INTERNAL-GEAR-TYPE OIL PUMP FOR VEHICLE

Abstract

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 said 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 said driven gear and which is disposed rotatably about an axis of rotation thereof eccentric with respect to an axis of rotation of said driven gear, to rotate said driven gear, characterized in that wherein: said driven gear has a plurality of first dynamic pressure generating grooves formed in local areas of its outer circumferential surface; and each of said 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 a distance of a gap from a bottom of said first dynamic pressure generating groove to said inner circumferential surface to a distance of a gap from the outer circumferential surface of said driven gear to said inner circumferential surface is held within a predetermined range in which a dynamic pressure generated by said first dynamic pressure generating grooves and changing as a function of said gap ratio has a maximal value and in which a fluid friction coefficient generated on the basis of said first dynamic pressure generating grooves and changing as a function of said gap ratio has a minimal value.

Description

TECHNICAL FIELD

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.

BACKGROUND ART

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. When the above-described driven gear is rotated, on the other hand, 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 circumferential directions toward the above-indicated portion of the gap. In this respect, it is noted that the above-indicated dynamic pressure acts on the outer circumferential surface of the above-described driven gear in a radially inward direction of the driven gear.

However, 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.

By the way, 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. In this internal-gear-type oil pump, 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. This higher dynamic pressure acting on the above-described driven gear promotes a function of automatic centering of the above-described driven gear, as compared with a driven gear of the vehicular internal-gear-type oil pump not having the above-indicated protrusions, so that the oscillation of the axis of rotation of the above-described driven gear is reduced. Alternatively, 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.

PRIOR ART DOCUMENTS

Patent Documents

• Patent Document 1: JP-2003-120550 A1
• Patent Document 2: JP-6-229448 A1
• Patent Document 3: JP-2011-052644 A1
• Patent Document 4: JP-2010-285979 A1
• Patent Document 5: JP-5-106632 A1

SUMMARY OF THE INVENTION

Object Achieved by the Invention

However, 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.

Means for Achieving the Object

The object indicated above is achieved according to the principle of the present invention, which provides 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 a gap from a bottom of the above-described first dynamic pressure generating groove to the above-described inner circumferential surface to a gap from the outer circumferential surface of the above-described driven gear to the above-described inner circumferential surface is held within a predetermined range in which a dynamic pressure generated by the first dynamic pressure generating grooves and changing as a function of the above-described gap ratio has a maximal value and in which a 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 a minimal value.

Advantages of the Invention

According to the vehicular internal-gear-type oil pump according to the present invention, (d) 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. Accordingly, 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.

According to one preferred form of this invention, 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. According to this form of the invention, 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.

According to another preferred form of the invention, 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.

According to a further preferred form of the invention, 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.

According to a still preferred form of the invention, (a) 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. According to this form of the invention, 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.

According to a yet further preferred form of the invention, (a) the above-described drive gear has a plurality of third dynamic pressure generating grooves formed in local areas of its opposite side surfaces, and (b) 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. According to this form of the invention, 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.

BRIEF DESCRIPTION OF THE DRAWINGS

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; and

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.

MODE FOR CARRYING OUT THE INVENTION

Referring to the drawings, embodiments of the present invention will be described in detail. It is to be understood that the drawings are simplified or transformed as needed for easy understanding, and do not necessarily accurately represent the dimensions and shapes of various elements of the embodiments.

First Embodiment

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.

As shown in FIG. 1, 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. In the thus constructed torque converter 16, 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 20a 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 20a 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. In the transmission casing 32, the input shaft 20 extends through a partition wall disposed between an accommodating space 32a accommodating the torque converter 16, and an accommodating space 32b 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 20a, and is fitted in a fitting hole 32c 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 34a 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, while the pump cover 36 is integrally fixed to the pump body 36 with second screws 40.

The pump body 34 has a cylindrical hole 34b which is open in the bottom surface of the fitting hole 34a and which has a smaller diameter and a larger depth than the fitting hole 34a. The cylindrical hole 34b has an axis O1 which is eccentric with respect to an axis of rotation C1 of the input shaft 22 and the sleeve 20a. In the oil pump 10, 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 20a, and which is defined by an inner circumferential surface 34c of the hole 34b, and inner wall surfaces 34d and 36a which are positioned at the respective opposite ends of the inner circumferential surface 34c in the direction of the axis O1. The pump chamber 42 has the axis O1 eccentric with respect to the axis of rotation C1 of the sleeve 20a.

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. In this respect, it is noted that FIG. 1 is the cross sectional view taken along lines I-I in FIG. 2. As shown in FIGS. 1 and 2, the oil pump 10 is provided with: an annular driven gear 46 which has internal teeth 46a, and an outer circumferential surface 46b opposed to the inner circumferential surface 34c defining the pump chamber 42, and which is rotatably supported by the inner circumferential surface 34c; and a drive gear 48 which has external teeth 48a engaging with the internal teeth 46a of the driven gear 46 and which is disposed rotatably about the axis of rotation C1 eccentric with respect to an axis of rotation C2 of the driven gear 46, to rotate the driven gear 46.

