EP4638963A1 - Expansion de volume pour une réduction de cavitation dans un engrènement de pompe à engrenages - Google Patents

Expansion de volume pour une réduction de cavitation dans un engrènement de pompe à engrenages

Info

Publication number
EP4638963A1
EP4638963A1 EP23844402.0A EP23844402A EP4638963A1 EP 4638963 A1 EP4638963 A1 EP 4638963A1 EP 23844402 A EP23844402 A EP 23844402A EP 4638963 A1 EP4638963 A1 EP 4638963A1
Authority
EP
European Patent Office
Prior art keywords
gear
root
length
teeth
bearing
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.)
Pending
Application number
EP23844402.0A
Other languages
German (de)
English (en)
Inventor
Thomas L. BURNHAM
William H. Dalton
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.)
Triumph Engine Control Systems LLC
Original Assignee
Riumph Engine Control Systems LLC
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 US18/372,480 external-priority patent/US12486843B2/en
Application filed by Riumph Engine Control Systems LLC filed Critical Riumph Engine Control Systems LLC
Publication of EP4638963A1 publication Critical patent/EP4638963A1/fr
Pending legal-status Critical Current

Links

Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C15/00—Component parts, details or accessories of machines, pumps or pumping installations, not provided for in groups F04C2/00 - F04C14/00
    • F04C15/0042—Systems for the equilibration of forces acting on the machines or pump
    • F04C15/0049—Equalization of pressure pulses
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C2/00—Rotary-piston machines or pumps
    • F04C2/08—Rotary-piston machines or pumps of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
    • F04C2/082—Details specially related to intermeshing engagement type machines or pumps
    • F04C2/084—Toothed wheels
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C2/00—Rotary-piston machines or pumps
    • F04C2/08—Rotary-piston machines or pumps of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
    • F04C2/082—Details specially related to intermeshing engagement type machines or pumps
    • F04C2/088—Elements in the toothed wheels or the carter for relieving the pressure of fluid imprisoned in the zones of engagement

