U.S. patent application number 13/011972 was filed with the patent office on 2011-08-04 for vane pump.
This patent application is currently assigned to Hitachi Automotive Systems, Ltd.. Invention is credited to Hideaki Ohnishi, Koji Saga, Yasushi WATANABE.
Application Number | 20110189043 13/011972 |
Document ID | / |
Family ID | 44341864 |
Filed Date | 2011-08-04 |
United States Patent
Application |
20110189043 |
Kind Code |
A1 |
WATANABE; Yasushi ; et
al. |
August 4, 2011 |
VANE PUMP
Abstract
A vane pump in which oil is drawn from at least one side, in an
axial direction, of a cam ring and is discharged from at least one
side, in the axial direction, of the cam ring, has a portion
defined by a groove formed on an inner circumference surface of the
cam ring. The groove extends along a circumferential direction of
the inner circumference surface of the cam ring and is arranged in
a position including a middle of an axial direction width on the
inner circumference surface. The groove is arranged in an oil
suction section or an oil discharge section of the cam ring.
Inventors: |
WATANABE; Yasushi;
(Aiko-gun, JP) ; Saga; Koji; (Ebina-shi, JP)
; Ohnishi; Hideaki; (Atsugi-shi, JP) |
Assignee: |
Hitachi Automotive Systems,
Ltd.
|
Family ID: |
44341864 |
Appl. No.: |
13/011972 |
Filed: |
January 24, 2011 |
Current U.S.
Class: |
418/133 |
Current CPC
Class: |
F04C 14/22 20130101;
F04C 2/3446 20130101; F04C 15/06 20130101; F01C 19/08 20130101 |
Class at
Publication: |
418/133 |
International
Class: |
F01C 19/08 20060101
F01C019/08 |
Foreign Application Data
Date |
Code |
Application Number |
Jan 29, 2010 |
JP |
2010-018201 |
Claims
1. A vane pump in which oil is drawn from at least one side, in an
axial direction, of a cam ring and is discharged from at least one
side, in the axial direction, of the cam ring, the vane pump
comprising: a portion defined by a groove formed on an inner
circumference surface of the cam ring, the groove extending along a
circumferential direction of the inner circumference surface and
arranged in a position including a middle of an axial direction
width on the inner circumference surface in an oil suction section
or an oil discharge section of the cam ring.
2. The vane pump as claimed in claim 1, wherein: the groove is
formed so that a depth of the groove is gradually shallower from a
middle, in the circumferential direction, of the groove to both
ends of the groove.
3. The vane pump as claimed in claim 1, wherein: the groove is
formed by cutting.
4. The vane pump as claimed in claim 1, wherein: the groove is
provided in both of the oil suction section and the oil discharge
section on the inner circumference surface of the cam ring.
5. The vane pump as claimed in claim 1, wherein: the vane pump is a
vane pump in which the oil is drawn from both sides, in the axial
direction, of the cam ring and is discharged from both sides, in
the axial direction, of the cam ring, and the groove is formed at a
circumferential portion except both circumferential edge sides of
the axial direction width on the inner circumference surface of the
cam ring.
6. The vane pump as claimed in claim 5, wherein: widths of the both
circumferential edge sides of the axial direction width on the
inner circumference surface, where no groove exists, are set to be
substantially the same.
7. A vane pump comprising: a rotor rotatably driven; a plurality of
vanes arranged at an outer circumference of the rotor and
extending/retracting in a radial direction; a cam ring housing, at
an inner circumferential side thereof, the rotor and the vanes, the
extending/retracting movement of the vanes occurring by the
rotation of the rotor and sliding contact of each top end edge of
the vanes with an inner circumference surface of the cam ring; a
housing housing the cam ring inside the housing and defining a
plurality of pump chambers by the housing, the vanes, the cam ring
and the rotor; an inlet port provided at least one side of both
side walls of the housing which respectively face axial direction
both sides of the cam ring and opening to a section where the vanes
extend; an outlet port provided at least one side of the both side
walls of the housing which respectively face the axial direction
both sides of the cam ring and opening to a section where the vanes
retract; and a communication portion formed at a circumferential
portion except both circumferential edge sides of an axial
direction width on the inner circumference surface of the cam ring
in the section where the vanes extend and connecting the pump
chambers.
8. The vane pump as claimed in claim 7, wherein: the communication
portion is formed by a groove that extends in a circumferential
direction of the inner circumference surface of the cam ring.
9. The vane pump as claimed in claim 8, wherein: the groove is
formed so that depths of both ends of the groove is shallower than
that of a middle, in the circumferential direction, of the
groove.
10. The vane pump as claimed in claim 8, wherein: the cam ring
moves with respect to the rotor and an eccentric amount of the cam
ring with respect to the rotor changes, and an oil amount
discharged from the outlet port is varied by the change of the
eccentric amount.
11. The vane pump as claimed in claim 10, wherein: the cam ring is
forced in a direction in which the eccentric amount with respect to
the rotor becomes large by a forcing member, and the cam ring is
moved in an opposite direction against the force of the forcing
member, for controlling the oil discharge amount.
12. The vane pump as claimed in claim 11, wherein: the cam ring is
configured to move in the opposite direction by receiving a
pressure of the outlet port.
13. The vane pump as claimed in claim 7, wherein: the communication
portion is formed by a communication conduit.
