US20140314608A1 - Internal-gear-type oil pump for vehicle - Google Patents

Internal-gear-type oil pump for vehicle Download PDF

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Publication number
US20140314608A1
US20140314608A1 US14/357,366 US201114357366A US2014314608A1 US 20140314608 A1 US20140314608 A1 US 20140314608A1 US 201114357366 A US201114357366 A US 201114357366A US 2014314608 A1 US2014314608 A1 US 2014314608A1
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United States
Prior art keywords
dynamic pressure
driven gear
pressure generating
circumferential surface
generating grooves
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Abandoned
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US14/357,366
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English (en)
Inventor
Hiroyasu Honda
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Toyota Motor Corp
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Toyota Motor Corp
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Assigned to TOYOTA JIDOSHA KABUSHIKI KAISHA reassignment TOYOTA JIDOSHA KABUSHIKI KAISHA ASSIGNMENT OF ASSIGNORS INTEREST (SEE DOCUMENT FOR DETAILS). Assignors: HONDA, HIROYASU
Publication of US20140314608A1 publication Critical patent/US20140314608A1/en
Abandoned legal-status Critical Current

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    • 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
    • F04C18/00—Rotary-piston pumps specially adapted for elastic fluids
    • F04C18/08—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
    • 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
    • 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
    • F04C2/10—Rotary-piston machines or pumps of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of internal-axis type with the outer member having more teeth or tooth-equivalents, e.g. rollers, than the inner member
    • F04C2/102—Rotary-piston machines or pumps of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of internal-axis type with the outer member having more teeth or tooth-equivalents, e.g. rollers, than the inner member the two members rotating simultaneously around their respective axes
    • 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
    • F04C2240/00—Components
    • F04C2240/50—Bearings
    • F04C2240/54—Hydrostatic or hydrodynamic bearing assemblies specially adapted for rotary positive displacement pumps or compressors
    • 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
    • F04C2240/00—Components
    • F04C2240/50—Bearings
    • F04C2240/56—Bearing bushings or details thereof

Definitions

  • the present invention relates to a vehicular internal-gear-type oil pump provided with a driven gear, and more particularly to techniques for optimizing a depth of a plurality of grooves formed in an outer circumferential surface of the driven gear.
  • the vehicular internal-gear-type oil pump is provided with (a) a pump chamber defined by a pump body and a pump cover, (b) an annular driven gear which has internal teeth, and an outer circumferential surface opposed to an inner circumferential surface defining the pump chamber, and which is rotatably supported by the inner circumferential surface defining the pump chamber, and (c) a drive gear which has external teeth engaging with the internal teeth of the driven gear and which is disposed rotatably about an axis of rotation thereof eccentric with respect to an axis of rotation of the driven gear, to rotate the driven gear.
  • Patent Documents 1 and 2 disclose an example of such a vehicular internal-gear-type oil pump.
  • the vehicular internal-gear-type oil pump as described above is generally configured such that the above-described driven gear is held, by gravity, in contact with the inner circumferential surface of the above-described pump chamber, when the above-described driven gear is not rotated.
  • the driven gear is supported by a working oil, without its outer circumferential surface being held in contact with the inner circumferential surface of the above-described pump chamber, such that the working oil existing in an annular gap formed between the outer circumferential surface of the above-described driven gear and the inner circumferential surface of the above-described pump chamber is moved through the gap in a circumferential direction, due to a rotary motion of the driven gear, with a generated dynamic pressure of the working oil being maximized in a portion of the gap in which the outer circumferential surface of the above-described driven gear and the inner circumferential surface of the above-described pump chamber are closest to each other, since the amount of the gap gradually decreases in the circumfer
  • the vehicular internal-gear-type oil pump configured as described above suffers from a problem of insufficient stability of balance of the dynamic pressure generated between the outer circumferential surface of the above-described driven gear and the inner circumferential surface of the above-described pump chamber when the oil pump is operated at a low speed, or operated to pressurize the working oil to a high pressure.
  • the insufficient stability of balance of the dynamic pressure may cause an oscillatory motion of the above-described driven gear, namely, oscillation of the axis of rotation of the driven gear.
  • This oscillation of the axis of rotation of the driven gear causes a friction loss due to boundary lubrication between the outer circumferential surface of the above-described driven gear and the inner circumferential surface of the above-described pump chamber, resulting in an increase of a resistance to the rotary motion of the above-described driven gear.
  • Patent Documents 3 and 4 describe a vehicular internal-gear-type oil pump wherein the above-described driven gear has protrusions projecting from its outer circumferential surface toward the inner circumferential surface of the above-described pump chamber.
  • the above-indicated protrusions generate a higher dynamic pressure of the working oil than in a vehicular internal-gear-type oil pump not having such protrusions, during rotation of the above-described driven gear.
  • the vehicular internal-gear-type oil pump is provided with grooves in the form of wedges formed in the inner circumferential surface of the pump body, as in an oil pump having a dynamic pressure bearing structure as disclosed in Patent Document 5, so that the oscillation of the axis of rotation of the above-described driven gear is reduced.
  • the vehicular internal-gear-type oil pump as disclosed in Patent Documents 3 and 4 and the vehicular internal-gear-type oil pump having the dynamic pressure bearing structure as disclosed in Patent Document 5 may have a risk of deterioration of the automatic centering function of the above-described driven gear due to decrease of the dynamic pressure, and an increase of a fluid friction acting on the above-described driven gear, and a consequent friction loss, depending upon a height of the above-indicated protrusions in the radial direction of the above-described driven gear, in other words, the depth of grooves formed between the above-indicated protrusions in the radially inward direction of the driven gear.
  • the present invention was made in view of the background art described above. It is therefore an object of the present invention to provide a vehicular internal-gear-type oil pump which is provided with a driven gear and which is configured to permit the driven gear to have an automatic centering function with a reduced increase of a fluid friction acting on the driven gear.