The drive gear 48 is fitted on the sleeve 20a such that drive gear 48 is rotatable with the sleeve 20a and movable relative to the sleeve 20a in the direction of the axis of rotation C1 of the sleeve 20a. When the sleeve 20a is rotated about the axis C1 in a direction indicated by an arrow “a” in FIG. 2, the driven gear 46 is rotated about the axis C2 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 48a of the drive gear 48 and the internal teeth 46a of the driven gear 46 the number of which is larger by one than that of the external teeth 48a 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 46a and the external teeth 48a 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, while 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 34a 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 34a of the pump body 34. It is noted that the 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, while the 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.

In the oil pump 10 constructed as described above, the drive gear 48 is rotated by the sleeve 20a, in the direction indicated by the arrow “a” in FIG. 2, and 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 46a and the external teeth 48a. 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.

As shown in FIGS. 4 and 5, the driven gear 46 has a plurality of first dynamic pressure generating grooves 46c formed in local areas of its outer circumferential surface 46b. As shown in FIG. 4, the plurality of first dynamic pressure generating grooves 46c formed in the outer circumferential surface 46b are equiangularly spaced apart from each other in the circumferential direction of the driven gear 46, about the axis of rotation C2 of the driven gear 46.

A depth D1 of each first dynamic pressure generating groove 46c in the radial direction of the driven gear 46, which is indicated in FIG. 5, is determined such that a gap ratio m1 (=h1/h2) is held within a predetermined range. The gap ratio m1 is a ratio of a distance h1 of a gap H1 from the bottom of the first dynamic pressure generating groove 46c to the inner circumferential surface 34c of the pump body 34 in the radial direction, to a distance h2 of a gap H2 from the outer circumferential surface 46b of the driven gear 46 to the inner circumferential surface 34c of the pump body 34 in the radial direction. It is noted that the depth D1 of the first dynamic pressure generating groove 46c in the radial direction of the driven gear 46 is equal to a difference (h1−h2), which is equal to the distance h1 minus the distance h2, as is apparent from FIG. 5. In the present embodiment, the distance h1 of the gap H1 is 125 μm, and the distance h2 of the gap H2 is 55 μm, while the depth D1 of the first dynamic pressure generating groove 46c is 70 μm.

As shown in FIG. 5, the first dynamic pressure generating groove 46c is a substantially triangular groove formed in the outer circumferential surface 46b of the driven gear 46. The first dynamic pressure generating groove 46c formed in the outer circumferential surface 46b of the driven gear 46 has a slant surface 46d inclined downwardly toward the bottom of the groove 46c in the direction of rotation of the driven gear 46 opposite to the direction indicated by the arrow “b”, and a slant surface 46h inclined from the bottom of the groove 46c such that a distance from the slant surface 46h to the inner circumferential surface 34b 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”. As shown in FIG. 5, the first dynamic pressure generating groove 46c has the slant surface 46d formed in the outer circumferential surface 46b of the driven gear 46 such that the slant surface 46d is inclined downwardly toward the bottom in the circumferential direction of the driven gear 46, and such that the slant surface 46d cooperates with the inner circumferential surface 34c of the pump chamber 42 to define a wedge space.

As shown in FIGS. 2-4, the driven gear 46 has a plurality of second dynamic pressure generating grooves 46g in the form of wedges formed in local areas of a side surface 46e thereof opposed to the inner wall surface 36a of the pump chamber 42 and a side surface 46f thereof opposed to the inner wall surface 34d of the pump chamber 42. As shown in FIG. 3, the second dynamic pressure generating grooves 46g have a shape as indicated in FIG. 7, and are equiangularly spaced apart from each other in the circumferential direction of the side surfaces 46e and 46f of the driven gear 46, about the axis of rotation C2 of the driven gear 46.

A depth D2 of each second dynamic pressure generating groove 46g in the direction of thickness of the driven gear 46, which is indicated in FIG. 6, is determined such that a gap ratio m2 (=h3/h4) is held within a predetermined range. The gap ratio m2 is a ratio of a distance h3 of a gap H3 from the bottom of the second dynamic pressure generating groove 46g to the inner wall surfaces 34d and 36a of the pump chamber 42, to a distance h4 of a gap H4 from the side surfaces 46e and 46f of the drive gear 46 to the inner wall surfaces 34d and 36a of the pump chamber 42. It is noted that the depth D2 of the second dynamic pressure generating groove 46g in the direction of thickness of the driven gear 46 is equal to a difference (h3−h4), which is equal to the distance h3 minus the distance h4, as is apparent from FIG. 6. In the present embodiment, the distance h3 of the gap H3 is 36 μm, and the distance h4 of the gap H4 is 16 μm, while the depth D2 of the second dynamic pressure generating groove 46g is 20 μm.