Definitions

  • the present disclosure relates generally to gear systems and, more particularly, to cavitation between meshing gear teeth.
  • All pumps operate by creating a low pressure at the inlet and allowing atmospheric (or system) pressure to push fluid into the pump.
  • the process makes all pumps susceptible to cavitation.
  • Cavitation is the formation of vapor cavities (bubbles) inside of a liquid when the local pressure is decreased rapidly below the vapor pressure of the liquid. This pressure decrease forms a vapor bubble inside the liquid which typically lasts for a short time before collapsing back into a liquid.
  • the collapse is violent, producing a loud popping noise and shockwaves that can damage nearby surfaces. These shockwaves can cause wear in some mechanical components.
  • Vapor cavities that implode near solid surfaces can cause cyclic stresses through repeated exposure to such implosions. Even strong metals will be pitted when subjected to the strong, localized jet resulting from the bubble implosion. Left unchecked, the damage can eventually destroy the pump.
  • Fig. 1 is a cross section side view schematic diagram.
  • Fig. 1 shows a conventional gear pump 10 having two identical spur gears 2 and 3 located within a pump housing 6.
  • Each gear 2, 3 has a plurality of gear teeth 1 (nine teeth 1 are shown for each gear 2, 3 as an example) arrayed about its outer periphery’.
  • the first gear 2 is mounted on, and integrally supported by, a corresponding gear shaft 4 so that the first gear 2 rotates in a clockwise direction 2a.
  • the second gear 3 is mounted on, and integrally supported by, a corresponding gear shaft 5 so that the second gear 3 rotates in a counterclockwise direction 3a.
  • a motor (not shown) causes the gears 2, 3 to rotate.
  • the gears 2, 3 continually trap portions of the fluid F at one location and displace those portions to another location, thereby pumping the fluid F.
  • Fig. 1A illustrates one common cavitation location in the gear pump 10. w here the teeth 1 of the first gear 2 approach the walls of the housing 6.
  • Fig. IB illustrates another common cavitation location in the gear pump 10, namely, proximate the mesh location 7. The cavitation that can occur at the mesh location 7 is explained more fully as follows.
  • the fluid F to be pumped is draw n into the inlet opening 8 by the gears 2, 3 coming out of mesh at a location relatively near to the inlet opening 8.
  • an expanding inter-tooth volume forms between the adjacent teeth 1 on each gear 2, 3 as the formerly meshed tooth 1 of the other gear exits those spaces.
  • These inter-tooth volumes in the spaces between adj acent teeth 1 on the gear coming out of mesh are filled by fluid F from the inlet opening 8 and, as indicated above, forced to move with each gear 2, 3 between its teeth 1 along the closely adjacent interior surface of the outer wall of the housing 6 to the outlet opening 9 at the discharge side of the gear pump 10.
  • fluid F is forced to exit the outlet opening 9 by gears 2, 3 going into mesh at a location relatively near to the outlet opening 9 to form shrinking inter-tooth volumes between those adjacent teeth 1 on each gear 2, 3 resulting from corresponding teeth 1 of the other gear entering those spaces.
  • the fluid discharge pressure is predominantly determined by the downstream conduit passageway cross sectional areas.
  • the meshing of the teeth 1 of the gears 2, 3 at the mesh location 7 which is more or less along an axis there joining the axes of symmetry' of the gear shafts 4, 5, and the presence of closely adjacent flat bearing surface portions there, has the effect of isolating the fluid F at the outlet opening 9 from that at the inlet opening 8.
  • Cavitation can occur in the gear pump 10 on the intake side of the gear pump 10 in the region where the teeth 1 of the gears 2, 3 separate in coming out of mesh with one another.
  • the expanding inter-tooth volume between adjacent teeth 1 on each gear 2, 3, where the tooth 1 of the other gear had just been and is exiting must be filled by the fluid F to be pumped that is entering from the inlet opening 8 under whatever is the inlet fluid pressure.
  • the rate of the expanding inter-tooth volumes can exceed the rate such volumes can be filled by the incoming fluid F at the inlet opening 8 under the inlet fluid pressure.
  • the local fluid pressure decreases below the vapor pressures of dissolved gases in the fluid F, or the vapor pressure of the pumped fluid F itself, so as to rupture the continuity of the fluid F at some particle or solid surface nucleation site and thereby form a cavity or bubble.
  • gases, or the vapors of the fluid F, or both evaporate into that cavity from the surrounding fluid medium.
  • Fig. 2 illustrates a computational fluid dynamics (CFD) analysis of a gear pump like the gear pump 10 done by CFX Berlin Software GmbH of Berlin, Germany.
  • the gears 2. 3 of the gear pump 10 each have nine teeth 1.
  • the gears of the gear pump analyzed in Fig. 2 each have eleven teeth.
  • CFD simulates fluid motion using numerical approaches; therefore, CFD acts as a virtual fluid dynamics simulator. The results show velocities, water vapor (cavitation), and air (aeration) behavior.
  • a pressurized fluid passageway is provided in at least one of the bearing structures across from the meshing region and extending between surface openings at the bearing surface of that bearing structure that are positioned on opposite sides of an alignment axis in that bearing surface. The surface openings are separated from one another by at least the width of the tooth 1 provided on the pair of gears 2, 3.