14. A vane pump comprising: a rotor rotatably driven and having, at
an outer circumference thereof, a plurality of opening slots; a
plurality of vanes provided in the respective slots; a cam ring
housing, at an inner circumferential side thereof, the rotor and
the vanes, extending/retracting movement of the vanes at the outer
circumference of the rotor occurring by the rotation of the rotor;
a housing housing the cam ring inside the housing and defining a
plurality of pump chambers by the housing, the vanes, the cam ring
and the rotor; an inlet port provided at least one side of both
side walls of the housing which respectively face axial direction
both sides of the cam ring and opening to a section where a volume
of the pump chamber increases; an outlet port provided at least one
side of the both side walls of the housing which respectively face
the axial direction both sides of the cam ring and opening to a
section where the volume of the pump chamber decreases; and a
communication portion formed at a circumferential portion except
both circumferential edge sides of an axial direction width on an
inner circumference surface of the cam ring in the section where
the volume of the pump chamber decreases and connecting the pump
chambers.
15. The vane pump as claimed in claim 14, wherein: the cam ring
moves with respect to the rotor and an eccentric amount of the cam
ring with respect to the rotor changes, and an oil amount
discharged from the outlet port is varied by the change of the
eccentric amount.
Description
BACKGROUND OF THE INVENTION
[0001] The present invention relates to a vane pump which supplies
oil to, for example, each sliding part in an internal combustion
engine of a vehicle and a variable valve timing control apparatus
that variably controls open/close timing of valves of the
engine.
[0002] This kind of related art vane pump has been disclosed in
Japanese Patent Provisional Publication to tokuhyou No. 2008-524500
(hereinafter is referred to as "JPA.sub.--2008524500")
corresponding to International Publication No. WO2006/066405.
[0003] In this related art vane pump, an inlet port and an outlet
port are each provided on both side walls of a housing where both
end surfaces, in an axial direction, of a rotor and vanes make
sliding contact with the both side walls, and the oil drawn from
the inlet port to each pump chamber is pressurized and discharged
to the outlet port.
SUMMARY OF THE INVENTION
[0004] In this related art vane pump, however, in a case where the
pump rotates at high speed, difference in pressure in a suction
section or a discharge section arises between both circumferential
edge side portions, in an axial direction, of an inner
circumference surface of a cam ring and a circumferential middle
portion of the inner circumference surface of the cam ring. For
this reason, there is a possibility that a stable pump operation
cannot be achieved.
[0005] It is therefore an object of the present invention to
provide a vane pump which is capable of achieving a stable action
of the cam ring even when the pump rotates at high speed.
[0006] According to one aspect of the present invention, a vane
pump in which oil is drawn from at least one side, in an axial
direction, of a cam ring and is discharged from at least one side,
in the axial direction, of the cam ring, the vane pump comprises: a
portion defined by a groove formed on an inner circumference
surface of the cam ring, the groove extending along a
circumferential direction of the inner circumference surface and
arranged in a position including a middle of an axial direction
width on the inner circumference surface in an oil suction section
or an oil discharge section of the cam ring.
[0007] According to another aspect of the present invention, a vane
pump comprises: a rotor rotatably driven; a plurality of vanes
arranged at an outer circumference of the rotor and
extending/retracting in a radial direction; a cam ring housing, at
an inner circumferential side thereof, the rotor and the vanes, the
extending/retracting movement of the vanes occurring by the
rotation of the rotor and sliding contact of each top end edge of
the vanes with an inner circumference surface of the cam ring; a
housing housing the cam ring inside the housing and defining a
plurality of pump chambers by the housing, the vanes, the cam ring
and the rotor; an inlet port provided at least one side of both
side walls of the housing which respectively face axial direction
both sides of the cam ring and opening to a section where the vanes
extend; an outlet port provided at least one side of the both side
walls of the housing which respectively face the axial direction
both sides of the cam ring and opening to a section where the vanes
retract; and a communication portion formed at a circumferential
portion except both circumferential edge sides of an axial
direction width on the inner circumference surface of the cam ring
in the section where the vanes extend and connecting the pump
chambers.
[0008] According to a further aspect of the invention, a vane pump
comprises: a rotor rotatably driven and having, at an outer
circumference thereof, a plurality of opening slots; a plurality of
vanes provided in the respective slots; a cam ring housing, at an
inner circumferential side thereof, the rotor and the vanes,
extending/retracting movement of the vanes at the outer
circumference of the rotor occurring by the rotation of the rotor;
a housing housing the cam ring inside the housing and defining a
plurality of pump chambers by the housing, the vanes, the cam ring
and the rotor; an inlet port provided at least one side of both
side walls of the housing which respectively face axial direction
both sides of the cam ring and opening to a section where a volume
of the pump chamber increases; an outlet port provided at least one
side of the both side walls of the housing which respectively face
the axial direction both sides of the cam ring and opening to a
section where the volume of the pump chamber decreases; and a
communication portion formed at a circumferential portion except
both circumferential edge sides of an axial direction width on an
inner circumference surface of the cam ring in the section where
the volume of the pump chamber decreases and connecting the pump
chambers.
[0009] According to the present invention, even when the pump
rotates at high speed, the stable action of the cam ring can be
achieved all the time.
[0010] The other objects and features of this invention will become
understood from the following description with reference to the
accompanying drawings.