  • a vehicular internal-gear-type oil pump provided with (a) a circular pump chamber defined by a pump body and a pump cover, (b) an annular driven gear which has internal teeth, and an outer circumferential surface opposed to an inner circumferential surface defining the pump chamber, and which is rotatably supported by the inner circumferential surface defining said pump chamber, and (c) a drive gear which has external teeth engaging with the internal teeth of the driven gear and which is disposed rotatably about an axis of rotation thereof eccentric with respect to an axis of rotation of the driven gear, to rotate the driven gear, characterized in that (d) the above-described driven gear has a plurality of first dynamic pressure generating grooves formed in local areas of its outer circumferential surface, and (e) each of the above-described first dynamic pressure generating grooves has a depth in a radial direction of the driven gear, which depth is determined such that a gap ratio which is a ratio of
  • the above-described driven gear has the plurality of first dynamic pressure generating grooves formed in the local areas of its outer circumferential surface, and (e) each of the above-described first dynamic pressure generating grooves has a depth in the radial direction of the driven gear, which depth is determined such that the gap ratio which is the ratio of the gap from the bottom of the above-described first dynamic pressure generating groove to the above-described inner circumferential surface to the gap from the outer circumferential surface of the above-described driven gear to the above-described inner circumferential surface is held within the predetermined range in which the dynamic pressure generated by the first dynamic pressure generating grooves and changing as a function of the above-described gap ratio has the maximal value and in which the fluid friction coefficient generated on the basis of the above-described first dynamic pressure generating grooves and changing as a function of the above-described gap ratio has the minimal value.
  • the fluid friction coefficient acting on the outer circumferential surface of the above-described driven gear is minimized, and the dynamic pressure generated by the above-described first dynamic pressure generating grooves is maximized while the driven gear is rotated, so that the above-described driven gear can be given the function of automatic centering in its radial direction while an increase of the fluid friction acting on the above-described driven gear is reduced or prevented.
  • the above-described first dynamic pressure generating groove has a slant surface formed downwardly toward the bottom thereof in the outer circumferential surface of the above-described driven gear, such that the slant surface cooperates with the inner circumferential surface of the above-described pump chamber to define a wedge space.
  • the fluid friction coefficient acting on the above-described driven gear is reduced, and the dynamic pressure generated by the above-described first dynamic pressure generating grooves is increased.
  • the above-described plurality of first dynamic pressure generating grooves are formed in the outer circumferential surface of the above-described driven gear such that the first dynamic pressure generating grooves are equiangularly spaced apart from each other, about the axis of rotation of the driven gear. According to this form of the invention, the function of automatic centering of the above-described driven gear is effectively improved.
  • the above-described first dynamic pressure generating grooves have a depth determined such that the above-described gap ratio is held within a range between 2 and 3. According to this form of the invention, the fluid friction coefficient acting on the above-described driven gear is almost minimized, and the dynamic pressure generated by the above-described first dynamic pressure generating grooves is almost maximized.
  • the above-described driven gear has a plurality of second dynamic pressure generating grooves formed in local areas of its opposite side surfaces, and (b) each of the above-described second dynamic pressure generating grooves has a depth in a thickness direction of the driven gear, which depth is determined such that a gap ratio which is a ratio of a gap from a bottom of the above-described second dynamic pressure generating groove to inner wall surfaces of the above-described pump chamber, to a gap from the side surfaces of the above-described driven gear to the inner wall surfaces of the above-described pump chamber is held within a predetermined range in which a dynamic pressure generated by the second dynamic pressure generating grooves has a maximal value and in which a fluid friction coefficient generated on the basis of the above-described second dynamic pressure generating grooves has a minimal value.
  • the fluid friction coefficient acting on the opposite side surfaces of the driven gear is minimized, and the dynamic pressure generated by the above-described second dynamic pressure generating grooves is maximized while the driven gear is rotated, so that the above-described driven gear can be given the function of automatic centering in its axial direction while an increase of the fluid friction acting on the above-described driven gear is reduced or prevented.
  • the above-described drive gear has a plurality of third dynamic pressure generating grooves formed in local areas of its opposite side surfaces
  • each of the above-described third dynamic pressure generating grooves has a depth in a thickness direction of the drive gear, which depth is determined such that a gap ratio which is a ratio of a gap from a bottom of the above-described third dynamic pressure generating groove to inner wall surfaces of the above-described pump chamber, to a gap from the side surfaces of the above-described drive gear to the inner wall surfaces of the above-described pump chamber is held within a predetermined range in which a dynamic pressure generated by the third dynamic pressure generating grooves has a maximal value and in which a fluid friction coefficient generated on the basis of the above-described third dynamic pressure generating grooves has a minimal value.
  • the fluid friction coefficient acting on the opposite side surfaces of the drive gear is minimized, and the dynamic pressure generated by the above-described third dynamic pressure generating grooves is maximized while the drive gear is rotated, so that the above-described drive gear can be given the function of automatic centering in its axial direction while an increase of the fluid friction acting on the above-described drive gear is reduced or prevented.