  • cavitation can accelerate the wear and reduce the pumping efficiency and lifespan of gear pump components, particularly gear teeth. Therefore, objects of the present disclosure are to minimize, if not eliminate, cavitation; reduce wear; and increase the efficiency and lifespan of gear pumps.
  • a related object is to significantly reduce or even to eliminate the pressure difference at the limits of the gear pump.
  • Another object is to provide a gear pump that minimizes fuel consumption by reducing energy losses.
  • the present disclosure provides a gear having volume expansion for cavitation reduction in a gear pump mesh.
  • the gear has a gear tooth profile; a body; a plurality of involute gear teeth extending radially outward from the body and including first and second neighboring gear teeth each having a respective tip and a root, the first and second neighboring gear teeth defining a space between them; and a root pocket formed directly into the roots of the gear teeth and in the space between the gear teeth, providing an increased gear root volume and adding trapped fluid compliance while leaving unaltered the gear tooth profile.
  • the gear pump comprises a first gear having (i) a gear tooth profile, (ii) a first body, (iii) a first plurality of involute gear teeth extending radially outward from the first body and including first and second neighboring gear teeth each having a respective tip and a root, the first and second neighboring gear teeth defining a space between them, and (iv) a root pocket formed directly into the roots of the first gear teeth and in the space betw een the first gear teeth, providing an increased gear root volume and adding trapped fluid compliance while leaving unaltered the gear tooth profile.
  • the gear pump also comprises a second gear being configured to mesh with the first gear in a mesh zone that defines a gear mesh volume that is increased by the root pocket and having a second body and a second plurality of involute gear teeth extending radially outw ard from the second body.
  • the gear pump further comprises a first gear bearing and a second gear bearing configured to position the first gear and the second gear along a bearing center hne, wherein the root pocket does not extend into either the first or the second gear bearing.
  • the gear pump still further comprises a housing having a fluid inlet and a fluid outlet and in which are disposed the first gear, the second gear, first gear bearing, and the second gear bearing.
  • Fig. 1 is a cross section side view schematic diagram showing a conventional gear pump
  • FIG. 1 A illustrates one common cavitation location in the gear pump shown in Fig. 1, where the teeth of the first gear approach the walls of the housing of the gear pump, in an enlarged view of the dashed area 1A shown in Fig. 1;
  • Fig. IB illustrates another common cavitation location in the gear pump shown in Fig. 1, namely proximate the mesh location, in an enlarged view of the dashed area IB shown in Fig. 1;
  • Fig. 2 illustrates a computational fluid dynamics (CFD) analysis of a gear pump like the conventional gear pump shown in Fig. 1;
  • FIG. 3 is a perspective view of a conventional high pressure gear pump illustrating the general layout of the gear pump
  • FIG. 4 is a perspective view of one of the bearings of the high pressure gear pump shown in Fig. 3;
  • FIG. 5 is a front view of one of the mobile or floating bearings of the high pressure gear pump shown in Figs. 3 and 4;
  • Fig. 6A show s a front offset angle perspective view of an example gear pump assembly
  • Fig. 6B is an enlarged view 7 of the bearing dam half that is highlighted in the dashed area 6B shown in Fig. 6A;
  • Fig. 7 is a side view of a collection of example gear teeth
  • Fig. 8 is side view schematic diagram showing a gear as modified by machining cavities or pockets directly into the roots of the gear teeth according to the present disclosure
  • Fig. 9 is a schematic diagram illustrating an example, oval shape and example dimensions for the pockets shown in Fig. 8;
  • Fig. 10 is a perspective view of a portion of a gear pump, including the gear having the pockets shown in Fig. 8, located between a stationary or fixed bearing and a floating or mobile bearing;
  • Fig. 11 is a perspective view of a portion of a gear having a pocket with a beam across the center of the pocket;
  • Fig. 12 is a schematic diagram illustrating the conventional bearing design used in the Model MFP-590 main fuel pump
  • Fig. 13 illustrates an analysis of trapped volume upon gear meshing for the conventional beanng design shown in Fig. 12;
  • Fig. 14 is a schematic diagram illustrating an aggressive bearing design that was analyzed and tested in the Model MFP-590 pump
  • Fig. 15 illustrates an analysis of trapped volume upon gear meshing for the aggressive bearing design shown in Fig. 14;
  • Fig. 16 is a schematic diagram illustrating a trapped volume analysis for a specific fluid fuel traveling through the aggressive bearing design shown in Fig. 14;
  • Fig. 17 is a schematic diagram illustrating a trapped volume analysis for the specific fluid fuel traveling through the conventional bearing design shown in Fig. 12;
  • Fig. 18A reflects calculated trapped volumes taken at an angle of rotation of the gears of 10° for the aggressive bearing design shown in Fig. 14;
  • Fig. 18B reflects calculated trapped volumes taken at an angle of rotation of the gears of 11° for the aggressive bearing design shown in Fig. 14;
  • Fig. 19A is a graph of measured pressure versus time for a MFP-590 pump with only the aggressive bearing design shown in Fig. 14, depicting the discharge pressure ripple;
  • Fig. 19B is a graph of measured pressure versus time for a MFP-590 pump w ith the aggressive bearing design shown in Fig. 14 and the pockets shown in Fig. 8, depicting the discharge pressure ripple.