BRIEF DESCRIPTION OF THE DRAWINGS
[0011] FIG. 1 is a perspective exploded view of a vane pump
according to a first embodiment.
[0012] FIG. 2 is a front view of the vane pump with a pump cover
removed.
[0013] FIG. 3 is a sectional view of the vane pump.
[0014] FIG. 4 is a longitudinal cross section taken along a plane
A-A in FIG. 2.
[0015] FIG. 5 is a front view of a pump housing.
[0016] FIG. 6 is a perspective view, viewed from one side, of a cam
ring.
[0017] FIG. 7 is a perspective view, viewed from the other side, of
the cam ring.
[0018] FIG. 8 is a sectional view of the cam ring, showing a main
part of the cam ring.
[0019] FIG. 9 is an enlarged view of a circle B in FIG. 8.
[0020] FIG. 10 is an enlarged view of a circle C in FIG. 8.
[0021] FIG. 11 is an enlarged view of a sectional view of the cam
ring at a discharge side.
[0022] FIG. 12 is an enlarged view of a sectional view of the cam
ring at a discharge side.
[0023] FIG. 13 is a drawing for explaining a cam ring action.
[0024] FIG. 14 is a drawing for explaining a cam ring action.
[0025] FIG. 15 is spring displacement-spring load characteristics
of first and second coil springs.
[0026] FIG. 16 is a characteristic showing a relationship between
an engine rpm and a discharge pressure in a related art vane
pump.
[0027] FIG. 17 is a characteristic showing a relationship between a
rotor rotation angle and a pump chamber internal pressure of the
present vane pump and the related art vane pump.
[0028] FIG. 18 is a characteristic showing a relationship between a
pump rotation speed and a discharge quantity of the present vane
pump and the related art vane pump.
[0029] FIG. 19 is a characteristic of a pump discharge pressure and
a discharge quantity of the present vane pump and the related art
vane pump.
[0030] FIG. 20 is a characteristic showing a relationship between a
pump rotation speed and a discharge pressure of the present vane
pump and the related art vane pump.
[0031] FIG. 21 is a sectional view of a cam ring of an other
embodiment, showing a main part of a communication groove of the
cam ring.
[0032] FIG. 22 is a front view of a vane pump of a second
embodiment with a pump cover removed.
DETAILED DESCRIPTION OF THE INVENTION
[0033] Embodiments of a vane pump of the present invention will be
explained below with reference to the drawings. The embodiments
show vane pumps applied to a variable displacement oil pump that
supplies lubricating oil to each sliding part in an internal
combustion engine of a vehicle.
First Embodiment
[0034] As shown in FIGS. 1 to 4, a vane pump is installed at a
front end portion of a cylinder block of the engine. The vane pump
has a pump housing 1 having a bottomed cylindrical shape, one end
opening of which is covered with a cover 2, a driving shaft 3 which
penetrates the pump housing 1 at a center of the pump housing 1 and
is driven and rotated by an engine crankshaft, a rotor 4 which is
rotatably housed inside the pump housing 1 and fixed to the driving
shaft 3 at its center, a cam ring 5 which is a movable member and
is rockably provided at an outer circumferential side of the rotor
4, and a pair of vane rings 6, 6 having a small diameter which are
slidably arranged on both side surfaces at an inner circumferential
side of the rotor 4.
[0035] The pump housing 1 is made of aluminum alloy material and is
integrally formed. As shown in FIG. 5, since one side surface, in
an axial direction, of the cam ring 5 makes sliding contact with a
depressed or hollow bottom surface 1a of the pump housing 1, a
sliding contact area of the bottom surface 1a is formed by
machining with high accuracy of flatness and surface roughness.
[0036] The pump housing 1 is provided with, at a certain position
on an inner circumference surface thereof, a hole into which one
end portion of a pivot pin 9 is inserted and a pivot groove 1c
having a semicircular shape in cross section. The pivot pin 9
serves as a pivot of the cam ring 5 for the rocking motion of the
cam ring 5.
[0037] In addition, as can be seen in FIGS. 2 and 5, a concave
arc-shaped seal surface is provided at an upper side with respect
to a line X (hereinafter called a cam ring reference line)
connecting an axial center of the pivot pin 9 and a center of the
pump housing 1 (a shaft center of the driving shaft 3) at the inner
circumference of the pump housing 1. More specifically, the seal
surface is positioned at an upper left side of the inner
circumference of the pump housing 1, as shown in FIGS. 1 to 3 and
5.
[0038] The seal surface 1s seals one end of an upper end side of an
after-mentioned control oil chamber 16 in cooperation with an
after-mentioned sealing member 14 that is provided in the cam ring
5 with the sealing member 14 making sliding contact with the seal
surface is. This seal surface is, as shown in FIG. 5, formed into
an arc-shaped surface having a certain radius of R1.
[0039] Furthermore, as can be seen in FIG. 5, an inlet port 7 is
provided at a left hand side of the driving shaft 3 on the bottom
surface is of the pump housing 1, also an outlet port 8 is provided
at a right hand side of the driving shaft 3 on the bottom surface
1a. These inlet and outlet ports 7 and 8 are arranged on
substantially opposite sides of the driving shaft 3.