  • FIG. 1 is a fragmentary cross sectional view showing a portion of a vehicular power transmitting system which includes a vehicular internal-gear-type oil pump according to one embodiment of this invention
  • FIG. 2 is a view showing the driven gear and the drive gear accommodated within the pump body shown in FIG. 1 , as seen from a surface of the pump body at which the pump body is fixed;
  • FIG. 3 is an enlarged view showing in enlargement the driven gear and the drive gear of FIG. 2 ;
  • FIG. 4 is a perspective view showing the driven gear of FIG. 3 ;
  • FIG. 5 is an enlarged view showing in enlargement an area within a circle indicated by a one-dot chain line in FIG. 2 , for explaining a shape of first dynamic pressure generating grooves formed in the driven gear of FIG. 3 ;
  • FIG. 6 is a cross sectional view taken along lines A-A in FIG. 3 , showing a shape of the second dynamic pressure generating grooves formed in opposite side surfaces of the driven gear of FIG. 3 ;
  • FIG. 7 is an enlarged view showing in enlargement an area within a circle indicated by a one-dot chain line in FIG. 3 , for explaining the shape of the second dynamic pressure generating grooves formed in the driven gear of FIG. 3 ;
  • FIG. 8 is a cross sectional view taken along lines A-A in FIG. 3 , showing a shape of the third dynamic pressure generating grooves formed in the opposite side surfaces of the drive gear of FIG. 3 ;
  • FIG. 9 is a view for explaining a thrust force generated in the radial direction of the driven gear during rotation of the driven gear
  • FIG. 10 is a view for explaining a thrust force generated in a direction of thickness of the driven gear during rotation of the driven gear;
  • FIG. 11 is a view for explaining a thrust force generated in a direction of thickness of the driven gear during rotation of the drive gear;
  • FIG. 12 is a view indicating a relationship between a distance of a gap between the outer circumferential surface of the driven gear and the inner circumferential surface of the pump chamber, and a dynamic pressure generated in the gap, when the driven gear is positioned eccentrically with respect to the axis of the pump chamber during rotation of the driven gear;
  • FIG. 13 is a view for explaining automatic centering forces acting on the driven gear in the radial direction when the driven gear is positioned eccentrically in the radial direction;
  • FIG. 14 is a view for explaining automatic centering forces acting on the driven gear in the thrust direction when the driven gear is positioned eccentrically in the direction of its thickness;
  • FIG. 15 is a view for explaining automatic centering forces acting on the driven gear in the thrust direction when the driven gear is positioned eccentrically in the direction of its thickness and when the centerline of the driven gear is inclined with respect to the centerline of the pump chamber;
  • FIG. 16 is a view for explaining automatic centering forces acting on the drive gear in the thrust direction when the drive gear is positioned eccentrically in the direction of its thickness;
  • FIG. 17 is a view indicating a relationship among a gap ratio, a dynamic pressure generated by the first dynamic pressure generating grooves, and a fluid friction coefficient;
  • FIG. 18 is a view corresponding to that of FIG. 5 , showing a shape of first dynamic pressure generating grooves formed in the outer circumferential surface of the driven gear in a vehicular internal-gear-type oil pump according to another embodiment of this invention
  • FIG. 19 is a view corresponding to that of FIG. 5 , showing a shape of first dynamic pressure generating grooves formed in the outer circumferential surface of the driven gear in a vehicular internal-gear-type oil pump according to a further embodiment of this invention
  • FIG. 20 is a view corresponding to that of FIG. 3 , showing a vehicular internal-gear-type oil pump according to a still further embodiment of this invention.
  • FIG. 21 is a perspective view corresponding to that of FIG. 4 , showing the driven gear provided in the vehicular internal-gear-type oil pump of FIG. 20 .
  • FIG. 1 is the fragmentary cross sectional view showing a portion of a vehicular power transmitting system 12 which includes a vehicular internal-gear-type oil pump (hereinafter referred to as an “oil pump”) 10 according to one embodiment of this invention.
  • the vehicular power transmitting system 12 is provided with a torque converter 16 and a step-variable automatic transmission 18 , which are disposed downstream of a crankshaft 14 of an engine provided as a vehicle drive power source.
  • the torque converter 16 is provided with a pump impeller 20 operatively connected to the crankshaft 14 , a turbine impeller 24 disposed rotatably relative to the pump impeller 20 and operatively connected to an input shaft 22 of the automatic transmission 18 , and a stator impeller 28 disposed between the pump impeller 20 and the turbine impeller 24 and rotatably supported via a one-way clutch 26 .
  • a rotary motion of the pump impeller 20 rotating together with the crankshaft 14 as a unit is transmitted to the turbine impeller 24 through a working fluid.
  • the pump impeller 20 is provided with a cylindrical sleeve 20 a disposed radially outwardly of the input shaft 22 so as to extend toward the automatic transmission 18 .
  • the oil pump 10 is driven by this sleeve 20 a of the pump impeller 20 .
  • the torque converter 16 and the automatic transmission 18 are accommodated within a cylindrical transmission casing 32 fixed to an engine block 30 indicated by two-dot chain lines in FIG. 1 .
  • the input shaft 20 extends through a partition wall disposed between an accommodating space 32 a accommodating the torque converter 16 , and an accommodating space 32 b accommodating the automatic transmission 18 .
  • the oil pump 10 is provided with a pump body 34 and a pump cover 36 , which constitute a part of the above-indicated partition wall.
  • the pump body 34 is formed annularly and disposed radially outwardly of the sleeve 20 a , and is fitted in a fitting hole 32 c which is a cylindrical portion of an inner circumferential surface of the transmission casing 32 .
  • the pump cover 36 is formed annularly and disposed radially outwardly of the input shaft 22 , and is fitted in a comparatively shallow fitting hole 34 a which is formed with a comparatively large diameter, in one of opposite side surfaces of the pump body 34 that is remote from the torque converter 16 .
  • the pump body 34 is integrally fixed to the transmission casing 32 with first screws 38
  • the pump cover 36 is integrally fixed to the pump body 36 with second screws 40 .
  • the pump body 34 has a cylindrical hole 34 b which is open in the bottom surface of the fitting hole 34 a and which has a smaller diameter and a larger depth than the fitting hole 34 a .
  • the cylindrical hole 34 b has an axis O 1 which is eccentric with respect to an axis of rotation C 1 of the input shaft 22 and the sleeve 20 a .
  • the pump body 34 and the pump cover 36 define a circular pump chamber 42 .
  • This pump chamber 42 is a cylindrical space which is formed radially outwardly of the sleeve 20 a , and which is defined by an inner circumferential surface 34 c of the hole 34 b , and inner wall surfaces 34 d and 36 a which are positioned at the respective opposite ends of the inner circumferential surface 34 c in the direction of the axis O 1 .
  • the pump chamber 42 has the axis O 1 eccentric with respect to the axis of rotation C 1 of the sleeve 20 a.
  • FIG. 2 is the view showing the oil pump 10 as seen from a surface of the pump body 34 shown in FIG. 1 at which the pump body 34 is fixed.
