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WO1998005890A1 - Ensemble joint d'etancheite/surface de roulement - Google Patents

Ensemble joint d'etancheite/surface de roulement Download PDF

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Publication number
WO1998005890A1
WO1998005890A1 PCT/CA1997/000548 CA9700548W WO9805890A1 WO 1998005890 A1 WO1998005890 A1 WO 1998005890A1 CA 9700548 W CA9700548 W CA 9700548W WO 9805890 A1 WO9805890 A1 WO 9805890A1
Authority
WO
WIPO (PCT)
Prior art keywords
interface
sleeve
exit
mouth
rotor
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Ceased
Application number
PCT/CA1997/000548
Other languages
English (en)
Inventor
Thomas Wilson Ramsay
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
Chesterton AW Co
Original Assignee
Chesterton AW Co
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Priority claimed from GBGB9616471.0A external-priority patent/GB9616471D0/en
Priority claimed from GBGB9618567.3A external-priority patent/GB9618567D0/en
Application filed by Chesterton AW Co filed Critical Chesterton AW Co
Priority to AU36906/97A priority Critical patent/AU3690697A/en
Publication of WO1998005890A1 publication Critical patent/WO1998005890A1/fr
Anticipated expiration legal-status Critical
Ceased legal-status Critical Current

Links

Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
    • F16C33/00—Parts of bearings; Special methods for making bearings or parts thereof
    • F16C33/02—Parts of sliding-contact bearings
    • F16C33/04—Brasses; Bushes; Linings
    • F16C33/06—Sliding surface mainly made of metal
    • F16C33/10—Construction relative to lubrication
    • F16C33/1025—Construction relative to lubrication with liquid, e.g. oil, as lubricant
    • F16C33/106—Details of distribution or circulation inside the bearings, e.g. details of the bearing surfaces to affect flow or pressure of the liquid
    • F16C33/107—Grooves for generating pressure
    • 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
    • F04D29/046—Bearings
    • F04D29/0467—Spherical 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
    • F04D29/046—Bearings
    • F04D29/047—Bearings hydrostatic; hydrodynamic
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
    • F16C17/00—Sliding-contact bearings for exclusively rotary movement
    • F16C17/10—Sliding-contact bearings for exclusively rotary movement for both radial and axial load
    • F16C17/102—Sliding-contact bearings for exclusively rotary movement for both radial and axial load with grooves in the bearing surface to generate hydrodynamic pressure
    • F16C17/105—Sliding-contact bearings for exclusively rotary movement for both radial and axial load with grooves in the bearing surface to generate hydrodynamic pressure with at least one bearing surface providing angular contact, e.g. conical or spherical bearing surfaces
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16J—PISTONS; CYLINDERS; SEALINGS
    • F16J15/00—Sealings
    • F16J15/16—Sealings between relatively-moving surfaces
    • F16J15/40—Sealings between relatively-moving surfaces by means of fluid
    • F16J15/406—Sealings between relatively-moving surfaces by means of fluid by at least one pump