[0040] As shown in FIGS. 4 and 5, the inlet port 7 communicates
with an inlet opening 7a where the oil in a oil pan (not shown) is
pumped up and flows in. On the other hand, the outlet port 8
communicates with an outlet opening 8a, then communicates with the
each sliding part in the engine and e.g. a variable valve timing
control apparatus that variably controls open/close timing of
valves of the engine through an oil main gallery (not shown).
[0041] The inlet port 7 has an arc-shaped inner-side port section
7b and a substantially rectangular outer-side port section 7c. The
outlet port 8 has an arc-shaped inner-side port section 8b and an
outer-side port section 8c that directly communicates with the
outlet opening 8a.
[0042] The bottom surface 1a of the pump housing 1 is provided
with, at the substantially center thereof, a shaft bearing bore if
for supporting the driving shaft 3. This shaft bearing bore if is
supplied with the oil that is discharged from the outlet port 8 via
a depressed groove tip 10a of a substantially L-shaped narrow oil
supply groove 10. Further, the oil supply groove 10 is configured
so that the oil is supplied to both side surfaces of the rotor 4
and a side surface of each vane 11 (described later) from an
opening of the oil supply groove 10 for securing lubrication of
these sliding parts.
[0043] The cover 2 is formed into a thick plate shape, as shown in
FIGS. 1 and 4. The cover 2 is provided with, on a substantially
flat inner side surface 2a thereof, an inlet port 7' and an outlet
port 8' respectively communicating with the inlet port 7 and the
outlet port 8, same as the bottom surface 1a of the pump housing 1.
Further, the cover 2 is provided with, at an edge of the inner side
surface 2a, a pin hole 2b into which the other end portion of the
pivot pin 9 is inserted. In addition, a shaft insertion hole 2c
into which the driving shaft 3 is inserted is formed at a
substantially center of the cover 2 for rotatably supporting the
driving shaft 3.
[0044] As shown in FIG. 1, the cover 2 is fixed to the pump housing
1 with a plurality of bolts B with its positioning in a
circumferential direction made by a plurality of positioning pins
IP.
[0045] The driving shaft 3 rotates the rotor 4 in a clockwise
direction in FIG. 2 by a turning force transmitted from the engine
crankshaft. In FIG. 2 and other drawings, a left half of the
driving shaft 3 is a suction section, while a right half of the
driving shaft 3 is a discharge section.
[0046] The rotor 4 has seven slits (slots) 4a formed in a radially
outward direction from the center side of the rotor 4, as shown in
FIGS. 1 to 3. The vanes 11 are each provided in the slits 4a, and
are supported in the slits 4a so that the vane 11 can move or
extend/retract in the radial direction. Further, a back-pressure
chamber 12 is formed at each base end portion of the slit 4a. This
back-pressure chamber 12 has a substantially circular shape in
cross section and hydraulic pressure (pressurized fluid or oil)
discharged to the outlet port 8 is introduced into the
back-pressure chamber 12 to force the vane 11 outwards in the
radial direction by the pressurized fluid.
[0047] Furthermore, the rotor 4 is provided with annular depressed
grooves or hollows 4b, 4b at the both side surfaces, in the axial
direction, of the inner circumferential side of the rotor 4, The
depressed hollows 4b, 4b support the vane rings 6, 6 so that the
vane rings 6, 6 eccentrically rotate at the inner circumferential
side of the rotor 4.
[0048] As shown in FIG. 2, each inner side base end edge of the
vane 11 makes sliding contact with an outer circumference surfaces
of the pair of vane rings 6, 6, while each top end edge of the vane
11 makes sliding contact with an inner circumference surface 5a of
the cam ring 5.
[0049] A plurality of sector-shaped pump chambers 13, each of which
is a working chamber, are defined by the adjacent vanes 11, the
inner circumference surface 5a of the cam ring 5, an outer
circumference surface of the rotor 4, the bottom surface 1a of the
pump housing 1 and the inner side surface 2a of the cover 2, and
liquid-tightness of each pump chamber 13 is ensured by these parts
and members.
[0050] Each vane ring 6 is set so as to push out each vane 11 in
the radially outward direction.
[0051] The cam ring 5 is made of sintered metal which is readily
formable and is integrally formed into a substantially cylindrical
shape. In FIG. 2, a pivot protrusion 5b is provided at an outer
circumference surface of the cam ring 5 on an upper right side of
the cam ring reference line X. Further, a supporting groove 5k
having a semicircular shape in cross section is formed along an
axial direction of the cam ring 5 at a center on an outer side
surface of the pivot protrusion 5b. The supporting groove 5k
supports the pivot pin 9 in cooperation with the pivot groove 1c,
namely that the pivot pin 9 is inserted and fitted into a
supporting hole formed by the supporting groove 5k and the pivot
groove 1c. The supporting groove 5k (or the pivot pin 9) serves as
an eccentric-rocking or moving fulcrum.
[0052] The cam ring 5 is provided with, at an upper position of the
cam ring reference line X, i.e. at an upper left side position in
FIG. 2, a substantially inverted U-shaped boss portion 5c formed
integrally with the cam ring 5. On an outer surface of the boss
portion 5c, a convex arc-shaped surface (an arc surface) 5d making
sliding contact with the concave arc-shaped seal surface 1s is
formed.