  • FIG. 1 is the cross sectional view taken along lines I-I in FIG. 2 .
  • the oil pump 10 is provided with: an annular driven gear 46 which has internal teeth 46 a , and an outer circumferential surface 46 b opposed to the inner circumferential surface 34 c defining the pump chamber 42 , and which is rotatably supported by the inner circumferential surface 34 c ; and a drive gear 48 which has external teeth 48 a engaging with the internal teeth 46 a of the driven gear 46 and which is disposed rotatably about the axis of rotation C 1 eccentric with respect to an axis of rotation C 2 of the driven gear 46 , to rotate the driven gear 46 .
  • the drive gear 48 is fitted on the sleeve 20 a such that drive gear 48 is rotatable with the sleeve 20 a and movable relative to the sleeve 20 a in the direction of the axis of rotation C 1 of the sleeve 20 a .
  • the driven gear 46 is rotated about the axis C 2 by the drive gear 48 in a direction indicated by an arrow “b” in FIG. 2 .
  • the oil pump 10 is of an internal-gear type wherein the external teeth 48 a of the drive gear 48 and the internal teeth 46 a of the driven gear 46 the number of which is larger by one than that of the external teeth 48 a are held in engagement with each other, as shown in FIGS. 2 and 3 , in a lower portion of the pump chamber 42 as seen in FIGS. 2 and 3 .
  • the internal teeth 46 a and the external teeth 48 a define a plurality of spaces, namely, pressure chambers in the pump chamber 42 , which are moved in the circumferential direction of the driven gear 46 when the drive gear 48 and the driven gear 46 are rotated.
  • the volumes of the pressure chambers increase as the pressure chambers are moved in the upward direction in the pump chamber 42 as seen in FIGS. 2 and 3 , and decrease as the pressure chambers are moved in the downward direction in the pump chamber 42 as seen in FIG. 2 .
  • the pump body 34 has a suction inlet 50 and a delivery outlet 52 formed in its radially outer portion fitted in the transmission casing 32 .
  • the suction inlet 50 is connected to a suction oil passage not shown, through which a working oil returned to an oil pan of the automatic transmission 18 , for example, is sucked into the suction inlet 50
  • the delivery outlet 52 is connected to a delivery oil passage not shown, through which the pressurized working oil is fed to a hydraulic control circuit for controlling hydraulically operated frictional coupling devices, for instance.
  • the pump body 34 further has a first inlet passage 56 for communication between the suction inlet 50 and a first suction port 54 formed on a pump body 34 side of the pump chamber 42 , and a first outlet (delivery) passage 60 for communication between the delivery outlet 52 and a first delivery port 58 formed on a pump body 34 side of the pump chamber 42 .
  • the pump cover 36 has a second suction passage not shown for communication between the suction inlet 50 and a second suction port not shown formed on a pump cover 36 side of the pump chamber 42 , and a second outlet (delivery) passage not shown formed for communication between the delivery outlet 52 and a second delivery port not shown formed on a pump cover 36 side of the pump chamber 42 .
  • the above-indicated second suction passage is held in communication with the first inlet passage 56 through a first communication port 62 open in the bottom surface of the fitting hole 34 a of the pump body 34 , while the above-indicated second outlet passage is held in communication with the first outlet passage 60 through a second communication port 64 open in the bottom surface of the fitting hole 34 a of the pump body 34 .
  • first suction port 54 and the above-indicated second suction port are positioned in the circumferential direction of the driven gear 46 , such that the volume of each of the above-described pressure chambers increases as the pressure chamber is moved in the circumferential direction of the driven gear 46
  • first delivery port 58 and the above-indicated second delivery port are positioned in the circumferential direction of the driven gear 46 , such that the volume of each pressure chamber decreases as the pressure chamber is moved in the circumferential direction of the driven gear 46 .
  • the drive gear 48 is rotated by the sleeve 20 a , in the direction indicated by the arrow “a” in FIG. 2
  • the driven gear 46 is rotated by the drive gear 48 , in the direction indicated by the arrow “b” in FIG. 2 , so that the working oil is sucked from the above-described oil pan into the pump chamber 42 through the suction inlet 50 , the first inlet passage 56 or the above-indicated second inlet passage, the first suction port 54 and the above-indicated second suction port.
  • the working oil sucked into the pump chamber 42 is admitted into one of the plurality of spaces defined by the internal teeth 46 a and the external teeth 48 a .
  • the working oil admitted in the space is pressurized with a decrease of the volume of the space as the space is moved in the circumferential direction of the drive gear 48 with a rotary motion of the drive gear 48 .
  • the thus pressurized working oil is delivered from the delivery outlet 52 into the above-described hydraulic control circuit through the first delivery port 58 or the above-indicated second delivery port, and the first outlet passage 60 or the above-indicated second outlet passage.
  • the driven gear 46 has a plurality of first dynamic pressure generating grooves 46 c formed in local areas of its outer circumferential surface 46 b .
  • the plurality of first dynamic pressure generating grooves 46 c formed in the outer circumferential surface 46 b are equiangularly spaced apart from each other in the circumferential direction of the driven gear 46 , about the axis of rotation C 2 of the driven gear 46 .
  • the gap ratio m 1 is a ratio of a distance h 1 of a gap H 1 from the bottom of the first dynamic pressure generating groove 46 c to the inner circumferential surface 34 c of the pump body 34 in the radial direction, to a distance h 2 of a gap H 2 from the outer circumferential surface 46 b of the driven gear 46 to the inner circumferential surface 34 c of the pump body 34 in the radial direction.
  • the depth D 1 of the first dynamic pressure generating groove 46 c in the radial direction of the driven gear 46 is equal to a difference (h 1 ⁇ h 2 ), which is equal to the distance h 1 minus the distance h 2 , as is apparent from FIG. 5 .
  • the distance h 1 of the gap H 1 is 125 ⁇ m
  • the distance h 2 of the gap H 2 is 55 ⁇ m
  • the depth D 1 of the first dynamic pressure generating groove 46 c is 70 ⁇ m.