Definitions

  • This invention relates to rotary seals and bearings, and is a development of the
  • the structures are characterised as 2 including a pair of sleeves, being a rotor-sleeve and a stator-sleeve, the sleeves having 3 respective interface surfaces; the rotor and stator interface surfaces are disposed in j male/female configuration; and the interface surfaces are so arranged as to define a 5 hydrodynamic-bearing-interface therebetween, during operation of the apparatus, and 6 upon supply of a liquid to the interface.
  • the interface is of a progressively-reducing-
  • One of the interface surfaces is provided with a groove, which extends spirally around
  • the groove has an entry-mouth and an exit-
  • the apparatus includes an entry-chamber, which is in liquid-flow-communication
  • the apparatus includes a supply of a liquid in the
  • the apparatus includes an exit-chamber, which is in liquid-flow-
  • two of the pairs of tapered sleeves may be
  • shaft can be contained between the two sleeve-pairs.
  • end-to-end configuration of sleeves can then serve as the whole bearing support required by the shaft.
  • the liquid or barrier-liquid provided for the hydrodynamic film at the sleeves interfaces is water, or water-based.
  • a 50/50 mixture of water with glycol has shown excellent performance.
  • Water, even when mixed with glycol, is not an ideal lubricant; that is to say, it does not form hydrodynamic films so readily as, say, lubricating oil.
  • a hydrodynamic film formed in water is fragile, compared with a film formed in oil; that is to say, it is all too easy to overload the surfaces defining the film, causing the film to break down, whereby the surfaces of the rotor and stator sleeves can touch, metal-to-metal.
  • the invention as described herein is aimed at providing a manner of arranging and profiling the sleeves, and the interactive surfaces thereof, whereby the possibility of a metal-to-metal contact occurrence is rendered less likely.
  • the interface At each location along the length of the interface, the interface has a respective cone angle, and, in respect of axially-spaced locations A and B along the axial length of the interface, the invention lies in providing that the cone angle of the interface at location A is substantially steeper than the cone angle at location B.
  • the interface has a respective film-thickness, being the thickness of the hydrodynamic film in the interface at that location, and, in respect of axially-spaced locations C and D along the axial length of the interface, the thickness of the film at location C is substantially thicker than the thickness of the film at location D.
  • the way in which this change in the steepness of the cone angle, and the change in the thickness of the film, is achieved is curving the profile of the rotor-sleeve / stator-sleeve interface.
  • Grooves in bearing surfaces are normally provided for purpose of equalising out a hydrostatic pressure in a lubricant.
  • the lubricant is being forced between the bearing surfaces, under pressure, and the grooves ensure the lubricant is spread evenly over the whole surface.
  • the pressure is at a maximum at the place where the lubricant enters the interface, and the pressure gradually drops off as the lubricant moves away from that location.
  • barrier-liquid is driven axially along the groove, by the rotation of the rotor, from near zero (gauge) pressure at the entry- mouth of the groove to a maximum pressure (in some cases, as high as 100 psi and more) at the exit-mouth of the groove.
  • the pressure of the barrier-liquid increases progressively (and, in some cases, substantially linearly) as the barrier-liquid passes through the bearing.
  • the designer preferably should seek to arrange, if possible, that in the portion of the interface where the pressure is lowest, the sleeve surfaces are a little further apart - or in other words, that the film is a little thicker in that portion.
  • the hydrodynamic-water-film interface between a pair of sleeves is supporting a heavy thrust force
  • the cone-angle of the interface is steep in one location and not-so-steep in another location, it may be surmised that the thrust forces are being supported more in the location where the cone is the steepest.