[0053] As shown in FIGS. 6 and 7, a holding groove 5e having a
rectangular shape in longitudinal cross section is formed on the
arc surface 5d. As mentioned above, the sealing member 14 sealing
one end side of the control oil chamber 16 is fitted into and fixed
to the holding groove 5e. On the other hand, the other end side of
the control oil chamber 16 is sealed by the supporting groove 5k of
the pivot protrusion 5b of the cam ring 5 and the pivot pin 9.
[0054] Here, with regard to the arc surface 5d, a radius of
curvature of the arc surface 5d is set to the almost same radius of
curvature as the seal surface 1s so as to form a constant minute or
slight gap between the arc surface 5d and the seal surface 1s.
[0055] The sealing member 14 is made of, for example, synthetic
resin having low friction, and formed into a long narrow shape
along the axial direction of the cam ring 5. The sealing member 14
is set so as to be pressed against the seal surface 1s by an
elastic force of an elastic member 15 (made of, i.e. rubber) which
is secured to a bottom side of the holding groove 5e. With this
sealing structure, good liquid-tightness of the control oil chamber
16 can be ensured all the time.
[0056] The cam ring 5 is provided with, on axial direction both end
surfaces thereof on the inlet port 7 side, a pair of inlet side cut
grooves 18a, 18b by which the oil (fluid) flows into each pump
chamber 13 in the suction section, as shown in FIGS. 4, 6 and 7.
Likewise, the cam ring 5 is provided with, on axial direction both
end surfaces thereof on the outlet port 8 side, a pair of outlet
side cut grooves 18c, 18d by which the oil (fluid) in each pump
chamber 13 flows out to the outlet port 8 in the discharge section.
As can be seen in the drawings, these inlet and outlet side cut
grooves 18a.about.18d are arranged along a circumferential
direction of the cam ring 5.
[0057] The control oil chamber 16 is defined between the outer
circumference surface of the cam ring 5, the pivot protrusion 5b
and the sealing member 14, and takes on a substantially arc shape.
The control oil chamber 16 is configured so as to move or rock the
cam ring 5 in a counterclockwise direction in FIG. 2 with the pivot
pin 9 being the fulcrum when the discharged hydraulic pressure
(pressurized fluid or oil) introduced into the control oil chamber
16 from the outlet port 8 acts on a pressure-receiving surface 5f
of the cam ring outer circumference surface then so as to decrease
an eccentric amount of the cam ring 5 with respect to the rotor
4.
[0058] The cam ring 5 is provided with an arm 17 that is an
extending part protruding radially outwards. The arm 17 is formed
integrally with the cam ring 5, and is placed on the opposite side
to the pivot protrusion 5b formed at the outer circumference
surface of the cylindrical body of the cam ring 5. The arm 17 has,
as shown in FIGS. 1 to 3, 6 and 7, a rectangular plate-shaped arm
body 17a radially extending from the cylindrical body of the cam
ring 5 and a protruding portion 17b formed integrally with an upper
surface, at a top end side, of the arm body 17a.
[0059] The arm body 17a is provided with, at a lower surface
thereof which is opposite side to the protruding portion 17b, a
convex portion 17c that is formed integrally with the arm body 17a
and has a rounded surface. The protruding portion 17b protrudes in
a direction substantially perpendicular to the arm body 17a, and is
provided with an upper surface 17d having a rounded surface whose
radius of curvature is small.
[0060] As can be seen in FIGS. 1 to 3 and 5, a lower side first
spring holder 19 and an upper side second spring holder 21 are
concentrically formed in the pump housing 1 at a position opposite
to the pivot groove 1c, i.e. at upper and lower positions of the
arm 17.
[0061] The first spring holder 19 has a substantially rectangular
shape and extends along an axial direction of the pump housing 1.
On the other hand, although the second spring holder 21 is set to
be shorter than that of the first spring holder 19, the second
spring holder 21 has a substantially rectangular shape and extends
along the axial direction of the pump housing 1, same as the first
spring holder 19.
[0062] As shown in FIG. 5, the second spring holder 21 has a pair
of stopper portions 23, 23 having a long rectangular plate shape.
The stopper portions 23, 23 are formed integrally with the second
spring holder 21, and extend inwards from a width direction of a
lower end opening part 21a so as to face each other. Further, the
stopper portions 23, 23 are formed so that the protruding portion
17b of the arm 17 can be inserted into and come out of the second
spring holder 21 through the opening part 21a between both stopper
portions 23, 23. More specifically, both stopper portions 23, 23
are formed so as to limit or restrict a maximum extension
deformation of an after-mentioned second coil spring 22.
[0063] A first coil spring 20 that is a forcing member is set
inside the first spring holder 19. The first coil spring 20 forces
the cam ring 5 in the clockwise direction in FIG. 2 through the arm
17. That is to say, the first coil spring 20 forces the cam ring 5
in a direction in which the eccentric amount of a center of the
inner circumference surface of the cam ring 5 with respect to a
rotation center of the rotor 4 becomes large.
[0064] The first coil spring 20 has a predetermined spring-load W3.
A lower end of the first coil spring 20 is elastically connected to
a bottom surface 19a of the first spring holder 19, while an upper
end of the first coil spring 20 is in contact with the rounded
surface convex portion 17c formed on the lower surface of the arm
body 17a all the time. The cam ring 5 is then forced in the
direction in which the eccentric amount of the center of the inner
circumference surface of the cam ring 5 with respect to the
rotation center of the rotor 4 becomes large, i.e. in the clockwise
direction in FIG. 2, by the first coil spring 20.