  • the term “about” means those amounts, sizes, formulations, parameters, and other quantities and characteristics are not and need not be exact, but may be approximate and/or larger or smaller, as desired, reflecting tolerances, conversion factors, rounding off, measurement error and the like, and other factors known to those of skill in the art.
  • a value is described to be about or about equal to a certain number, the value is within ⁇ 10% of the number.
  • a value that is about 10 refers to a value between 9 and 11, inclusive.
  • the term “about” is used in describing a value or an end-point of a range, the disclosure should be understood to include the specific value or end-point.
  • indefinite article “a” or “an” and its corresponding definite article “the” as used in this disclosure means at least one, or one or more, unless specified otherwise.
  • “Include.” “includes.” “including,” “have.” “has,” “having,” comprise,” “comprises,” “comprising,” or like terms mean encompassing but not limited to, that is, inclusive and not exclusive.
  • Fig. 3 is a perspective view of the gear pump 10 typical of an aerospace fluid pump that operates to pump fuel, lubricant, or other fluids.
  • Aircraft engines include a main fuel pump, for example, that is at the heart of their regulation system.
  • Such pumps supply fuel to the combustion chamber by pumping the necessary flow from the fuel tanks.
  • the output flow from these fuel pumps is also used as a hydraulic fluid to operate actuators, like those used to open air flow discharge gates from the engine core flow to the fan flow.
  • Fig. 3 illustrates the general layout of the high pressure gear pump 10, which is described in U.S. Patent No. 10,094,291 assigned to Safran Aircraft Engines of Paris, France.
  • the gear pump 10 includes the first gear 2 and the second gear 3 that mesh together and discharge fuel between their teeth 1 to achieve pumping.
  • the first gear 2 is driven by the gear shaft 4 (see Fig. 1).
  • each of the gears (or pinions) 2 and 3 have stub shafts 13, 14 and 15, 16 at its two opposite sides, the first of which at the right in Fig. 3 are supported by first bearings 17, 18 respectively, called fixed or stationary bearings with first gaps, and the second at the left in Fig.
  • second bearings called mobile or floating bearings 19, 20 with second gaps.
  • These bearings 17, 18, 19, and 20 are all smooth bearings, but the fixed bearings 17 and 18 are retained with a smaller gap in the housing 6 than the mobile bearings 19 and 20, and can thus be displaced in the axial direction to squeeze the gears 2 and 3 and reduce gaps that could enable recirculation of the pumped fluid towards low pressures.
  • the high pressure dish 33 communicates with a curved high pressure groove 36 that opens up on its inner axial face 35, and on a high pressure groove 37 through a drilling not shown, that opens up on an inner radial face 38 of bearings 17 to 20.
  • a low pressure groove 39 extends to the junction of the inner axial face 35 and the inner radial face 38 and communicates with the low pressure dish 34 through a collective groove 40.
  • the stub shafts 13 to 16 are supported by these hydrodynamic layers in the inner radial faces 38 that occupy the gaps, and the hydrodynamic layers on the inner axial faces 35 form adjacent to the flanks of the gears 2 and 3, holding them slightly separated from the bearings 17 to 20 and therefore preventing the gaps from being entirely closed, despite the springs 30.
  • the mobile bearings 19 and 20 in their known construction, have a special feature at their outer axial face remote from the gears 2 and 3: this face is divided into two crescent shaped portions 42 and 43 that are located in different planes, separated by a shoulder fitted with a seal that exposes the crescent shaped portion 42 to high fluid pressure and the crescent shaped portion 43 to low pressure.
  • the mobile bearings 19 and 20 are held in place at the shoulder separating the crescent shaped portions 42 and 43, in grooves reamed in the housing 6, offset from the rotation axes of the gears 2 and 3.
  • a plurality of springs 30 are compressed between the mobile bearings 19 and 20 and one face of the housing 6.
  • the springs 30 are installed into corresponding compartments 41 only where the crescent shaped portion 42 is largest, over approximately a quarter of the circumference of the mobile bearings 19 and 20.
  • the thrust in the axial direction is then unbalanced and comprises a moment about a transverse axis of the mobile bearings 19 and 20. so as to balance an opposing moment created by pressure differences in the pumped fluid on the inner axial faces 35; therefore, there is no tilting of the mobile bearings 19 and 20 about this axis, so that they remain coaxial with the stub shafts 15 and 16 that they support, despite assembly gaps of the mobile bearings 19 and 20 in the housing 6.
  • the bearing dam substantially seals the inlet from the discharge side to maintain pumping efficiency.
  • the shape of the bearing dam can have a significant impact on gear venting and filling, and therefore mayimpact the cavitation performance of the gear pump.
  • the gear pump described in the ‘980 patent includes a bearing dam with a geometry that ostensibly reduces fluid cavitation and the damage that can result.
  • a driving gear bearing 104 includes a driving gear bearing half 204a and a driving gear bearing half 204b.
  • a driving gear 114 includes driving gear teeth 134, a central shaft portion 234a (e.g., a journal) extending axially from the driving gear teeth 134, and a central shaft portion 234b extending axially from the driving gear teeth 134 opposite the central shaft portion 234a.