  • the designer should seek to provide, at the locations where the cone angle is steepest, that that is a location where the surfaces are a little further apart.
  • the properties of the interface from the standpoint of thrust- support, and the thickness of the film can be tailored by the designer to suit the circumstances.
  • Fig 1 is a cross-sectioned side elevation of a rotary impeller pump, which embodies the invention
  • Fig 2 is a similar view of another pump that embodies the invention
  • Fig 3a is a close up of a pair of sleeves from the pump of Fig 1
  • Fig 3b is the same view as Fig 3a, but shows the sleeves in a separated condition
  • Fig 4a is a close up of a pair of sleeves from the pump of Fig 2
  • Fig 4b is the same view as Fig 4a, but shows the sleeves in a separated condition
  • Fig 5 is a cross-sectioned side elevation of a rotary machine, which embodies the invention
  • Fig 6a is a close up of a pair of sleeves from the machine of Fig 5
  • Fig 6b is the same view as Fig 6a, but shows the sleeves in a separated condition
  • Fig 7 is a cross-sectioned side elevation of another rotary impeller pump, which embodies the invention
  • the pump 20 as shown in Fig 1 includes a rotor sleeve 23 mounted on a shaft 24, and a stator sleeve 25.
  • the outward-facing surface 26 of the male rotor-sleeve 23 and the inward-facing surface 27 of the female stator-sleeve 25 together define an interface.
  • a spiral groove 28 is provided in the male rotor surface 26, the groove having an entry- mouth 29 and an exit-mouth 30.
  • the entry-mouth 29 communicates with an entry- chamber 32, and the exit-mouth 30 communicates with an exit-chamber 34.
  • the surface 26 contains only the one groove 28, the groove having several turns.
  • the hand of the spiral groove 28 is such that upon rotation of the shaft 24, liquid present in the groove 28 travels along the groove from left to right in Fig 1 . Barrier-liquid is supplied at ambient pressure to the entry-chamber 32 from a reservoir 35.
  • the exit-chamber 34 is open to the process-chamber 36. In other cases, the exit-chamber can be sealed from the process-chamber, as will be observed in some of the accompanying drawings.
  • the shaft 24 is provided with bearings shown diagrammatically at 37.
  • the interface between the surfaces 26,27 serves as a seal, for sealing the process fluid from the atmosphere.
  • the female stator-sleeve 25 is spring- (and pressure-) biassed into contact with the male rotor-sleeve 23. Even though the interaction of the sleeves 23,25 is intended as a seal, not as a bearing, of course the sleeves do provide some support for the shaft.
  • the grooved male rotor surface 26 is curved convexly as shown in Fig 1. In Fig 2, the situation is similar, except that the grooved male rotor sleeve is curved concavely.
  • Fig 3a shows the sleeves 23,25 of Fig 1.
  • the cone angle of the interface at location A is steeper than the cone angle at location B.
  • the cone angle at A is substantially more than ten degrees (included) and the cone angle at B is less than ten degrees.
  • Fig 3b shows the sleeves somewhat separated. Although Fig 3 exaggerates the effect of axially separating the sleeves, it will be noted that the thickness of the film at axial location C increases more than the thickness of the film at location D, as the sleeves are separated. Figs 4a and 4b show the same thing in respect of Fig 2.
  • Fig 5 shows a machine with bearings supporting a shaft.
  • two pairs of rotor- and stator-sleeve are shown, in an in-series end-to-end or back-to-back configuration.
  • the sleeves define a spherical interface, which allows the shaft to be in angular- misalignment with respect to the frame of the machine (in the manner of conventional spherical bearings).
  • Figs 6a and 6b illustrate the progressive nature of the changes in film thickness as the sleeves of Fig 5 are axially separated.
  • Fig 7 shows a pump, having an impeller 40 operating in a process-fluid-chamber 42.
  • the impeller is driven by a shaft 43, which takes torque from a motor (not shown) to the left.
  • the shaft 43 receives only torque from the motor, and the shaft has no other bearings, other than the sleeves as shown, which support the shaft against journal (radial) and thrust (axial) forces.
  • the rotor sleeves 45L,45R are solid with the shaft 43.
  • the stator sleeves 47L.47R are solid with the housing 49.