[0065] On the other hand, the second coil spring 22 that is a
forcing member is set inside the second spring holder 21. The
second coil spring 22 forces the cam ring 5 in the counterclockwise
direction in FIG. 2 through the arm 17.
[0066] An upper end of the second coil spring 22 is elastically
connected to an upper wall surface 21b of the second spring holder
21. A lower end of the second coil spring 22 is elastically
connected to the protruding portion 17b of the arm 17 from a
maximum eccentric moving position, in the clockwise direction, of
the arm 17 of the cam ring 5 until the lower end of the second coil
spring 22 is stopped by both the stopper portions 23, 23, then
provides an urging force in the counterclockwise direction to the
cam ring 5.
[0067] Although the second coil spring 22 has a predetermined
spring-load against the first coil spring 20, this spring-load is
set to be smaller than the spring-load W3 of the first coil spring
20. That is, by a load difference W1 in the spring-load between the
first coil spring 20 and the second coil spring 22, the cam ring 5
is set to an initial position (the maximum eccentric position).
[0068] More specifically, the cam ring 5 is forced all the time in
a direction in which the cam ring 5 eccentrically moves upwards,
i.e. in a direction in which a volume of the pump chamber 13
becomes large, by the first coil spring 20 and the second coil
spring 22 with the spring-load W1 provided to the arm 17.
[0069] The spring-load W1 is a load by which the cam ring 5 starts
to move when the hydraulic pressure becomes a required hydraulic
pressure P1 for the variable valve timing control apparatus or
more.
[0070] As described above, the second coil spring 22 is in contact
with the arm 17 when the eccentric amount of the center of the
inner circumference surface of the cam ring 5 with respect to the
rotation center of the rotor 4 is a predetermined value or more.
However, as shown in FIG. 13, when the eccentric amount of the cam
ring 5 with respect to the rotor 4 is less than the predetermined
value, the lower end of the second coil spring 22 is stopped by
both the stopper portions 23, 23 with a compression state of the
second coil spring 22 being held, then the second coil spring 22 is
in almost no contact with the arm 17.
[0071] Here, a spring-load W2 of the first coil spring 20 when the
second coil spring 22 is stopped by both the stopper portions 23,
23 and the load acting on the arm 17 becomes zero is a load by
which the cam ring 5 starts to move when hydraulic pressure becomes
a required hydraulic pressure P2 for a piston oil jet etc. or a
required hydraulic pressure P3 at a maximum rotation of the engine
crankshaft.
[0072] As shown in FIGS. 2 and 8, a root portion upper surface 17f
is formed between the arm body 17a and the cylindrical body of the
cam ring 5. When the cam ring 5 rotates in the clockwise direction
by the spring force of the first coil spring 20, the root portion
upper surface 17f makes contact with a lower surface of the one
stopper portion 23, the further rotation, in the clockwise
direction, of the cam ring 5 is then limited. That is, the moving
or rocking position of the cam ring 5 is limited or set to the
initial set position (the maximum eccentric position) by the spring
force of the first coil spring 20.
[0073] As shown in FIGS. 3, 6 and 7, the cam ring 5 is provided
with, on the inner circumference surface 5a thereof, an inlet side
communication groove 24 and an outlet side communication groove 25
respectively in the suction section (in which each vane 11 extends)
where the arc-shaped inner-side port section 7b of the inlet port 7
is formed and in the discharge section (in which each vane 11
retracts) where the arc-shaped inner-side port section 8b of the
outlet port 8 is formed. These inlet and outlet side communication
grooves 24 and 25 are communication portions.
[0074] Both the inlet and outlet side communication grooves 24 and
25 are formed parallel to the inlet side cut grooves 18a, 18b and
the outlet side cut grooves 18c, 18d respectively. More
specifically, each of the inlet and outlet side communication
grooves 24 and 25 is formed into a band shape, and extends along
the circumferential direction of the inner circumference surface 5a
at a circumferential middle portion except both circumferential
edge side portions 5a', 5a', in the axial direction, of the inner
circumference surface 5a. Lengths L and L1 (see FIG. 3) of the
inlet and outlet side communication grooves 24 and 25 are
respectively set to almost the same arc lengths as the inner-side
port sections 7b and 8b. Axial direction widths of the both
circumferential edge side portions 5a', 5a' of the inner
circumference surface 5a are set to be almost the same as each
other. With this structure, the adjacent pump chambers 13
communicate with each other.
[0075] Further, as can be seen in FIGS. 3 and 8 to 12, each depth d
of the inlet and outlet side communication grooves 24 and 25 is set
to approximately 1/3 (one third) of a thickness of the cam ring
5.
[0076] In addition, starting-point portions 24a, 25a and endpoint
portions 24b, 25b of the inlet and outlet side communication
grooves 24 and 25 are set so that their depths are gradually
shallower from the middle of the inlet and outlet side
communication grooves 24 and 25. That is, as seen in FIGS. 9, 10
and 11, 12 which are enlarged views of circles B and C in FIG. 8
and enlarged views of the discharge side, each of the
starting-point portions and the endpoint portions 24a, 24b and 25a,
25b is formed into an arc-shaped surface having a relatively small
radius R, then gradually becomes shallower from the middle of the
inlet and outlet side communication grooves 24 and 25.