  • the first dynamic pressure generating groove 46 c is a substantially triangular groove formed in the outer circumferential surface 46 b of the driven gear 46 .
  • the first dynamic pressure generating groove 46 c formed in the outer circumferential surface 46 b of the driven gear 46 has a slant surface 46 d inclined downwardly toward the bottom of the groove 46 c in the direction of rotation of the driven gear 46 opposite to the direction indicated by the arrow “b”, and a slant surface 46 h inclined from the bottom of the groove 46 c such that a distance from the slant surface 46 h to the inner circumferential surface 34 b of the pump chamber 42 decreases in the direction of rotation of the driven gear 46 opposite to the direction indicated by the arrow “b”.
  • the first dynamic pressure generating groove 46 c has the slant surface 46 d formed in the outer circumferential surface 46 b of the driven gear 46 such that the slant surface 46 d is inclined downwardly toward the bottom in the circumferential direction of the driven gear 46 , and such that the slant surface 46 d cooperates with the inner circumferential surface 34 c of the pump chamber 42 to define a wedge space.
  • the driven gear 46 has a plurality of second dynamic pressure generating grooves 46 g in the form of wedges formed in local areas of a side surface 46 e thereof opposed to the inner wall surface 36 a of the pump chamber 42 and a side surface 46 f thereof opposed to the inner wall surface 34 d of the pump chamber 42 .
  • the second dynamic pressure generating grooves 46 g have a shape as indicated in FIG. 7 , and are equiangularly spaced apart from each other in the circumferential direction of the side surfaces 46 e and 46 f of the driven gear 46 , about the axis of rotation C 2 of the driven gear 46 .
  • the gap ratio m 2 is a ratio of a distance h 3 of a gap H 3 from the bottom of the second dynamic pressure generating groove 46 g to the inner wall surfaces 34 d and 36 a of the pump chamber 42 , to a distance h 4 of a gap H 4 from the side surfaces 46 e and 46 f of the drive gear 46 to the inner wall surfaces 34 d and 36 a of the pump chamber 42 .
  • the depth D 2 of the second dynamic pressure generating groove 46 g in the direction of thickness of the driven gear 46 is equal to a difference (h 3 ⁇ h 4 ), which is equal to the distance h 3 minus the distance h 4 , as is apparent from FIG. 6 .
  • the distance h 3 of the gap H 3 is 36 ⁇ m
  • the distance h 4 of the gap H 4 is 16 ⁇ m
  • the depth D 2 of the second dynamic pressure generating groove 46 g is 20 ⁇ m.
  • the drive gear 48 has a plurality of third dynamic pressure generating grooves 48 d in the form of wedges formed in local areas of a side surface 48 b thereof (shown in FIG. 1 ) opposed to the inner wall surface 36 a of the pump chamber 42 and a side surface 48 c thereof (shown in FIG. 1 ) opposed to the inner wall surface 34 d of the pump chamber 42 .
  • the third dynamic pressure generating grooves 48 d are equiangularly spaced apart from each other in the circumferential direction of the side surfaces 48 b and 48 c , about the axis of rotation C 1 of the drive gear 48 .
  • the gap ratio m 3 is a ratio of a distance h 5 of a gap H 5 from the bottom of the third dynamic pressure generating groove 48 d to the inner wall surfaces 34 d and 36 a of the pump chamber 42 , to a distance h 6 of a gap H 6 from the side surfaces 48 b and 48 c of the drive gear 48 to the inner wall surfaces 34 d and 36 a of the pump chamber 42 .
  • the depth D 3 of the third dynamic pressure generating groove 48 d in the direction of thickness of the drive gear 48 is equal to a difference (h 5 ⁇ h 6 ), which is equal to the distance h 5 minus the distance h 6 , as is apparent from FIG. 8 .
  • the distance h 5 of the gap H 5 is 36 ⁇ m
  • the distance h 6 of the gap H 6 is 16 ⁇ m
  • the depth D 3 of the third dynamic pressure generating groove 48 d is 20 ⁇ m.
  • the gap H 2 formed between the outer circumferential surface 46 b of the driven gear 46 having the first dynamic pressure generating grooves 46 c and the inner circumferential surface 34 c of the pump body 34 , as indicated in FIG. 5 , is filled with the viscous working oil, so that a dynamic pressure (generated dynamic pressure) P 1 of the working oil is maximized in a circumferential portion at which the gap distance is smallest.
  • the gap H 4 formed between the side surface 46 e of the driven gear 46 having the second dynamic pressure generating grooves 46 g and the inner wall surface 36 a of the pump chamber 42 , and the gap H 4 formed between the side surface 46 f of the driven gear 46 having the second dynamic pressure generating grooves 46 g and the inner wall surface 34 d of the pump chamber 42 , as indicated in FIG. 6 , are filled with the viscous working oil, so that a dynamic pressure (generated dynamic pressure) P 2 of the working oil is maximized in the circumferential portion at which the gap distance is smallest.
  • the gap H 6 formed between the side surface 48 b of the drive gear 48 having the third dynamic pressure generating grooves 48 d and the inner wall surface 36 a of the pump chamber 42 , and the gap H 6 formed between the side surface 48 c of the drive gear 48 having the third dynamic pressure generating grooves 48 d and the inner wall surface 34 d of the pump chamber 42 , as indicated in FIG. 8 , are filled with the viscous working oil, so that a dynamic pressure (generated dynamic pressure) P 3 of the working oil is maximized in circumferential portions at which the gap distance is smallest.
  • the dynamic pressure P 1 generates a thrust force acting on the outer circumferential surface 46 b of the driven gear 46 toward the axis of rotation C 2 of the driven gear 46 , as indicated in FIG. 9 , so that the driven gear 46 is rotated with its outer circumferential surface 46 b being held spaced apart from the inner circumferential surface 34 c of the pump body 34 , as indicated in FIG. 9 .
