugal pump in accordance with claim 1, wherein the throttling is arranged downstream of a cavity downstream of the balancing disc.

11. The centrifugal pump in accordance with claim 1, wherein the throttling downstream of the balancing device is one of a slide bearing, a valve and a pipeline having a suitable cross sectional flow area.

12. A method of balancing an axial thrust of a centrifugal pump in accordance with claim 1, the method comprising:

arranging the throttling downstream of the balancing disc for controlling the amount of liquid leaking via the balancing device,

providing the counter member with the counter ring having the second non-axial surface, and

providing the balancing disc with the annular ring having the first non-axial surface.

13. (canceled)