As shown in FIGS. 2 and 3, the drive gear 48 has a plurality of third dynamic pressure generating grooves 48d in the form of wedges formed in local areas of a side surface 48b thereof (shown in FIG. 1) opposed to the inner wall surface 36a of the pump chamber 42 and a side surface 48c thereof (shown in FIG. 1) opposed to the inner wall surface 34d of the pump chamber 42. As indicated in FIG. 3, the third dynamic pressure generating grooves 48d are equiangularly spaced apart from each other in the circumferential direction of the side surfaces 48b and 48c, about the axis of rotation C1 of the drive gear 48.

A depth D3 of each third dynamic pressure generating groove 48d in the direction of thickness of the drive gear 48, which is indicated in FIG. 8, is determined such that a gap ratio m3 (=h5/h6) is held within a predetermined range. The gap ratio m3 is a ratio of a distance h5 of a gap H5 from the bottom of the third dynamic pressure generating groove 48d to the inner wall surfaces 34d and 36a of the pump chamber 42, to a distance h6 of a gap H6 from the side surfaces 48b and 48c of the drive gear 48 to the inner wall surfaces 34d and 36a of the pump chamber 42. It is noted that the depth D3 of the third dynamic pressure generating groove 48d in the direction of thickness of the drive gear 48 is equal to a difference (h5−h6), which is equal to the distance h5 minus the distance h6, as is apparent from FIG. 8. In the present embodiment, the distance h5 of the gap H5 is 36 μm, and the distance h6 of the gap H6 is 16 μm, while the depth D3 of the third dynamic pressure generating groove 48d is 20 μm.

In the oil pump 10 constructed as described above, rotary motions of the driven gear 46 and the drive gear 48 by a rotary motion of the sleeve 20a cause flows of the working oil in the circumferential direction of the rotated driven and drive gears 46, 48, through the annular gap H2 formed between the outer circumferential surface 46b of the driven gear 46 and the inner circumferential surface 34c of the pump body 34, the pair of annular gaps H4 formed between the side surface 46f of the driven gear 46 and the inner wall surface 34d of the pump chamber 42 and between the side surface 46e of the driven gear 46 and the inner wall surface 36a of the pump chamber 42, and the pair of annular gaps H6 formed between the side surface 48c of the drive gear 48 and the inner wall surface 34d of the pump chamber 42 and between the side surface 48b of the drive gear 48 and the inner wall surface 36a of the pump chamber 42.

As a result, the gap H2 formed between the outer circumferential surface 46b of the driven gear 46 having the first dynamic pressure generating grooves 46c and the inner circumferential surface 34c 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) P1 of the working oil is maximized in a circumferential portion at which the gap distance is smallest. Further, the gap H4 formed between the side surface 46e of the driven gear 46 having the second dynamic pressure generating grooves 46g and the inner wall surface 36a of the pump chamber 42, and the gap H4 formed between the side surface 46f of the driven gear 46 having the second dynamic pressure generating grooves 46g and the inner wall surface 34d 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) P2 of the working oil is maximized in the circumferential portion at which the gap distance is smallest. Further, the gap H6 formed between the side surface 48b of the drive gear 48 having the third dynamic pressure generating grooves 48d and the inner wall surface 36a of the pump chamber 42, and the gap H6 formed between the side surface 48c of the drive gear 48 having the third dynamic pressure generating grooves 48d and the inner wall surface 34d 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) P3 of the working oil is maximized in circumferential portions at which the gap distance is smallest.

Accordingly, the dynamic pressure P1 generates a thrust force acting on the outer circumferential surface 46b of the driven gear 46 toward the axis of rotation C2 of the driven gear 46, as indicated in FIG. 9, so that the driven gear 46 is rotated with its outer circumferential surface 46b being held spaced apart from the inner circumferential surface 34c of the pump body 34, as indicated in FIG. 9. Further, the dynamic pressure P2 generates a thrust force which acts on the side surface 46f of the driven gear 46 toward the inner wall surface 34d of the pump chamber 42, and which also acts on the side surface 46e of the driven gear 46 toward the inner wall surface 36a of the pump chamber 42, as indicated in FIG. 10, so that the driven gear 46 is rotated with its side surfaces 46e and 46f being held spaced apart from the inner wall surfaces 34d and 36a of the pump chamber 42. Further, the dynamic pressure P3 generates a thrust force which acts on the side surface 48b of the drive gear 48 toward the inner wall surface 34d of the pump chamber 42, and which also acts on the side surface 48c of the drive gear 48 toward the inner wall surface 36a of the pump chamber 42, as indicated in FIG. 11, so that the drive gear 48 is rotated with its side surfaces 48b and 48c being held spaced apart from the inner wall surfaces 34d and 36a of the pump chamber 42.