  • the dynamic pressure P 2 generates a thrust force which acts on the side surface 46 f of the driven gear 46 toward the inner wall surface 34 d of the pump chamber 42 , and which also acts on the side surface 46 e of the driven gear 46 toward the inner wall surface 36 a of the pump chamber 42 , as indicated in FIG.
  • the driven gear 46 is rotated with its side surfaces 46 e and 46 f being held spaced apart from the inner wall surfaces 34 d and 36 a of the pump chamber 42 .
  • the dynamic pressure P 3 generates a thrust force which acts on the side surface 48 b of the drive gear 48 toward the inner wall surface 34 d of the pump chamber 42 , and which also acts on the side surface 48 c of the drive gear 48 toward the inner wall surface 36 a of the pump chamber 42 , as indicated in FIG. 11 , so that the drive gear 48 is rotated with its side surfaces 48 b and 48 c being held spaced apart from the inner wall surfaces 34 d and 36 a of the pump chamber 42 .
  • FIG. 12 is the view indicating a relation between the distance h 2 of the gap H 2 between the outer circumferential surface 46 b of the driven gear 46 and the inner circumferential surface 34 d of the pump body 34 , and the dynamic pressure P 1 generated in the gap H 2 filled with the viscous working oil, when the axis of rotation C 2 of the driven gear 46 is positioned eccentrically moved with respect to the driven gear center position A 1 coincident with the axis O 1 of the pump chamber 42 indicated in FIG. 9 , due to the eccentric force F indicated in FIG. 9 , during rotation of the driven gear 46 .
  • FIG. 12 is the view indicating a relation between the distance h 2 of the gap H 2 between the outer circumferential surface 46 b of the driven gear 46 and the inner circumferential surface 34 d of the pump body 34 , and the dynamic pressure P 1 generated in the gap H 2 filled with the viscous working oil, when the axis of rotation C 2 of the driven gear 46 is positioned eccentrically moved with respect to the driven gear center position
  • “HIGH DYNAMIC PRESSURE ECCENTRIC SIDE” is one of opposite sides of the center position A 1 of the driven gear 46 , on which the gap H 2 between the outer circumferential surface 46 b of the driven gear 46 and the inner circumferential surface 34 c of the pump body 34 is smaller so that the dynamic pressure P 1 is higher
  • “LOW DYNAMIC PRESSURE ECCENTRIC SIDE” is the other side of the center position A 1 of the driven gear 46 , on which the gap H 2 between the outer circumferential surface 46 b of the driven gear 46 and the inner circumferential surface 34 c of the pump body 34 is larger so that the dynamic pressure P 1 is lower.
  • the dynamic pressure P 1 which increases along a curve of the second order with an amount of eccentricity of the axis of rotation C 2 of the driven gear 46 with respect to the axis O 1 of the pump chamber 42 is generated in the reduced gap between the outer circumferential surface 46 b of the driven gear 46 and the inner circumferential surface 34 d of the pump body 34 , so that a radial automatic centering force acts on the driven gear 46 so that the gap H 2 is made constant in the circumferential direction of the driven gear 46 , that is, so as to move the axis of rotation C 2 of the driven gear 46 toward the axis O 1 of the pump chamber 42 .
  • the above-indicated radial automatic centering force causes a change of the lubricating condition from the boundary lubrication back to fluid lubrication.
  • the side surfaces 46 e and 46 f of the driven gear 46 which define the thickness of the driven gear 46 and the side surfaces 48 b and 48 c of the drive gear 48 which define the thickness of the drive gear 48 respectively have the second dynamic pressure generating grooves 46 g and the third dynamic pressure generating grooves 48 d , which generate a thrust automatic centering force, like the first dynamic pressure generating grooves 46 c.
  • the dynamic pressure P 2 which increases along a curve of the second order with an amount of a distance (i.e., an amount of biasing) from the centerline C 4 of the driven gear 46 to the centerline C 3 of the pump chamber 42 is generated in the reduced gap between the side surface 46 e of the driven gear 46 and the inner circumferential surface 34 d of the pump body 34 , so that a thrust automatic centering force acts on the driven gear 46 so that the gap H 4 in the thickness direction of the driven gear 46 is made constant, that is, so as to move the centerline C 4 of the driven gear 46 toward the centerline C 3 of the pump chamber 42 .
  • the centerline C 4 of the driven gear 46 is a straight line intermediate between the side surfaces 46 e and 46 f of the driven gear 46 in its thickness direction, while the centerline C 3 of the pump chamber 42 is a straight line intermediate between the inner wall surfaces 34 d and 36 a of the pump chamber 42 in the thickness direction of the driven gear 46 .
  • the dynamic pressure P 3 which increases along a curve of the second order with an amount of a distance from the centerline C 5 of the drive gear 48 to the centerline C 3 of the pump chamber 42 is generated in the reduced gap between the side surface 48 c of the drive gear 48 and the inner circumferential surface 34 d of the pump body 34 , so that a thrust automatic centering force acts on the drive gear 48 so that the gap H 6 in the thickness direction of the drive gear 46 is made constant, that is, so as to move the centerline C 5 of the drive gear 48 toward the centerline C 3 of the pump chamber 42 .
  • the centerline C 5 of the drive gear 48 is a straight line intermediate between the side surfaces 48 b and 48 c of the drive gear 48 in its thickness direction.
  • FIG. 17 is the view indicating a relationship among the gap ratio m 1 , the dynamic pressure P 1 generated by the first dynamic pressure generating grooves 46 c having the gap ratio m 1 , and a fluid friction coefficient ⁇ 1 . It will be understood from FIG. 17 that the dynamic pressure P 1 changes as a function of the gap ratio m 1 , and is maximal when the gap ratio m 1 is in a predetermined range, while the fluid friction coefficient ⁇ 1 changes as a function of the gap ratio m 1 , and is minimal when the gap ratio m 1 is in a predetermined range.
