Classifications

• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
  • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
  • F04D29/00—Details, component parts, or accessories
  • F04D29/04—Shafts or bearings, or assemblies thereof
  • F04D29/041—Axial thrust balancing
  • F04D29/0413—Axial thrust balancing hydrostatic; hydrodynamic thrust bearings
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
  • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
  • F04D29/00—Details, component parts, or accessories
  • F04D29/04—Shafts or bearings, or assemblies thereof
  • F04D29/041—Axial thrust balancing
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
  • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
  • F04D29/00—Details, component parts, or accessories
  • F04D29/04—Shafts or bearings, or assemblies thereof
  • F04D29/041—Axial thrust balancing
  • F04D29/0416—Axial thrust balancing balancing pistons
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
  • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
  • F04D29/00—Details, component parts, or accessories
  • F04D29/04—Shafts or bearings, or assemblies thereof
  • F04D29/046—Bearings
  • F04D29/047—Bearings hydrostatic; hydrodynamic
  • F04D29/0473—Bearings hydrostatic; hydrodynamic for radial pumps
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
  • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
  • F04D29/00—Details, component parts, or accessories
  • F04D29/18—Rotors
  • F04D29/22—Rotors specially for centrifugal pumps
  • F04D29/2261—Rotors specially for centrifugal pumps with special measures
  • F04D29/2266—Rotors specially for centrifugal pumps with special measures for sealing or thrust balance
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
  • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
  • F04D1/00—Radial-flow pumps, e.g. centrifugal pumps; Helico-centrifugal pumps
  • F04D1/06—Multi-stage pumps
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F05—INDEXING SCHEMES RELATING TO ENGINES OR PUMPS IN VARIOUS SUBCLASSES OF CLASSES F01-F04
  • F05D—INDEXING SCHEME FOR ASPECTS RELATING TO NON-POSITIVE-DISPLACEMENT MACHINES OR ENGINES, GAS-TURBINES OR JET-PROPULSION PLANTS
  • F05D2250/00—Geometry
  • F05D2250/20—Three-dimensional
  • F05D2250/29—Three-dimensional machined; miscellaneous
  • F05D2250/294—Three-dimensional machined; miscellaneous grooved