FIG. 12 is the view indicating a relation between the distance h2 of the gap H2 between the outer circumferential surface 46b of the driven gear 46 and the inner circumferential surface 34d of the pump body 34, and the dynamic pressure P1 generated in the gap H2 filled with the viscous working oil, when the axis of rotation C2 of the driven gear 46 is positioned eccentrically moved with respect to the driven gear center position A1 coincident with the axis O1 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. In FIG. 12, “HIGH DYNAMIC PRESSURE ECCENTRIC SIDE” is one of opposite sides of the center position A1 of the driven gear 46, on which the gap H2 between the outer circumferential surface 46b of the driven gear 46 and the inner circumferential surface 34c of the pump body 34 is smaller so that the dynamic pressure P1 is higher, while “LOW DYNAMIC PRESSURE ECCENTRIC SIDE” is the other side of the center position A1 of the driven gear 46, on which the gap H2 between the outer circumferential surface 46b of the driven gear 46 and the inner circumferential surface 34c of the pump body 34 is larger so that the dynamic pressure P1 is lower.

Accordingly, when the driven gear 46 is positioned eccentrically in the radial direction, as indicated in FIG. 13, the dynamic pressure P1 which increases along a curve of the second order with an amount of eccentricity of the axis of rotation C2 of the driven gear 46 with respect to the axis O1 of the pump chamber 42 is generated in the reduced gap between the outer circumferential surface 46b of the driven gear 46 and the inner circumferential surface 34d of the pump body 34, so that a radial automatic centering force acts on the driven gear 46 so that the gap H2 is made constant in the circumferential direction of the driven gear 46, that is, so as to move the axis of rotation C2 of the driven gear 46 toward the axis O1 of the pump chamber 42.

Therefore, even if boundary lubrication takes place between the outer circumferential surface 46b of the driven gear 46 and the inner circumferential surface 34c of the pump body 34 due to eccentric positioning of the driven gear 46, the above-indicated radial automatic centering force causes a change of the lubricating condition from the boundary lubrication back to fluid lubrication. Further, the automatic centering of the driven gear 46 permits an increase of the distance h2 of the gap H2, and a decrease of a viscosity stress (τ=η(du/dy) of the above-indicated fluid lubrication. In this respect, it is noted that the side surfaces 46e and 46f of the driven gear 46 which define the thickness of the driven gear 46, and the side surfaces 48b and 48c of the drive gear 48 which define the thickness of the drive gear 48 respectively have the second dynamic pressure generating grooves 46g and the third dynamic pressure generating grooves 48d, which generate a thrust automatic centering force, like the first dynamic pressure generating grooves 46c.

When the driven gear 46 is moved in its thickness direction such that a centerline C4 of the driven gear 46 is spaced from a centerline C3 of the pump chamber 42, as indicated in FIG. 14, the dynamic pressure P2 which increases along a curve of the second order with an amount of a distance (i.e., an amount of biasing) from the centerline C4 of the driven gear 46 to the centerline C3 of the pump chamber 42 is generated in the reduced gap between the side surface 46e of the driven gear 46 and the inner circumferential surface 34d of the pump body 34, so that a thrust automatic centering force acts on the driven gear 46 so that the gap H4 in the thickness direction of the driven gear 46 is made constant, that is, so as to move the centerline C4 of the driven gear 46 toward the centerline C3 of the pump chamber 42. The centerline C4 of the driven gear 46 is a straight line intermediate between the side surfaces 46e and 46f of the driven gear 46 in its thickness direction, while the centerline C3 of the pump chamber 42 is a straight line intermediate between the inner wall surfaces 34d and 36a of the pump chamber 42 in the thickness direction of the driven gear 46.

When the centerline C4 of the driven gear 46 is inclined with respect to the centerline C3 of the pump chamber 42, as indicated in FIG. 15, the comparatively high dynamic pressure P2 is generated in a reduced upper portion of the gap between the side surface 46e of the driven gear 46 and the inner wall surface 34d of the pump chamber 42, and a reduced lower portion of the gap between the side surface 46f of the driven gear 46 and the inner wall surface 36a of the pump chamber 42, so that a thrust automatic centering force acts on the driven gear 46 so that the gap H4 in the thickness direction of the driven gear 46 is made constant, that is, so as to move the centerline C4 of the driven gear 46 toward the centerline C3 of the pump chamber 42.

When the drive gear 48 is moved in its thickness direction such that a centerline C5 of the drive gear 48 is spaced from the centerline C3 of the pump chamber 42, as indicated in FIG. 16, the dynamic pressure P3 which increases along a curve of the second order with an amount of a distance from the centerline C5 of the drive gear 48 to the centerline C3 of the pump chamber 42 is generated in the reduced gap between the side surface 48c of the drive gear 48 and the inner circumferential surface 34d of the pump body 34, so that a thrust automatic centering force acts on the drive gear 48 so that the gap H6 in the thickness direction of the drive gear 46 is made constant, that is, so as to move the centerline C5 of the drive gear 48 toward the centerline C3 of the pump chamber 42. The centerline C5 of the drive gear 48 is a straight line intermediate between the side surfaces 48b and 48c of the drive gear 48 in its thickness direction.

FIG. 17 is the view indicating a relationship among the gap ratio m1, the dynamic pressure P1 generated by the first dynamic pressure generating grooves 46c having the gap ratio m1, and a fluid friction coefficient μ1. It will be understood from FIG. 17 that the dynamic pressure P1 changes as a function of the gap ratio m1, and is maximal when the gap ratio m1 is in a predetermined range, while the fluid friction coefficient μ1 changes as a function of the gap ratio m1, and is minimal when the gap ratio m1 is in a predetermined range.