  • Mechanical seal 50 seals the process-chamber 42 from the sleeves, and mechanical seal 52 seals the sleeves from the atmosphere.
  • the sleeves are arranged in end-to-end configuration.
  • the sleeves are centre- fed, in that barrier-liquid is admitted (substantially not under pressure) into a chamber 53, which serves as the entry chamber for both sleeve-pairs 48L (45L+47L) and 48R (45R+47R) for conveying liquid to the entry-mouths of the grooves.
  • the spiral grooves 54L.54R are oppositely handed, whereby the liquid is forced to the left in the left sleeve- pair 48L, and to the right in the right sleeve-pair 48R, during rotation. From the exit- mouths of the grooves, the liquid emerges, now under pressure, into the respective exit- chambers 56L.56R.
  • Fig 7a shows a similar arrangement of sleeves, but now the barrier-liquid is end-fed. That is to say, the liquid is supplied to an entry-chamber 57 at the extreme left, passes along the interface of the first sleeve-pair 48L, progressively gaining pressure, and emerges into the intermediate chamber 58. From there, the liquid passes along the interface of the second sleeve-pair 48R, gaining more pressure, and emerges into the exit-chamber 59.
  • the intermediate chamber 58 may be regarded as the exit-chamber for the first sleeve-pair 48L and as the entry-chamber for the second sleeve-pair 48R.
  • cavitation refers to the condition where the impeller is rotating but there is an interruption in the supply of process fluid in the intake to the impeller. Cavitation can be partial or (rarely) complete. When cavitation is partial, the thrust force due to impeller-draw is reduced: however, in practice, cavitation can cause vibrations, and spurious effects, whereby the thrust force might even be (momentarily) increased. If cavitation is complete, the impeller-induced thrust-force on the shaft theoretically drops to zero.
  • the shaft in Fig 7, the centre-fed case, the shaft is (nearly) neutrally- thrust-balanced under complete-cavitation conditions, i.e when there is no process fluid in the intake to the impeller, but the shaft is urged to the right by impeller-draw under normal pumping; whereas in Fig 7a, the end-fed case, the shaft is (nearly) neutrally- thrust-balanced under normal pumping, but suffers a thrust force to the left, because of the pressure in the chambers, under cavitation.
  • the designer can choose to feed the barrier-liquid into the sleeves according to which mode of operation best suits the particular installation.
  • the film in the interface of the right sleeve-pair 48R which under normal pumping is the sleeve-pair that is reacting the thrust force due to impeller-draw, is thicker near the impeller, where the cone angle of the interface is steeper, and the film is thinner at the other end of the right sleeve-pair 48R, where the cone angle is less steep.
  • the incremental increase in pressure per incremental change in axial location is inversely proportional to the film thickness at the location. That is to say, the pressure increases in the shortest distance where the film is thinnest. The thicker the film, the less the pressure increases with length; the thinner the film, the more the pressure increases with length. So, for an interface that is to take thrust, the designer should see to it that the film is at the higher pressure at the location where the film is predominantly taking the thrust, and also that the film is thickest at the location where the film is predominantly taking the thrust.
  • the left sleeve-pair 48L as shown in Fig 7a does not meet this condition.
  • the exit end of the left sleeve-pair is the intermediate chamber 58. That is where the pressure in the left sleeve-pair interface is higher, and that is where the film should be thicker, and the slope should be steeper.
  • the shape of the curvature of the left-sleeve should rather be designed as shown in Fig 8.
  • Fig 8a shows a version that follows the second design rule.
  • Fig 7a For the left sleeve pair in Fig 7a (end-fed), which comes under increasing thrust load with increasing cavitation of the pump, according to the second rule, the configuration shown in Fig 7a, i.e the rotor-is-convex configuration, would be preferred over that of Fig 8. Again, it does not matter so much what shape the right sleeve-pair has, since that is not under thrust during cavitation, and is not under so much thrust, even under normal pumping. Besides, in Fig 7a, the entry pressure of the liquid fed to the entry mouth of the groove of the right sleeve-pair has already been elevated in pressure by the action of passing through the left sleeve-pair.