  • the driving gear bearing half 204a includes a bore 250a
  • the driving gear bearing half 204b includes a bore 250b.
  • the bore 250a is formed to accept insertion of and rotationally support the central shaft portion 234a
  • the bore 250b is formed to accept insertion of and rotationally support the central shaft portion 234b, when the assembly 100 is in its assembled form.
  • a driven gear bearing 106 includes a driven gear bearing half 206a and a driven gear bearing half 206b.
  • a driven gear 116 includes driven gear teeth 136, a central shaft portion 236a extending axially from the driven gear teeth 136, and a central shaft portion 236b extending axially from the driven gear teeth 136 opposite the central shaft portion 236a.
  • the driven gear bearing half 206a includes a bore 250c, and the driven gear bearing half 206b includes a bore 250d.
  • the bore 250c is formed to accept insertion of and rotationally support the central shaft portion 236a, and the bore 250d is formed to accept insertion of and rotationally support the central shaft portion 236b, when the assembly 100 is in its assembled form.
  • the assembly 100 includes the central fluid dam half 258a within the area 6B in Fig. 6A proximate the bores 250a and 250c in the bearing half 204a and the bearing half 206a, respectively.
  • Fig. 6B is an enlarged view of the bearing dam half 258a shown in that area.
  • a corresponding central fluid dam half (not shown) is provided proximate the bores 250b and 250d in the bearing half 204b and the bearing half 206b, respectively. Together, the central fluid dam half 258a and the corresponding central fluid dam half form the beanng dam.
  • the central fluid dam half 258a of the bearing dam includes an inlet face 260 and an outlet face 261.
  • the inlet face 260 includes a slot 262 formed as a relief cut in the inlet face 260.
  • the outlet face 261 includes a vent 263 formed as a relief cut in the outlet face 261 .
  • the central fluid dam halves, the driving gear teeth 134, and the driven gear teeth 136 provide a barrier that substantially blocks the flow' of fluid between the fluid inlet cavity 160 and the fluid discharge cavity 180 along the bearing split line across the bearing center line.
  • the configuration of the inlet face 260, the outlet face 261, the slot 262, and the vent 263 is designed to address cavitation.
  • Fig. 7 is a side view of a collection of example gear teeth 300.
  • the gear teeth 300 can represent the driving gear teeth 134 and/or the driven gear teeth 136 of the example gear pump assembly 100.
  • the gear teeth 300 extend radially from a gear 302.
  • the gear 302 can be the driving gear 114 or the driven gear 116.
  • the gear 302 has a root diameter 304, which is the diameter at the base of a tooth space 306.
  • the gear 302 also includes a pitch circle 308.
  • the pitch circle 308 can be the circle derived from the number of the gear teeth 300 and a predetermined diametral or circular pitch, and can be the circle on which spacing or tooth profiles is established and from which the tooth proportions can be constructed.
  • the profile of the gear tooth is the shape of the gear tooth curve and is measured from the root to the tip of the gear tooth.
  • the functional, or operating, portion of the profile is the area that is in actual contact during tooth mesh.
  • Gears generally have an involute curve tooth profile. This involute curve helps the gears transmit power smoothly during the rolling action.
  • the tooth thickness, diametral pitch, and pressure angle all help determine the gear tooth profile. These factors are determined by the desired contact ratio between mating parts of the gear.
  • the gear tooth profile also varies by the number of teeth on the gear such that the larger the amount of teeth the straighter the profile of the gear eventually forming what is called a rack gear.
  • Each of the gear teeth 300 includes an addendum 310 and a dedendum 312.
  • the addendum 310 is the height by which the gear tooth 300 projects beyond the pitch circle 308; the dedendum 312 is the depth of the tooth space 306 between the pitch circle 308 and the root diameter 304.
  • Each of the gear teeth 300 also includes a pressure angle 320.
  • the pressure angle 320 is the angle at a pitch point 322 on the pitch circle 308 between the line of pressure which is normal to the tooth surface at the pitch point 322, and the plane tangent to the pitch circle 308.
  • the pressure angle 320 can be also described as the angle between a line of action 324 and a line 326 tangent to the pitch circle 308.
  • the gear teeth 300 of the gear 302 illustrated in Fig. 7 depict a standard full fillet root profile.
  • a “standard full fillet root profile” is that which provides a constant radius which extends in a continuous arc from one tooth 300 to the next tooth 300.
  • the typical geometry for a spur gear tooth root is a full fillet which is tangent to the involute tooth profile and simultaneously tangent to the root diameter. The lowest point of the constant radius fillet establishes the root diameter.
  • the geometry is generated by the path the tool tip follows as the teeth 300 are cut. For form ground teeth, the radius is formed on the extremity of the grinding wheel.
  • U.S. Patent No. 9,057,372 assigned to Hamilton Sundstrand Corporation of Windsor Locks, Connecticut discloses a modified gear root profile or geometry.
  • the modified gear root profile provides a desired enlarged carry-over fluid volume as compared to the standard full fillet root profile to mitigate the effects of fluid displacement. Ostensibly, the effects from the enlarged carry-over volume of the modified gear root geometry tend to reduce the phenomenon of cavitation within the gear mesh zone. Because other factors can be affected by profile changes such as leakage across the pump thus reducing volumetric efficiency, however, this approach has limited effectiveness.