[0077] In the following description, basic operation or working of
the present embodiment will be explained. Before the explanation, a
relationship between a control hydraulic pressure by the related
art variable displacement vane pump employing inner/outer double
coil springs and a required hydraulic pressure for the sliding
parts in the engine or the variable valve timing control apparatus
or a piston cooling device will be explained with reference to FIG.
16.
[0078] With respect to the hydraulic pressure required for the
internal combustion engine, in the case where the variable valve
timing control apparatus is employed for the improved fuel economy
and the exhaust emission control, as an actuating source of this
apparatus, the hydraulic pressure of the oil pump is used. Thus, in
order to improve actuation (operation) response of the apparatus,
as shown by a broken line b in FIG. 16, as an actuating hydraulic
pressure, a high hydraulic pressure P1 is required from an early
stage where an engine rpm is low.
[0079] Further, in the case where the oil jet device for the piston
cooling is employed, a high hydraulic pressure P2 is required at a
middle rpm of the engine. A required hydraulic pressure at the
maximum rpm of the engine is determined mainly by a hydraulic
pressure P3 required for the lubrication of a bearing portion of
the engine crankshaft. Therefore, the hydraulic pressure required
for the whole internal combustion engine is shown by a
characteristic of the broken line connecting lines b and c.
[0080] Here, regarding the middle rpm range required hydraulic
pressure P2 and the high rpm range required hydraulic pressure P3
for the internal combustion engine, its relationship is generally
P2<. P3. Also in many cases, both P2 and P3 are close values.
Thus, in a range D from the middle rpm to the high rpm in FIG. 16,
it is desirable to set the hydraulic pressure so that the hydraulic
pressure does not increase even though the rpm increases.
[0081] In the present embodiment, as shown by a solid line in FIG.
16, in a range from an engine start to the low rpm including engine
idling, since a pump discharge pressure does not reach P1, the cam
ring 5 is in a non-operative state (operation halt state) with the
arm body 17a of the arm 17 of the cam ring 5 making contact with
the one stopper portion 23 of the pump housing 1 by the difference
in the spring-load between the first coil spring 20 and the second
coil spring 22 (see FIG. 2).
[0082] At this time, the eccentric amount of the cam ring 5 is the
maximum and a pump capacity becomes the maximum, then the discharge
pressure quickly or rapidly rises with increase in the engine rpm
as shown by a characteristic of A on the solid line in FIG. 16.
[0083] Subsequently, when the pump discharge pressure further
increases with the further increase in the engine rpm and reaches
Pf, the introduction hydraulic pressure in the control oil chamber
16 increases. The cam ring 5 then starts pressing down or
compressing the first coil spring 20 that acts on the arm 17, and
eccentrically moves or rocks in the counterclockwise direction with
the pivot pin 9 being the fulcrum. Pf is a first cam actuation or
operation pressure, and this Pf is set to be greater than the
required hydraulic pressure P1 for the variable valve timing
control apparatus.
[0084] With this cam ring action, the pump capacity decreases, and
as shown by a characteristic of a range B, increase of the
discharge pressure becomes small or gentle. Then as shown in FIG.
13, the second coil spring 22 is stopped by both the stopper
portions 23, 23 with the compression state of the second coil
spring 22 being held, and the cam ring 5 moves in the
counterclockwise direction with the upper surface 17d of the
protruding portion 17b provided with no load of the second coil
spring 22.
[0085] In this state shown in FIG. 13, the load of the second coil
spring 22 does not act on the cam ring 5 from this point, and the
cam ring 5 is brought in a holding state in which the cam ring 5
does not move until the discharge pressure reaches P2 (hydraulic
pressure P2 in the control oil chamber 16) and exceeds the
spring-load W2 of the first coil spring 20. As a consequence, the
discharge pressure increases with the increase in the engine rpm as
shown by a characteristic of C. However, since the eccentric amount
of the cam ring 5 becomes small and the pump capacity decreases,
this increase of the range C does not show the rapidly rising
characteristic like the range A.
[0086] Further, when the engine rpm increases and the discharge
pressure becomes Ps (P2) or higher, as shown in FIG. 14, the cam
ring 5 moves in the counterclockwise direction while compressing
the first coil spring 20 against the spring force of the
spring-load W2 of the first coil spring 20 through the arm 17. With
this moving action of the cam ring 5, the pump capacity further
decreases and the increase in the discharge pressure becomes small
or gentle, and the engine rpm reaches the maximum rpm while keeping
a discharge pressure characteristic shown by D.
[0087] Consequently, since it is possible to adequately bring the
discharge pressure (the solid line) at the pump high rotation speed
to the required hydraulic pressure (the broken line), loss of power
can be effectively suppressed.
[0088] FIG. 15 shows a relationship between spring displacement of
each of the first and second coil springs 20, 22 or a rocking (or
moving) angle of the cam ring 5 and the spring-loads W1, W2. In the
early stage from the engine start to the low engine rpm, since the
spring-load W1 obtained by subtracting the set load of the second
coil spring 22 from the set load W3 of the first coil spring 20 is
provided to the cam ring 5, the cam ring 5 can not move until the
discharge pressure exceeds the spring-load W1.
[0089] When the discharge pressure exceeds the spring-load W1, the
first coil spring 20 is compressed then its load increases. The
second coil spring 22, meanwhile, gets closer to free length and
its load decreases. As a result, the spring-load increases. This
inclination is a spring constant.