  • the depth D 1 of the first dynamic pressure generating grooves 46 c in the radial direction of the driven gear 46 is determined such that the above-indicated gap ratio m 1 is held within a predetermined range in which the dynamic pressure P 1 generated by the first dynamic pressure generating grooves 46 c has a maximal value and in which the fluid friction coefficient ⁇ 1 generated on the basis of the first dynamic pressure generating grooves 46 c has a minimal value, as indicated in FIG. 17 . It will be understood from FIG.
  • the dynamic pressure P 1 is calculated according to Mathematical Equation 2 by replacing a non-dimensional pressure Kp by a value calculated according to Mathematical Equation 1 which is a three-dimensional Reynolds equation.
  • L represents the thickness of the driven gear 46 indicated in FIG. 4
  • B represents a length of the first dynamic pressure generating grooves 46 c in the form of wedges indicated in FIG. 5
  • U represents an outer circumferential flow velocity of the driven gear indicated in FIG. 5
  • ⁇ represents the viscosity of the working oil.
  • Mathematical Equation 3 is solved by differentiating Mathematical Equation 1 with respect to “x”.
  • the non-dimensional pressure Kp is calculated by mathematical analysis of Mathematical Equation 4 by a difference method.
  • the fluid friction coefficient ⁇ 1 is calculated according to Mathematical Equation 5.
  • K W and “K F0 ” in Mathematical Equation 5 are respectively calculated according to Mathematical Equations 6 and 7.
  • the depth D 2 of the second dynamic pressure generating grooves 46 g in the direction of thickness of the driven gear 46 is determined such that the above-indicated gap ratio m 2 is held within a predetermined range in which the dynamic pressure P 2 generated by the second dynamic pressure generating grooves 46 g has a maximal value and in which the fluid friction coefficient ⁇ 2 generated on the basis of the second dynamic pressure generating grooves 46 g has a minimal value.
  • the gap ratio m 2 is determined according to a relationship similar to the relationship indicated in FIG.
  • the fluid friction coefficient ⁇ 2 generated on the basis of the second dynamic pressure generating grooves 46 g is minimum while the dynamic pressure P 2 generated by the second dynamic pressure generating grooves 46 g is maximum, when the gap ratio m 2 is in a range between 1.5 and 4, preferably, between 2 and 3.
  • the depth D 3 of the third dynamic pressure generating grooves 48 d in the direction of thickness of the drive gear 48 is determined such that the above-indicated gap ratio m 3 is held within a predetermined range in which the dynamic pressure P 3 generated by the third dynamic pressure generating grooves 48 d has a maximal value and in which the fluid friction coefficient ⁇ 3 generated on the basis of the third dynamic pressure generating grooves 48 d has a minimal value.
  • the gap ratio m 3 is determined according to a relationship similar to the relationship indicated in FIG.
  • the fluid friction coefficient ⁇ 3 generated on the basis of the third dynamic pressure generating grooves 48 d is minimum while the dynamic pressure P 3 generated by the third dynamic pressure generating grooves 48 d is maximum, when the gap ratio m 3 is in a range between 1.5 and 4, preferably, between 2 and 3.
  • the fluid friction coefficient ⁇ 1 acting on the outer circumferential surface 46 b of the driven gear 46 is almost minimized, and the dynamic pressure P 1 generated by the first dynamic pressure generating grooves 46 c is almost maximized while the driven gear 46 is rotated, so that an automatic centering function of the driven gear 46 in its radial direction can be performed due to the automatic centering force in the radial direction while an increase of a fluid friction acting on the driven gear 46 is reduced or prevented.
  • the oil pump 10 is further configured such that the first dynamic pressure generating groove 46 c has the slant surface 46 d formed downwardly toward the bottom thereof in the outer circumferential surface 46 b of the driven gear 46 , such that the slant surface 46 d cooperates with the inner circumferential surface 34 c of the pump chamber 42 to define a wedge space. Accordingly, the fluid friction coefficient ⁇ 1 acting on the driven gear 46 is reduced, and the dynamic pressure P 1 generated by the first dynamic pressure generating grooves 46 c is increased.
  • the oil pump 10 is also configured such that the first dynamic pressure generating grooves 46 c are formed in the outer circumferential surface 46 b of the driven gear 46 such that the first dynamic pressure generating grooves 46 c are equiangularly spaced apart from each other, about the axis of rotation C 2 of the driven gear 46 . Accordingly, the function of automatic centering of the above-described driven gear 46 is effectively improved.
  • the oil pump 10 is further configured such that the first dynamic pressure generating grooves 46 c have the depth P 1 determined such that the gap ratio m 1 is held within the range between 2 and 3. Accordingly, the fluid friction coefficient ⁇ 1 acting on the driven gear 46 is almost minimized, and the dynamic pressure P 1 generated by the first dynamic pressure generating grooves 46 c is almost maximized.
  • the fluid friction coefficient ⁇ 2 acting on the opposite side surfaces 46 e and 46 f of the driven gear 46 is minimized, and the dynamic pressure P 2 generated by the second dynamic pressure generating grooves 46 g is maximized while the driven gear 46 is rotated, so that the driven gear 46 can be given a function of automatic centering in a thickness direction of the driven gear 46 , i.e., the direction of its axis of rotation C 2 due to the automatic centering force in the thrust direction while an increase of the fluid friction acting on the driven gear 46 is reduced or prevented.
  • the fluid friction coefficient ⁇ 3 acting on the opposite side surfaces 48 b and 48 c of the drive gear 48 is minimized, and the dynamic pressure P 3 generated by the third dynamic pressure generating grooves 48 d is maximized while the drive gear 48 is rotated, so that the drive gear 48 can be given a function of automatic centering in the direction of its axis of rotation C 1 while an increase of the fluid friction acting on the driven gear 48 is reduced or prevented.
  • the oil pump according to the present embodiment is different from the oil pump 10 according to the first embodiment described above, in the shape of first dynamic pressure generating grooves 46 i which is different from that of the first dynamic pressure generating grooves 46 c in the first embodiment.
  • the present oil pump is identical in construction with the oil pump 10 .