  • the added volume is accomplished, as illustrated in Fig. 8, by modifying a gear 500 having a plurality of teeth 502, each tooth 502 having a contact side 503a, a non-contact side 503b, a tip 504, and a root 506, extending radially outward from a body 512 and a spline 508 formed in the center of the body 512.
  • the gear 500 is modified by machining cavities or pockets 510 directly into the roots 506 of the gear teeth 502. Material is removed from the non-contacting sides 503b of the teeth 502, only as shown, to form the pockets 510.
  • eighteen gear teeth 502 and pockets 510 are illustrated in the example of Fig. 8, an artisan would recognize that the number of gear teeth 502 and corresponding pockets 510 can be increased or decreased depending upon the application.
  • the gear 500 as modified with the pockets 510 can be incorporated into the gear pump 10, into the gear pump assembly 100, or into another application.
  • the pockets 510 are shaped to maximize the volume of liquid (e.g., fuel) which will occupy each pocket 510, while taking into account the structural considerations (e.g., the strength requirements) of the gear 500.
  • the pockets 510 must be sized and shaped to maximize their volume while minimizing the adverse impact of the pockets on the integrity' of the gear 500.
  • a preferred shape for the pockets 510 is an oval.
  • An oval shape is defined by a curve that is closed and always concave toward the center; a closed curve bounding a convex domain. Common objects like a football or an egg have oval-shaped sections.
  • an oval shape is preferred for at least one particular application.
  • the application included a gear 500 configured for use in the gear pump 10.
  • the example gear 500 had eighteen teeth 502, a pressure angle of about 30°, a root diameter of about 2.925 inches, a true involute form (TIF) diameter of about 3.059 inches, a circular tooth thickness of about 0.280 inches, and an outside diameter of about 3.700 inches.
  • the pockets 510 have a depth, D, of about 0.350 ⁇ 0.005 inches; a length, L, of about 0.450 ⁇ 0.001 inches; a width, W, of about 0.179 ⁇ 0.001 inches, and radii of curvature, R, of about 0.060 inches. The radii of curvature help to avoid stress concentrations.
  • the pockets 510 do not extend the full length of the gear 500.
  • the gear 500 has a first length
  • the root pockets 510 have a second length
  • the second length is less than the first length.
  • the gear 500 of the highlighted application has a length between the gear teeth 501, where the pockets 510 are formed, of about 0.700 inches. Therefore, the length of the pockets 510 (about 0.450 inches) is about 65% of the length of the gear 500.
  • the length of the pockets 510 is between about 50% and about 80% of the length of the gear 500. More preferably, the length of the pockets 510 is between about 55% and about 75% of the length of the gear 500. Still more preferably, the length of the pockets 510 is between about 60% and about 70% of the length of the gear 500.
  • Fig. 10 illustrates a portion of a gear pump 520, including the gear 500 having the pockets 510, located between a stationary or fixed bearing 517 (which is positioned on a fixed gear journal 517a) and a mobile or floating bearing 519 (which is positioned on a floating gear journal 519a).
  • the design tradeoff represented by the disclosed size and shape of the pockets 510 is just one example. Those skilled in the art should recognize that other configurations, which include multiple pockets, might achieve the best design tradeoff in different applications.
  • One alternative shape for the pockets 510 is shown in Fig. 11 and includes a beam 530 across the center of one or more of the pockets 510.
  • the beam 530 can extend partially or fully into the depth of the pocket 510; full extension would essentially divide the pocket 510 into two, separate pockets while maximizing the strength of the gear 500.
  • the beam 530 can be added for structural considerations, thus trading a small amount of volume for added gear strength.
  • Model MFP-590 main fuel pump design which has a conventional bearing design 550 with vent cuts or bearing channels 552, 554, and 556 in the driven bearing 560 and in the driver bearing 562 like the channels disclosed in the "980 patent and discussed above.
  • the Model MFP-590 fuel pump is relatively large in size, has relatively high gear tip speeds, and operates at relatively high pressures.
  • the conventional bearing design 550 is illustrated in Fig. 12, and has an inlet 564 and an outlet or discharge 566.
  • the initial tests showed, after about eight hours of operation, severe cavitation erosion in both the bearing bridges (breaching the bearing in localized areas) and the bearing journals.
  • the conventional bearing design 550 is required to operate for more than about 5,400 hours without overhaul or service. Therefore, these tests raised concern that the cavitation erosion is life limiting to the component and may result in premature failure.
  • Fig. 13 illustrates an analysis of trapped volume upon gear meshing for the conventional bearing design 550 shown in Fig. 12, and helps to explain why the conventional bearing design 550 experienced severe cavitation erosion.
  • a small amount of fluid volume is carried from the discharge 566 back to the inlet 564 through the mesh.
  • a small amount or volume of fluid is trapped (as the trapped fluid 570) in the mesh of the gears between a first seal point 572 and a second seal point 574.