  • cavitation is a common bane of pumps; it is all too usual that if a pump fails, it fails because of what happens during cavitation. In a case where that is so, the centre-fed mode should be favoured. If the cavitation is complete, and there is no process fluid at all in the impeller intake, the shaft being then under neutral thrust, both sleeve-pairs are providing only a minimum of pressure, and the second rule assumes more importance. In that case, the thickest films should be at the entry ends of the sleeve-pairs; therefore, both sleeve-pairs should be rotor-concave (Fig 8a).
  • the normal-pumping thrust force due to impeller-draw is reacted in the right hand sleeve-pair, i.e in the sleeve-pair nearest the impeller.
  • This is preferred, in that the thrust-loaded sleeve-pair has a smaller running clearance than the thrust-unloaded sleeve-pair.
  • the left sleeve-pair is the one that takes the thrust force on the shaft due to impeller-draw.
  • the left and right sleeve- pairs are arranged end-to-end in "bow-tie" configuration, as shown in Fig 9.
  • the end- to-end configuration shown in Figs 7,7a,8,8a may be termed "football" configuration, by contrast.
  • Fig 9a also shows a variety of bow-tie configuration.
  • Figs 10a, 10b, 10c, 10d show diagrammatically the arranged convexity / concavity of the curvatures of the sleeve-pair interfaces, in following the first design rule, i.e the cone angle should be steepest, and the film should be thickest, at the high-pressure end of the sleeve-pair.
  • Fig 10a shows football / end-fed;
  • Fig 10b shows football / centre-fed;
  • Fig 10c shows bow-tie / end-fed; and
  • Fig 10d shows bow-tie / centre-fed.
  • Fig 1 1 a shows such an arrangement, which corresponds to Fig 8.
  • Fig 1 1 b shows an arrangement that corresponds to Fig 7.
  • Fig 1 1 c shows the thicker film at the entry ends of both sleeves, but the steeper slope is at the exit end. This arrangement might be favoured in some cases, especially since it can in this case be arranged that, if the sleeves were to come together in metal-to-metal contact, they would touch first at the exit end, where the pressure is higher.
  • the curvatures i.e the cone-angles
  • the curvature can be set so that the film thickness varies from thickest-at-the-low-pressure-end-when- the-fiim-is-thin, to thicker-at-the-high-pressure-end-when-the-film-is-thick, within the same sleeve-pair.
  • Fig 12a shows a sleeve-pair in which the curvature changes between rotor-convex and rotor-concave, within the one interface. It may be noted that, in the absence of barrier- liquid, the interface surfaces would come together simultaneously all-over, and yet, as shown in Fig 12b, the configuration is very favourable for handling thrust on a sleeve- pair.
  • the film thickness is wide at the entry, or low-pressure, left end region of the interface, so that there is plenty of opportunity for the liquid to establish itself into the entry-end of the interface even when contact is imminent. That is to say, the interface is not starved or cut off from the supply of liquid in the entry chamber.
  • the liquid then passes to the central region of the interface, where the clearance is small, and therefore where the pressure rises rapidly as the liquid flows from left to right.
  • the pressure in the film is high, the film is thick, and the cone angle is steep, which, as mentioned, are excellent conditions for supporting heavy thrust forces.
  • Fig 13 shows a compressor, in which the big-end bearings on the crank- pin are curve-profiled.
  • the spherical configuration shown here is advantageous, because it permits the connecting rod of the machine to exactly follow the line of action of the piston in the cylinder, even if the cylinder might not be quite square to the axis of the crank-pin.
  • the arrows on the drawing indicate the direction of flow of the lubricant through the various zones.
  • Fig 14 shows curved-profiling of the main-bearings, and a rotor-concave profile for the big-end bearings.