  • the gap ratio m 1 is a ratio of a distance h 1 of a gap H 1 from the bottom of the first dynamic pressure generating groove 46 i to the inner circumferential surface 34 c of the pump body 34 , to a distance h 2 of a gap H 2 from the outer circumferential surface 46 b of the driven gear 46 to the inner circumferential surface 34 c of the pump body 34 .
  • the first dynamic pressure generating groove 46 i is a groove in the form of a wedge formed in the outer circumferential surface 46 b of the driven gear 46 .
  • the first dynamic pressure generating groove 46 i formed in the outer circumferential surface 46 b of the driven gear 46 has a slant surface 46 j inclined upwardly from the bottom of the groove 46 i such that a distance from the slant surface 46 j to the inner circumferential surface 34 b of the pump chamber 42 decreases in the direction of rotation of the driven gear 46 opposite to the direction indicated by the arrow “b”.
  • the first dynamic pressure generating groove 46 c according to the first embodiment is qualitatively advantageous over the first dynamic pressure generating groove 46 i according to the present embodiment, from the standpoint of flow of the working fluid, because the working fluid suffers from partial separation due to abrupt enlargement of a gap of a fluid passage by the first dynamic pressure generating groove 46 i .
  • the optimum depth of the first dynamic pressure generating groove 46 i is on the order of ⁇ m in the oil pump 10 used for the automatic transmission 18 , so that there is not a significant quantitative advantage on the side of the first dynamic pressure generating groove 46 c in view of the fluid flow through the gap of this size.
  • the oil pump according to the present embodiment is different from the oil pump 10 according to the first embodiment described above, in the shape of first dynamic pressure generating grooves 46 k which is different from that of the first dynamic pressure generating grooves 46 c in the first embodiment.
  • the present oil pump is identical in construction with the oil pump 10 .
  • the gap ratio m 1 is a ratio of a distance h 1 of a gap H 1 from the bottom of the first dynamic pressure generating groove 46 k to the inner circumferential surface 34 c of the pump body 34 , to a distance h 2 of a gap H 2 from the outer circumferential surface 46 b of the driven gear 46 to the inner circumferential surface 34 c of the pump body 34 .
  • the first dynamic pressure generating groove 46 k is an elongate rectangular groove formed in the outer circumferential surface 46 b of the driven gear 46 . It is noted that the first dynamic pressure generating groove 46 c according to the first embodiment is qualitatively advantageous over the first dynamic pressure generating groove 46 k according to the present embodiment, from the standpoint of flow of the working fluid, because the working fluid suffers from partial separation due to abrupt enlargement of a gap of a fluid passage by the first dynamic pressure generating groove 46 k .
  • the optimum depth of the first dynamic pressure generating groove 46 k is on the order of ⁇ m in the oil pump 10 used for the automatic transmission 18 , so that there is not a significant quantitative advantage on the side of the first dynamic pressure generating groove 46 c in view of the fluid flow through the gap of this size.
  • An oil pump 66 according to the present embodiment is different from the oil pump 10 according to the first embodiment described above, in that the oil pump 66 is provided with a driven gear 68 not having the second dynamic pressure generating grooves 46 g provided in the first embodiment, and a drive gear 70 not having the third dynamic pressure generating grooves 48 d provided in the first embodiment, as shown in FIGS. 20 and 21 .
  • the oil pump 66 is identical in construction with the oil pump 10 .
  • the oil pump 66 constructed as described above has a lower degree of an automatic centering function of the driven gear 68 in its thickness direction, and a lower degree of an automatic centering function of the drive gear 70 in its thickness direction, than the automatic centering functions of the driven gear 46 and the drive gear 48 in the first embodiment.
  • the fluid friction coefficient ⁇ 1 acting on the driven gear 68 is minimized and the dynamic pressure P 1 generated by the first dynamic pressure generating grooves 46 c is maximized, while the driven gear 68 is rotated, so that the automatic centering function of the driven gear 68 in its radial direction is maximized, while an increase of the fluid friction acting on the driven gear 68 is reduced or prevented.
  • the gap ratios m 1 , m 2 and m 3 which are the ratios of the distances h 1 , h 3 and h 5 of the depths, i.e., gaps H 1 , H 3 and H 5 at the bottom of the first, second and third dynamic pressure generating grooves 46 c , 46 g and 48 d , to the distances h 2 , h 4 and h 6 of the gaps H 2 , H 4 and H 6 , are important, but the shapes per se of those grooves 46 c , 46 g and 48 d may be selected as desired.
  • the performance of the oil pump 10 provided for an automatic transmission does not substantially vary depending upon the shape of the dynamic pressure generating grooves the gaps of which are on the order of ⁇ m.
  • the second dynamic pressure generating grooves 46 g are formed in the opposite side surfaces 46 e and 46 f of the driven gear 46 .
  • the second dynamic pressure generating grooves 46 g may be formed in only one of the opposite side surfaces 46 e and 46 f .
  • the third dynamic pressure generating grooves 48 d which are formed in the opposite side surfaces 48 b and 48 c of the drive gear 48 may be formed in only one of the opposite side surfaces 48 b and 48 c.
  • the second dynamic pressure generating grooves 46 g have the shape as shown in FIG. 7 .
  • the second dynamic pressure generating grooves 46 g may have any other shapes, as long as the second dynamic pressure generating grooves 46 g assure the provision of sealing portions on the side surfaces 46 e and 46 f of the driven gear 46 . If the second dynamic pressure generating grooves 46 g are formed so as to extend through the sealing portions, the working oil tends to leak through the second dynamic pressure generating grooves 46 g , resulting in deterioration of the volumetric efficiency of the pump 10 .
  • oil pump 10 is used for the step-variable automatic transmission
  • the oil pump 10 may be used for a CVT or an automatic transmission for a hybrid vehicle.
US14/357,366 2011-11-10 2011-11-10 Internal-gear-type oil pump for vehicle Abandoned US20140314608A1 (en)

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JP5747999B2 (ja) 2015-07-15
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CN103917785B (zh) 2016-01-06
WO2013069144A1 (ja) 2013-05-16

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