The depth D1 of the first dynamic pressure generating grooves 46c in the radial direction of the driven gear 46 is determined such that the above-indicated gap ratio m1 is held within a predetermined range in which the dynamic pressure P1 generated by the first dynamic pressure generating grooves 46c has a maximal value and in which the fluid friction coefficient μ1 generated on the basis of the first dynamic pressure generating grooves 46c has a minimal value, as indicated in FIG. 17. It will be understood from FIG. 17 that the fluid friction coefficient μ1 generated on the basis of the first dynamic pressure generating grooves 46c is minimum while the dynamic pressure P1 generated by the first dynamic pressure generating grooves 46c is maximum, when the gap ratio m1 is in a range between 1.5 and 4, preferably, between 2 and 3.

It is noted that the dynamic pressure P1 and the fluid friction coefficient μ1 corresponding to the gap ratio m1, which are indicated in FIG. 17, are calculated as described below.

The dynamic pressure P1 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. It is noted that “L” represents the thickness of the driven gear 46 indicated in FIG. 4, “B” represents a length of the first dynamic pressure generating grooves 46c in the form of wedges indicated in FIG. 5, and “U” represents an outer circumferential flow velocity of the driven gear indicated in FIG. 5, and “η” represents the viscosity of the working oil.

The method of calculation of the non-dimensional pressure Kp will be described. Initially, Mathematical Equation 3 is solved by differentiating Mathematical Equation 1 with respect to “x”. Then, Mathematical Equation 4 is solved by inserting therein a non-dimensional film thickness H (=h/h2), a non-dimensional coordinate value X (=x/B), a non-dimensional coordinate value Z (=z/L), a non-dimensional pressure P (=ph2²)/(ηUB), and an oil film formation equation dH/dX=1−m. The non-dimensional pressure Kp is calculated by mathematical analysis of Mathematical Equation 4 by a difference method.

The fluid friction coefficient μ1 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.

$\begin{array}{cc}\frac{\partial }{\partial x}\left(h^3\frac{\partial p}{\partial x}\right)+h^3\frac{{\partial }^2p}{\partial z^2}-6\eta U\frac{h}{x}& \left[\mathrm{Mathematical}\mathrm{Equation}1\right]\\ P1=\frac{\eta \mathrm{UB}}{h^2}\times \mathrm{Kp}& \left[\mathrm{Mathematical}\mathrm{Equation}2\right]\\ h^3\frac{{\partial }^2p}{\partial x^2}+3h^2\frac{h}{x}\frac{\partial p}{\partial x}+h^3\frac{{\partial }^2p}{\partial z^2}=6\eta U\frac{h}{x}& \left[\mathrm{Mathematical}\mathrm{Equation}3\right]\\ \frac{{\partial }^2p}{\partial x^2}+\frac{3\left(1-m\right)}{H}\frac{\partial p}{\partial X}+{\left(\frac{1}{L/B}\right)}^2\frac{{\partial }^2p}{\partial z^2}=\frac{6\left(1-m\right)}{H^3}& \left[\mathrm{Mathematical}\mathrm{Equation}4\right]\\ \mu 1=\frac{h_2}{B}\frac{K_{F0}}{K_W}& \left[\mathrm{Mathematical}\mathrm{Equation}5\right]\\ K_W=\frac{6}{{\left(m-1\right)}^2}\left\{\ln m-\frac{2\left(1-m\right)}{m+1}\right\}& \left[\mathrm{Mathematical}\mathrm{Equation}6\right]\\ K_{F0}=\frac{1}{m-1}\left\{4\ln m-\frac{6\left(1-m\right)}{m+1}\right\}& \left[\mathrm{Mathematical}\mathrm{Equation}7\right]\end{array}$

In the present embodiment, the depth D2 of the second dynamic pressure generating grooves 46g in the direction of thickness of the driven gear 46 is determined such that the above-indicated gap ratio m2 is held within a predetermined range in which the dynamic pressure P2 generated by the second dynamic pressure generating grooves 46g has a maximal value and in which the fluid friction coefficient μ2 generated on the basis of the second dynamic pressure generating grooves 46g has a minimal value. In the present embodiment, the gap ratio m2 is determined according to a relationship similar to the relationship indicated in FIG. 17, among the gap ratio m2, the dynamic pressure P2 generated by the second dynamic pressure generating grooves 46g having the depth D2 determined by the gap ratio m2, and the fluid friction coefficient μ2 generated on the basis of the second dynamic pressure generating grooves 46g, wherein the dynamic pressure P2 and the fluid friction coefficient μ2 are calculated according to the above-described methods of calculation of the dynamic pressure P1 and the fluid friction coefficient μ1. It is noted that the fluid friction coefficient μ2 generated on the basis of the second dynamic pressure generating grooves 46g is minimum while the dynamic pressure P2 generated by the second dynamic pressure generating grooves 46g is maximum, when the gap ratio m2 is in a range between 1.5 and 4, preferably, between 2 and 3.