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  • Engineering & Computer Science (AREA)
  • General Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • Physics & Mathematics (AREA)
  • Fluid Mechanics (AREA)
  • Chemical & Material Sciences (AREA)
  • Oil, Petroleum & Natural Gas (AREA)
  • Structures Of Non-Positive Displacement Pumps (AREA)

Abstract

On a monté des manchons de rotor et de stator sur l'arbre, par exemple, d'une pompe à roue-hélice, afin d'étanchéifier l'arbre et constituer des surfaces de roulement pour celui-ci. Le manchon du rotor présente une rainure en spirale destinée à produire une pression dans le liquide formant une barrière. L'interface du manchon présente une configuration à diamètre se réduisant, et le profil de l'interface est courbe. Cette courbure a pour effet de changer l'épaisseur du film liquide et l'angle du cône de l'interface, selon des proportions différentes, lorsque les manchons se déplacent ensemble et s'éloignent l'un de l'autre, de manière axiale. Grâce à cette conception, il est ainsi possible de repousser la survenue d'une rupture hydrodynamique du film, dans bien des conditions de fonctionnement.
PCT/CA1997/000548 1996-08-05 1997-08-05 Ensemble joint d'etancheite/surface de roulement Ceased WO1998005890A1 (fr)

Priority Applications (1)

Application Number Priority Date Filing Date Title
AU36906/97A AU3690697A (en) 1996-08-05 1997-08-05 Seal/bearing assembly

Applications Claiming Priority (4)

Application Number Priority Date Filing Date Title
GB9616471.0 1996-08-05
GBGB9616471.0A GB9616471D0 (en) 1996-08-05 1996-08-05 Rotary bearing and/or seal device
GB9618567.3 1996-09-05
GBGB9618567.3A GB9618567D0 (en) 1996-09-05 1996-09-05 Rotary seal assembly

Publications (1)

Publication Number Publication Date
WO1998005890A1 true WO1998005890A1 (fr) 1998-02-12

Family

ID=26309822

Family Applications (1)

Application Number Title Priority Date Filing Date
PCT/CA1997/000548 Ceased WO1998005890A1 (fr) 1996-08-05 1997-08-05 Ensemble joint d'etancheite/surface de roulement

Country Status (2)

Country Link
AU (1) AU3690697A (fr)
WO (1) WO1998005890A1 (fr)

Cited By (15)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
GB2333331A (en) * 1997-12-31 1999-07-21 Lewis Patrick A marine propeller shaft seal
EP1213517A3 (fr) * 2000-12-05 2002-12-18 Emu Unterwasserpumpen Gmbh Joint mécanique pour machines rotatives
WO2003058098A1 (fr) * 2002-01-14 2003-07-17 Aes Engineering Ltd Dispositif d'ejection
WO2009056242A1 (fr) * 2007-10-29 2009-05-07 Grundfos Management A/S Ensemble pompe
WO2014007971A1 (fr) * 2012-06-14 2014-01-09 Inpro/Seal Llc Ensemble joints d'arbres
US8979093B2 (en) 2002-06-21 2015-03-17 Inpro/Seal, LLC Pressure balanced shaft seal assembly
US9004491B2 (en) 2002-06-21 2015-04-14 Inpro/Seal Llc Shaft seal assembly
US9048603B2 (en) 2011-04-18 2015-06-02 Inpro/Seal, LLC Current diverter ring
US9071092B2 (en) 2005-06-25 2015-06-30 Inpro/Seal, LLC Current diverter ring
US9525327B2 (en) 2011-04-18 2016-12-20 Inpro/Seal, LLC Current diverter ring
US9634547B1 (en) 2005-06-25 2017-04-25 Inpro/Seal Llc Motor grounding seal
US9831739B2 (en) 2012-06-18 2017-11-28 Inpro/Seal Llc Explosion-proof current diverting device
WO2018140731A1 (fr) 2017-01-27 2018-08-02 Regal Beloit America, Inc. Ensembles pompes centrifuges ayant un moteur électrique à flux axial et leurs procédés d'assemblage
US11543031B2 (en) 2018-09-28 2023-01-03 Inpro/Seal Llc Shaft seal assembly
US12163590B2 (en) 2013-09-30 2024-12-10 Inpro/Seal Llc Shaft seal assembly

Citations (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US1132759A (en) 1914-03-17 1915-03-23 Graphite Lubricating Company Floating bearing.
DE3326415A1 (de) * 1983-07-19 1985-02-07 Mannesmann AG, 4000 Düsseldorf Axial-radiallager, insbesondere fuer schiffsantriebsanlagen
US4614445A (en) 1983-11-08 1986-09-30 U.S. Philips Corporation Metal-lubricated helical-groove bearing comprising an anti-wetting layer
WO1995035457A1 (fr) 1994-06-20 1995-12-28 Ramsay Thomas W Ensemble joint d'etancheite/surface de roulement
WO1997013084A1 (fr) 1995-10-02 1997-04-10 A.W. Chesterton Co. Joint d'etancheite et/ou palier rotatif