Definitions

• the present invention relates to a centrifugal pump with balancing means in accordance with the preamble of claim 1 and a method of balancing axial forces of the centrifugal pump in accordance with the preamble of claim 12. More specifically the present invention relates to single- or multi-stage centrifugal pumps having novel disctype means for balancing the axial forces of the pump.
• the means for balancing axial forces of centrifugal pumps are normally in use in multistage pumps, which have a high pressure head, and are provided with 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.
• means for balancing the axial force have been developed.
• the disc-type balancing means may be considered as the preferred choice of the two basic balancing means as it adjusts its operation automatically, i.e. slight wear does not affect the operation of the balancing means at all, whereas even the slightest wear of the drum-type balancing means results in a change in the balancing capability of the balancing means. Furthermore, the disc-type balancing means occupies also less space in the axial direction than drum-type balancing means.
• the disc-type balancing means the present invention discusses later on in more detail is formed of a balancing disc fastened on the shaft of the pump and a stationary counter member.
• 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.
• 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.
• the balancing means 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 means 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.
• 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.
• the axial thrust loading the bearings of the pump is the difference of the two axial forces having opposite directions.
• 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.
• the first problem that may be seen in the operation of the balancing means is high power consumption combined with fluctuations in the power consumption due to the balancing means operating, alternatingly, in both low-friction and high-friction situations.
• an object of the present invention is to design such a novel balancing means for a centrifugal pump that reduces the power consumption of the balancing means.
• Another object of the present invention is to design such a novel balancing means for a centrifugal pump that considerably reduces or obviates totally the fluctuations in the power consumption of the balancing means.
• 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 means for a centrifugal pump that adjusts automatically its operating clearance.
• 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 impellerfastened on the shaft for rotation therewith and a means for balancing axial forcesf
• the balancing means 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 is being provided with at least one annular groove, and a throttling downstream of the balancing means
• the balancing disc is capable of carrying load for the entire radial extension thereof
• the balancing means may be easily adjusted for different pump properties, especially to varying head, and
• Fig. 1 illustrates schematically, and in an axial cross section, a multi-stage centrifugal pump including a disc-type balancing means in accordance with a first preferred embodiment of the present invention
• Fig. 2 illustrates schematically an axial, more detailed cross section of the balancing means in accordance with a first preferred embodiment of the present invention
• Fig. 3 illustrates a detailed axial cross section of the balancing means 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 means 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 means 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 means in accordance with a second preferred embodiment of the present invention
• Fig. 7 illustrates a detailed axial cross section of the balancing means in accordance with a third preferred embodiment of the present invention.
• Fig. 8 illustrates a detailed axial cross section of the balancing means in accordance with a fourth preferred embodiment of the present invention
• Fig. 9 illustrates a detailed axial cross section of the balancing means in accordance with a fifth preferred embodiment of the present invention.
• Fig. 10 illustrates a detailed axial cross section of the balancing means in accordance with a sixth preferred embodiment of the present invention
• Fig. 1 1 illustrates a detailed axial cross section of the balancing means in accordance with a seventh preferred embodiment of the present invention.
• Fig. 12 illustrates a detailed axial cross section of the balancing means in accordance with an eighth preferred embodiment of the present invention.
• 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 means 20.
• FIG. 2 illustrates schematically an axial, more detailed cross section of the balancing means 20 and the end part of the centrifugal pump in accordance with a first preferred embodiment of the present invention.
• the balancing means 20 is formed of a balancing disc 22 attached on the shaft 18 for rotation therewith.
• the balancing disc 22 there may be a separate sleeve or the balancing disc may be provided with 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 means 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 is, preferably but not necessarily, provided with 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 means of 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.
• 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 means 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.
• 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 means 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 means of 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.
• the cross sectional flow area, or in fact the radial depth thereof 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.
• 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 means require is the smallest.
• 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.
• the basic requirement for the direction of the counter surfaces 50 and 52 is that the direction thereof is non-axial.
• 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 means in accordance with a second variation of the first preferred embodiment of the present invention.
• the counter ring 26' is not radial but somewhat inclined as was discussed already above.
• the counter surface in the balancing disc 22' is inclined in a corresponding manner.
• 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 means in accordance with a third variation of the first preferred embodiment of the present invention.
• the counter ring 26" is not radial but somewhat inclined as was discussed already above.
• the counter surface in the balancing disc 22" is inclined in a corresponding manner.
• the entire balancing disc 22" is also inclined.
• the inclination of the machine elements discussed above may be needed for some constructional reasons. For instance, with the structure of Figure 5 the shaft bearing may be brought somewhat closer to the impeller.
• FIG. 6 illustrates a detailed axial cross section of the balancing means in accordance with a second preferred embodiment of the present invention.
• 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 means in accordance with a third preferred embodiment of the present invention.
• 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 means in accordance with a fourth preferred embodiment of the present invention.
• 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 means in accordance with a fifth preferred embodiment of the present invention.
• the balancing disc 22 is here provided with 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 provided with 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 means of 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 means in accordance with a sixth preferred embodiment of the present invention.
• the balancing disc 22 is provided 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 provided 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 means of screws, adhesives, welding, riveting, just to name a few examples without any intention of limiting the options to the listed ones.
• Fig. 1 1 illustrates a detailed axial cross section of the balancing means in accordance with a seventh preferred embodiment of the present invention.
• the balancing disc is provided with 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 means in accordance with an eighth preferred embodiment of the present invention.
• the balancing disc is provided with an annular ring and the counter member with the counter ring. Both the annular ring and the counter ring have annular grooves.
• 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 balancing means 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.
• the power consumption of the balancing disc is also proportional to the radius of the balancing disc. (031 ) 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.
• 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.
• 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.
• 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.
• 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.
• the local pressure in the thin gap 36' is high enough for not allowing the liquid to evaporate.
• 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.
• 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 means, i.e. a device that controls the liquid flow from the cavity downstream of the balancing disc further.
• a device may be a manual or otherwise controlled valve or a pipeline having a suitable cross sectional flow area for the throttling purpose.
• a preferable throttling device is the slide bearing shown in Figure 2, for instance.

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• Engineering & Computer Science (AREA)
• Mechanical Engineering (AREA)
• General Engineering & Computer Science (AREA)
• Physics & Mathematics (AREA)
• Fluid Mechanics (AREA)
• Structures Of Non-Positive Displacement Pumps (AREA)
• Control Of Non-Positive-Displacement Pumps (AREA)

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• 2017
  • 2017-03-06 BR BR112018068267A patent/BR112018068267A2/pt not_active Application Discontinuation
  • 2017-03-06 US US16/081,471 patent/US20190063450A1/en not_active Abandoned
  • 2017-03-06 EP EP17707915.9A patent/EP3430270B1/en active Active
  • 2017-03-06 RU RU2018132444A patent/RU2018132444A/ru not_active Application Discontinuation
  • 2017-03-06 CN CN201780014882.9A patent/CN108779782A/zh active Pending
  • 2017-03-06 WO PCT/EP2017/055118 patent/WO2017157702A1/en active Application Filing

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