In the present embodiment, the depth D3 of the third dynamic pressure generating grooves 48d in the direction of thickness of the drive gear 48 is determined such that the above-indicated gap ratio m3 is held within a predetermined range in which the dynamic pressure P3 generated by the third dynamic pressure generating grooves 48d has a maximal value and in which the fluid friction coefficient μ3 generated on the basis of the third dynamic pressure generating grooves 48d has a minimal value. In the present embodiment, the gap ratio m3 is determined according to a relationship similar to the relationship indicated in FIG. 17, among the gap ratio m3, the dynamic pressure P3 generated by the third dynamic pressure generating grooves 48d having the depth D3 determined by the gap ratio m3, and the fluid friction coefficient μ3 generated on the basis of the third dynamic pressure generating grooves 48d, wherein the dynamic pressure P3 and the fluid friction coefficient μ3 are calculated according to the above-described methods of calculation of the dynamic pressure P1 and the fluid friction coefficient μ1. It is noted that the fluid friction coefficient μ3 generated on the basis of the third dynamic pressure generating grooves 48d is minimum while the dynamic pressure P3 generated by the third dynamic pressure generating grooves 48d is maximum, when the gap ratio m3 is in a range between 1.5 and 4, preferably, between 2 and 3.

The internal-gear-type oil pump 10 according to the present embodiment is configured such that the driven gear 46 has the plurality of first dynamic pressure generating grooves 46c formed in the local areas of its outer circumferential surface 46b, and each of the first dynamic pressure generating groove 46c has the depth D1 in the radial direction of the driven gear 46, which depth D1 is determined such that the gap ratio m1 (=h1/h2) which is the ratio of the distance h1 of the gap H1 from the bottom of the first dynamic pressure generating groove 46c to the inner circumferential surface 34b of the pump body 34 to the distance h2 of the gap H2 from the outer circumferential surface 46b of the driven gear 46 to the inner circumferential surface 34c of the pump body 34 is held within the predetermined range, for example, between 1.5 and 4, preferably between 2 and 3, in which the dynamic pressure P1 generated by the first dynamic pressure generating grooves 46c and changing as a function of the gap ratio m1 has the maximal value and in which the fluid friction coefficient μ1 generated on the basis of the first dynamic pressure generating grooves 46c and changing as a function of the gap ratio m1 has the minimal value. Accordingly, the fluid friction coefficient μ1 acting on the outer circumferential surface 46b of the driven gear 46 is almost minimized, and the dynamic pressure P1 generated by the first dynamic pressure generating grooves 46c 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 according to the present embodiment is further configured such that the first dynamic pressure generating groove 46c has the slant surface 46d formed downwardly toward the bottom thereof in the outer circumferential surface 46b of the driven gear 46, such that the slant surface 46d cooperates with the inner circumferential surface 34c of the pump chamber 42 to define a wedge space. Accordingly, the fluid friction coefficient μ1 acting on the driven gear 46 is reduced, and the dynamic pressure P1 generated by the first dynamic pressure generating grooves 46c is increased.

The oil pump 10 according to the present embodiment is also configured such that the first dynamic pressure generating grooves 46c are formed in the outer circumferential surface 46b of the driven gear 46 such that the first dynamic pressure generating grooves 46c are equiangularly spaced apart from each other, about the axis of rotation C2 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 according to the present embodiment is further configured such that the first dynamic pressure generating grooves 46c have the depth P1 determined such that the gap ratio m1 is held within the range between 2 and 3. Accordingly, the fluid friction coefficient μ1 acting on the driven gear 46 is almost minimized, and the dynamic pressure P1 generated by the first dynamic pressure generating grooves 46c is almost maximized.

The oil pump 10 according to the present embodiment is also configured such that the driven gear 46 has the plurality of second dynamic pressure generating grooves 46g formed in the local areas of its opposite side surfaces 46f and 46e, and such that each of the second dynamic pressure generating grooves 46g has the depth D2 in the thickness direction of the driven gear 46, which depth D2 is determined such that the gap ratio m2 (=h3/h4) which is the ratio of the distance h3 of the gap H3 from the bottom of the second dynamic pressure generating groove 46g to the inner wall surfaces 36a and 34d of the pump chamber 42, to the distance h4 of the gap H4 from the side surfaces 46f and 46e of the driven gear 46 to the inner wall surfaces 36a and 34d of the pump chamber 42 is held within the predetermined range in which the dynamic pressure P2 generated by the second dynamic pressure generating grooves 46g has the maximal value and in which the fluid friction coefficient μ2 generated on the basis of the second dynamic pressure generating grooves 46g has the minimal value. Accordingly, the fluid friction coefficient μ2 acting on the opposite side surfaces 46e and 46f of the driven gear 46 is minimized, and the dynamic pressure P2 generated by the second dynamic pressure generating grooves 46g 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 C2 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 oil pump 10 according to the present embodiment is further configured such that the driven gear 48 has the plurality of third dynamic pressure generating grooves 48d formed in the local areas of its opposite side surfaces 48b and 48c, and such that each of the third dynamic pressure generating grooves 48d has the depth D3 in the thickness direction of the drive gear 48, which depth D3 is determined such that the gap ratio m3 (=h5/h6) which is the ratio of the distance h5 of the gap H5 from the bottom of the third dynamic pressure generating groove 48d to the inner wall surfaces 36a and 34d of the pump chamber 42, to the distance h6 of the gap H6 from the side surfaces 48b and 48c of the drive gear 48 to the inner wall surfaces 36a and 34d of the pump chamber 42 is held within the predetermined range in which the dynamic pressure P3 generated by the third dynamic pressure generating grooves 48d has the maximal value and in which the fluid friction coefficient μ3 generated on the basis of the third dynamic pressure generating grooves 48d has the minimal value. Accordingly, the fluid friction coefficient μ3 acting on the opposite side surfaces 48b and 48c of the drive gear 48 is minimized, and the dynamic pressure P3 generated by the third dynamic pressure generating grooves 48d 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 C1 while an increase of the fluid friction acting on the driven gear 48 is reduced or prevented.