Patent Citations (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US1132759A (en) 1914-03-17 1915-03-23 Graphite Lubricating Company Floating bearing.
DE3326415A1 (de) * 1983-07-19 1985-02-07 Mannesmann AG, 4000 Düsseldorf Axial-radiallager, insbesondere fuer schiffsantriebsanlagen
US4614445A (en) 1983-11-08 1986-09-30 U.S. Philips Corporation Metal-lubricated helical-groove bearing comprising an anti-wetting layer
WO1995035457A1 (fr) 1994-06-20 1995-12-28 Ramsay Thomas W Ensemble joint d'etancheite/surface de roulement
WO1997013084A1 (fr) 1995-10-02 1997-04-10 A.W. Chesterton Co. Joint d'etancheite et/ou palier rotatif

Cited By (22)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
GB2333331A (en) * 1997-12-31 1999-07-21 Lewis Patrick A marine propeller shaft seal
EP1213517A3 (fr) * 2000-12-05 2002-12-18 Emu Unterwasserpumpen Gmbh Joint mécanique pour machines rotatives
US6612805B2 (en) 2000-12-05 2003-09-02 Emu Unterwasserpumpen Gmbh Hydrodynamic machine
WO2003058098A1 (fr) * 2002-01-14 2003-07-17 Aes Engineering Ltd Dispositif d'ejection
US8979093B2 (en) 2002-06-21 2015-03-17 Inpro/Seal, LLC Pressure balanced shaft seal assembly
US9004491B2 (en) 2002-06-21 2015-04-14 Inpro/Seal Llc Shaft seal assembly
US9634547B1 (en) 2005-06-25 2017-04-25 Inpro/Seal Llc Motor grounding seal
US10270320B1 (en) 2005-06-25 2019-04-23 Inpro/Seal Llc Motor grounding seal
US9071092B2 (en) 2005-06-25 2015-06-30 Inpro/Seal, LLC Current diverter ring
WO2009056242A1 (fr) * 2007-10-29 2009-05-07 Grundfos Management A/S Ensemble pompe
EP2063139A1 (fr) * 2007-10-29 2009-05-27 Grundfos Management A/S Groupe motopompe
US8807920B2 (en) 2007-10-29 2014-08-19 Grundfos Management A/S Pump assembly
US9048603B2 (en) 2011-04-18 2015-06-02 Inpro/Seal, LLC Current diverter ring
US9614339B2 (en) 2011-04-18 2017-04-04 Inpro/Seal, LLC Current diverter ring
US9525327B2 (en) 2011-04-18 2016-12-20 Inpro/Seal, LLC Current diverter ring
WO2014007971A1 (fr) * 2012-06-14 2014-01-09 Inpro/Seal Llc Ensemble joints d'arbres
US9831739B2 (en) 2012-06-18 2017-11-28 Inpro/Seal Llc Explosion-proof current diverting device
US12163590B2 (en) 2013-09-30 2024-12-10 Inpro/Seal Llc Shaft seal assembly
WO2018140731A1 (fr) 2017-01-27 2018-08-02 Regal Beloit America, Inc. Ensembles pompes centrifuges ayant un moteur électrique à flux axial et leurs procédés d'assemblage
EP3574220A4 (fr) * 2017-01-27 2020-12-02 Regal Beloit America, Inc. Ensembles pompes centrifuges ayant un moteur électrique à flux axial et leurs procédés d'assemblage
US11543031B2 (en) 2018-09-28 2023-01-03 Inpro/Seal Llc Shaft seal assembly
US12000484B2 (en) 2018-09-28 2024-06-04 Inpro/Seal Llc Shaft seal assembly

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