  • the trapped fluid 570 is at high pressure.
  • the aggressive bearing design 580 is illustrated in Fig. 14, and has an inlet 564 and an outlet or discharge 566. As shown in Fig. 14, the aggressive bearing design 580 includes an added notch 582, an added channel 584. and an added discharge vent 586. Further, inlet dam material is removed from the area 588.
  • Fig. 15 illustrates an analysis of trapped volume upon gear meshing for the aggressive bearing design 580 shown in Fig. 14. and helps to explain why the aggressive bearing design 580 was insufficient to avoid cavitation damage.
  • Fuel in Bucket “B” at high pressure
  • Bucket “A” at low pressure
  • the fuel moves in the direction of arrow “C.” This results in high compression, pressure spikes, and cavitation.
  • Fig. 16 is a schematic diagram illustrating a trapped volume analysis for a specific fluid fuel traveling through the aggressive bearing design 580 shown in Fig. 14. The analysis was done for a fluid fuel having the following properties: a specific gravity of 0.818, a discharge coefficient of 0.75, and a bulk modulus of 150,000 PSIG.
  • PSIG refers to the gauge pressure, expressed in pounds per square inch gauge. It is a unit of pressure relative to the ambient pressure or atmospheric pressure, and measures pressure without factoring in local atmospheric pressure. The gauge pressure is applied when the pressure inside the system is greater than the atmospheric pressure.
  • Fig. 17 is a schematic diagram illustrating a trapped volume analysis for the specific fluid fuel traveling through the conventional bearing design shown in Fig. 12.
  • fuel in Bucket “B” (at high pressure) is compressed or squeezed into Bucket “A” (at low pressure) through the connection 596.
  • the connection 596 provides a small, narrow ⁇ opening between the HP area 592 and the LP area 594.
  • the fuel travels from the HP area 592 to the LP area 594.
  • Visible in Fig. 17 are the gear teeth positioned over two bearings; the bearings that would appear over the gear teeth are omitted for clarity.
  • the line of action 324 is shown.
  • Fig. 18A reflects calculations made at an angle of rotation of the gears of 10° and Fig. 18B reflects calculations made at an angle of rotation of the gears of 11° for the aggressive bearing design 580.
  • Figs. 18A and 18B each depict the driven gear 561, the driver gear 563, the line of action 324, the pitch circle 308, and five areas of trapped fluid volumes: mesh area 1, mesh area 2, mesh area 3, mesh area 4, and mesh area 5.
  • the following table summarizes the trapped mesh areas for the tw o angles of rotation.
  • PSID stands for “pounds per square inch differential.” and is used when measuring pressure relative to something other than atmosphenc pressure.
  • B bulk modulus
  • the amount of volume increase can also be calculated for the modification of adding the pockets 510 in comparison to the baseline of the aggressive bearing design 580.
  • the trapped volume in area 1 at an angle of rotation of 10° is 0.004932 in 3 for the aggressive bearing design 580 and 0.03443 in 3 for the modification of adding the pockets 510.
  • the trapped volume in area 1 at an angle of rotation of 11° is 0.004792 in 3 for the aggressive bearing design 580 and 0.03429 in 3 for the modification of adding the pockets 510.
  • the improvement achieved by the modification of adding the pockets 510, in comparison to the baseline of the aggressive bearing design 580, can also be measured by the effect of the modification on both pressure ripple and gear stress.
  • the pressure ripple can be calculated as the pressure fluctuation or amplitude of pressure deflection up (positive) or down (negative) divided by the total system pressure times 100. Therefore, for example, if the total pressure were 200 psi and the pressure were to fluctuate by 10 psi then the pressure ripple would be 5%.
  • FIG. 19A is a graph of measured pressure versus time for a MFP-590 pump with only the aggressive bearing design 580, and shows a discharge pressure ripple of about ⁇ 30 psi or 6.5%.
  • Fig. 19B is a graph of measured pressure versus time for a MFP-590 pump with the aggressive bearing design 580 and the pockets 510, and shows a discharge pressure ripple of about ⁇ 5 psi or 1 %.
  • the discharge pressure ripple was significantly quieted by adding the pockets 510.
  • the addition of the pockets 510 converted the pump being tested from a pump that failed to meet performance targets to a pump that met the discharge pressure ripple goal.
  • the gear was made of CPM-10V steel, which is strong and resistant to wear.
  • the gear teeth were subject to a nitriding surface treatment to provide further wear resistance.
  • Nitriding is a thermochemical treatment process applied to enrich the surface with nitrogen for the purpose of increasing the surface hardness. The process is based on the low sol ubi li ty of nitrogen in the ferritic cry stal structure to promote the precipitation of iron nitrides or alloy nitrides.
  • the yield strength of the gear material was about 70.4 ksi; yield stress is the stress limit after which the material starts deforming.
  • the maximum principal stress or major principal stress is the maximum value of normal stress acting on one of the principal planes of a component (such as the gear under analysis) where the value of shear stress is zero.
  • High cycle fatigue is a type of fatigue caused by small elastic strains under a high number of load cycles before failure occurs. The stress comes from a combination of mean and alternating stresses. Thus, the FEA done for a gear with the pockets 510 showed not only an acceptable FoS, but a relatively high FoS.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Rotary Pumps (AREA)