Second Embodiment

Other embodiments of this invention will be described. It is to be understood that the same reference signs will be used to identify the corresponding elements in the different embodiments, which will not be described redundantly.

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 46i which is different from that of the first dynamic pressure generating grooves 46c in the first embodiment. In the other aspects, the present oil pump is identical in construction with the oil pump 10.

A depth D1 of each first dynamic pressure generating groove 46i in the radial direction of the driven gear 46, which is indicated in FIG. 18, is determined such that a gap ratio m1 (=h1/h2) is held within a predetermined range, as in the first embodiment. The gap ratio m1 is a ratio of a distance h1 of a gap H1 from the bottom of the first dynamic pressure generating groove 46i to the inner circumferential surface 34c of the pump body 34, to a distance h2 of a gap H2 from the outer circumferential surface 46b of the driven gear 46 to the inner circumferential surface 34c of the pump body 34.

As shown in FIG. 18, the first dynamic pressure generating groove 46i is a groove in the form of a wedge formed in the outer circumferential surface 46b of the driven gear 46. The first dynamic pressure generating groove 46i formed in the outer circumferential surface 46b of the driven gear 46 has a slant surface 46j inclined upwardly from the bottom of the groove 46i such that a distance from the slant surface 46j to the inner circumferential surface 34b 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”. It is noted that the first dynamic pressure generating groove 46c according to the first embodiment is qualitatively advantageous over the first dynamic pressure generating groove 46i 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 46i. However, the optimum depth of the first dynamic pressure generating groove 46i is on the order of μm 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 46c in view of the fluid flow through the gap of this size.

Third Embodiment

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 46k which is different from that of the first dynamic pressure generating grooves 46c in the first embodiment. In the other aspects, the present oil pump is identical in construction with the oil pump 10.

A depth D1 of each first dynamic pressure generating groove 46k in the radial direction of the driven gear 46, which is indicated in FIG. 19, is determined such that a gap ratio m1 (=h1/h2) is held within a predetermined range, as in the first embodiment. The gap ratio m1 is a ratio of a distance h1 of a gap H1 from the bottom of the first dynamic pressure generating groove 46k to the inner circumferential surface 34c of the pump body 34, to a distance h2 of a gap H2 from the outer circumferential surface 46b of the driven gear 46 to the inner circumferential surface 34c of the pump body 34.

As shown in FIG. 19, the first dynamic pressure generating groove 46k is an elongate rectangular groove formed in the outer circumferential surface 46b of the driven gear 46. It is noted that the first dynamic pressure generating groove 46c according to the first embodiment is qualitatively advantageous over the first dynamic pressure generating groove 46k 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 46k. However, the optimum depth of the first dynamic pressure generating groove 46k is on the order of μm 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 46c in view of the fluid flow through the gap of this size.

Fourth Embodiment

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 46g provided in the first embodiment, and a drive gear 70 not having the third dynamic pressure generating grooves 48d provided in the first embodiment, as shown in FIGS. 20 and 21. In the other aspects, 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. As in the case of the driven gear 46 according to the first embodiment, however, the fluid friction coefficient μ1 acting on the driven gear 68 is minimized and the dynamic pressure P1 generated by the first dynamic pressure generating grooves 46c 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.

While the embodiments of this invention have been described in detail by reference to the drawings, it is to be understood that the present invention may be otherwise embodied.

In the oil pump 10 according to the illustrated embodiments, the gap ratios m1, m2 and m3, which are the ratios of the distances h1, h3 and h5 of the depths, i.e., gaps H1, H3 and H5 at the bottom of the first, second and third dynamic pressure generating grooves 46c, 46g and 48d, to the distances h2, h4 and h6 of the gaps H2, H4 and H6, are important, but the shapes per se of those grooves 46c, 46g and 48d 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 μm.