Abstract

L'invention concerne un engrenage ayant une expansion de volume pour une réduction de cavitation dans un engrènement de pompe à engrenages. L'engrenage a un profil de dent d'engrenage ; un corps ; une pluralité de dents d'engrenage à développante s'étendant radialement vers l'extérieur à partir du corps et comportant des première et seconde dents d'engrenage voisines ayant chacune une pointe respective et une racine, les première et seconde dents d'engrenage voisines définissant un espace entre elles ; et une poche de racine formée directement dans les racines des dents d'engrenage et dans l'espace entre les dents d'engrenage, fournissant un volume de racine d'engrenage accru et ajoutant une compliance de fluide piégé tout en laissant inchangé le profil de dent d'engrenage. L'invention divulgue également une pompe à engrenages comprenant l'engrenage.
EP23844402.0A 2022-12-19 2023-12-12 Expansion de volume pour une réduction de cavitation dans un engrènement de pompe à engrenages Pending EP4638963A1 (fr)

Applications Claiming Priority (3)

Application Number Priority Date Filing Date Title
US202263433587P 2022-12-19 2022-12-19
US18/372,480 US12486843B2 (en) 2022-12-19 2023-09-25 Volume expansion for cavitation reduction in a gear pump mesh
PCT/US2023/083470 WO2024137265A1 (fr) 2022-12-19 2023-12-12 Expansion de volume pour une réduction de cavitation dans un engrènement de pompe à engrenages

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EP4638963A1 true EP4638963A1 (fr) 2025-10-29

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EP (1) EP4638963A1 (fr)
JP (1) JP2025542275A (fr)
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CN121723608B (zh) * 2026-02-25 2026-04-21 东北大学 改进的弧齿锥齿轮承载接触分析方法、介质及电子装置

Family Cites Families (9)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US2344628A (en) * 1940-12-26 1944-03-21 Gar Wood Ind Inc Gear pump
GB1467441A (en) * 1973-03-15 1977-03-16 Lucas Industries Ltd Hydraulic gear pumps and motors
US4233005A (en) * 1978-01-18 1980-11-11 Lucas Industries Limited Hydraulic gear pump with recesses in non-working gear flanks
GB1547944A (en) * 1978-04-05 1979-07-04 Rolls Royce Gear pumps or motors
US7878781B2 (en) 2007-12-11 2011-02-01 Hamilton Sundstrand Corporation Gear pump cavitation reduction
US9057372B2 (en) 2010-12-06 2015-06-16 Hamilton Sundstrand Corporation Gear root geometry for increased carryover volume
US9303644B2 (en) 2013-11-26 2016-04-05 Woodward, Inc. Gear pump bearing dam
FR3019856B1 (fr) 2014-04-15 2019-05-31 Safran Aircraft Engines Pompe a engrenages, prevue notamment comme pompe a haute pression a carburant
US11149729B2 (en) * 2017-07-18 2021-10-19 Eaton Intelligent Power Limited Pump with bleed mechanism for reducing cavitation

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IL321604A (en) 2025-08-01
JP2025542275A (ja) 2025-12-25

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