In the oil pump 10 according to the illustrated embodiments, the second dynamic pressure generating grooves 46g are formed in the opposite side surfaces 46e and 46f of the driven gear 46. However, the second dynamic pressure generating grooves 46g may be formed in only one of the opposite side surfaces 46e and 46f. Similarly, the third dynamic pressure generating grooves 48d which are formed in the opposite side surfaces 48b and 48c of the drive gear 48 may be formed in only one of the opposite side surfaces 48b and 48c.

In the oil pump 10 according to the illustrated embodiments, the second dynamic pressure generating grooves 46g have the shape as shown in FIG. 7. However, the second dynamic pressure generating grooves 46g may have any other shapes, as long as the second dynamic pressure generating grooves 46g assure the provision of sealing portions on the side surfaces 46e and 46f of the driven gear 46. If the second dynamic pressure generating grooves 46g are formed so as to extend through the sealing portions, the working oil tends to leak through the second dynamic pressure generating grooves 46g, resulting in deterioration of the volumetric efficiency of the pump 10.

While the oil pump 10 according to the illustrated embodiments 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.

While the embodiments of this invention have been described for illustrative purpose only, it is to be understood that the invention may be embodied with various changes and improvements which may occur to those skilled in the art.

NOMENCLATURE OF REFERENCE SIGNS

• 10, 66: Oil pump
• 34: Pump body
• 34 c: Inner circumferential surface
• 34 d: Inner wall surface
• 36: Pump cover
• 36 a: Inner wall surface
• 42: Pump chamber
• 46: Driven gear
• 46 a: Internal teeth
• 46 b: Outer circumferential surface
• 46 c, 46i, 46k: First dynamic pressure generating grooves
• 46 d: Slant surface
• 46 e, 46f: Side surfaces
• 46 g: Second dynamic pressure generating grooves
• 48: Drive gear
• 48 a: External teeth
• 48 b, 48c: Side surface
• 48 d: Third dynamic pressure generating grooves
• C1: Axis of rotation of drive gear
• C2: Axis of rotation of driven gear
• D1: Radial depth of first dynamic pressure generating grooves
• D2: Radial depth of second dynamic pressure generating grooves
• D3: Radial depth of third dynamic pressure generating grooves
• H1-H6: Gaps
• m1-m2: Gap ratios
• P1-P3: Dynamic pressures (generating dynamic pressures)
• μ1-μ3: Fluid friction coefficients

Claims

1. 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 said 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 said driven gear and which is disposed rotatably about an axis of rotation thereof eccentric with respect to an axis of rotation of said driven gear, to rotate said driven gear, wherein: said driven gear has a plurality of first dynamic pressure generating grooves formed in local areas of its outer circumferential surface; andeach of said 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 a distance of a gap from a bottom of said first dynamic pressure generating groove to said inner circumferential surface to a distance of a gap from the outer circumferential surface of said driven gear to said inner circumferential surface is held within a predetermined range in which a dynamic pressure generated by said first dynamic pressure generating grooves and changing as a function of said gap ratio has a maximal value and in which a fluid friction coefficient generated on the basis of said first dynamic pressure generating grooves and changing as a function of said gap ratio has a minimal value.

2. The vehicular internal-gear-type oil pump according to claim 1, wherein said first dynamic pressure generating groove has a slant surface formed downwardly toward the bottom thereof in the outer circumferential surface of said driven gear, such that said slant surface cooperates with the inner circumferential surface of said pump chamber to define a wedge space.

3. The vehicular internal-gear-type oil pump according to claim 1, wherein said plurality of first dynamic pressure generating grooves are formed in the outer circumferential surface of said driven gear such that the first dynamic pressure generating grooves are equiangularly spaced apart from each other, about the axis of rotation of said driven gear.

4. The vehicular internal-gear-type oil pump according to claim 1, wherein said first dynamic pressure generating grooves have a depth such that said gap ratio is held within a range between 2 and 3.

5. The vehicular integral-gear-type oil pump according to claim 1, wherein said driven gear has a plurality of second dynamic pressure generating grooves formed in local areas of its opposite side surfaces, each of said second dynamic pressure generating grooves having 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 distance of a gap from a bottom of said second dynamic pressure generating groove to inner wall surfaces of said pump chamber, to a distance of a gap from the side surfaces of said driven gear to the inner wall surfaces of said pump chamber is held within a predetermined range in which a dynamic pressure generated by said second dynamic pressure generating grooves has a maximal value and in which a fluid friction coefficient generated on the basis of said second dynamic pressure generating grooves has a minimal value.

6. The vehicular internal-gear-type oil pump according to claim 1, wherein said drive gear has a plurality of third dynamic pressure generating grooves formed in local areas of its opposite side surfaces, each of said third dynamic pressure generating grooves having 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 distance of a gap from a bottom of said third dynamic pressure generating groove to inner wall surfaces of said pump chamber, to a distance of a gap from the side surfaces of said drive gear to the inner wall surfaces of said pump chamber is held within a predetermined range in which a dynamic pressure generated by said third dynamic pressure generating grooves has a maximal value and in which a fluid friction coefficient generated on the basis of said third dynamic pressure generating grooves has a minimal value.