[0090] At a position, shown in FIG. 13, of the cam ring 5, the
spring-load is the spring-load W2 of the first coil spring 20, and
it discontinuously changes (increases). However, when the discharge
pressure exceeds the spring-load W2, although the first coil spring
20 is compressed then its load increases, since the number of the
coil springs acting on the cam ring 5 is changed from two (the both
coil springs 20, 22) to one (the first coil spring 20), the
inclination is changed, namely that the spring constant
decreases.
[0091] As explained above, although when the engine rpm increases
and the discharge pressure reaches P1, the cam ring 5 starts moving
then suppresses the increase in the discharge pressure, when a
moving amount of the cam ring 5 reaches a predetermined moving
amount in the counterclockwise direction shown in FIG. 13, no
spring force of the second coil spring 22 is generated and the
spring constant becomes small, while the spring-load W2 of the
first coil spring 20 discontinuously increases. Therefore, after
the discharge pressure increases to P2, the moving action of the
cam ring 5 starts again (see FIG. 14). That is, since the relative
spring-load of the first and second coil springs 20, 22 acts on the
cam ring 5 and the spring load shows a nonlinear spring
characteristic, the cam ring 5 has specific or unique moving
(rocking) characteristic.
[0092] In this manner, by the nonlinear spring characteristic of
the spring forces of the both coil springs 20, 22, the discharge
pressure has the characteristic shown by the ranges A.about.D in
FIG. 16, and it is possible to adequately bring the control
hydraulic pressure (the solid line) to the required hydraulic
pressure (the broken line). As a result, loss of power caused by
unnecessary increase of the hydraulic pressure can be sufficiently
reduced.
[0093] Further, since the two springs of the first and second coil
springs 20, 22 facing each other are employed, each spring-load of
the coil springs 20, 22 can be arbitrarily set according to the
change of the discharge pressure. It is therefore possible to set
an optimum spring force for the discharge pressure.
[0094] The arm 17 does not make contact with the upper end of the
first coil spring 20 and the lower end of the second coil spring 22
through plunger etc., but directly makes contact with these upper
end and lower end and compresses them. Thus this structure is
simple and increase in parts count can be suppressed. This
facilitates assembly and leads to cost reduction.
[0095] In addition, since the convex portion 17c of the arm body
17a of the arm 17 and the upper surface 17d of the protruding
portion 17b are formed into the arc-shaped rounded surface, change
of their contact angles or contact points with the upper and lower
ends of the first and second coil springs 20, 22 can be small. With
this, displacement of the first coil spring 20 and the lower end of
the second coil spring 22 can be stable.
[0096] Moreover, in the present embodiment, the lubricating oil
discharged from the outlet opening 8a through the outlet port 8 is
used not only for the lubrication of the sliding parts in the
engine, but used as the actuating source of the variable valve
timing control apparatus. As described above and shown in FIG. 16,
since the good rising (range A) of the discharge pressure can be
achieved at the early stage, operation response of a relative
rotational phase control between a timing sprocket and a camshaft,
by which the rotational phase is advanced or retarded, can be
improved even just after the engine start.
[0097] In the present embodiment, by providing the inlet side
communication groove 24 and the outlet side communication groove 25
on the inner circumference surface 5a of the cam ring 5, the
following operation and effect are obtained.
[0098] FIG. 17 is a characteristic showing a relationship between a
rotor rotation angle and a pump chamber internal pressure. A
pressure sensor is provided between the two slits 4a formed on the
outer circumference surface of the rotor 4, and change of the
internal pressure of each pump chamber 13 during the rotation of
the rotor 4 is measured by this pressure sensor. In the drawing, a
vertical axis is the internal pressure of the pump chamber 13
(gauge pressure), and a horizontal axis is the rotation angle (deg)
of the rotor 4. When the rotation angle of the rotor 4 is 0.degree.
(=360.degree.), a volume of the pump chamber 13 becomes a minimum.
When the rotation angle of the rotor 4 is 180.degree., the volume
of the pump chamber 13 becomes a maximum. A range of
0.degree..about.180.degree. is a suction stroke (the suction
section), while a range of 180.degree..about.360.degree. is a
discharge stroke (the discharge section).
[0099] In FIG. 17, a solid line shows an internal pressure change
of a case where no communication grooves 24, 25 are provided. A
broken line shows an internal pressure change of the present
embodiment, i.e. the internal pressure change of the case where the
inlet and outlet side communication grooves 24 and 25 are
provided.
[0100] Regarding the vane pump having no communication grooves, as
can be seen by the solid line, at the beginning of the suction
stroke around rotation angle 0.degree., since shapes of both
openings of the pump chamber 13 which face the inlet ports 7, 7'
are thin long crescent shape, an opening area of each opening is
insufficient for change of volume expansion or volume increase of
the pump chamber 13 at high speed of the pump rotation. That is,
the oil flowing into the pump chamber 13 through the inlet side cut
grooves 18a, 18b can not sufficiently be drawn, and an suction
negative pressure in the pump chamber 13 becomes large (shown by a
in FIG. 17).
[0101] Around rotation angle 90.degree. of the rotor 4, since the
opening area of the opening expands or increases, the oil can be
sufficiently drawn, and the suction negative pressure becomes small
once.
[0102] Around the end of the suction stroke just before the
rotation angle 180.degree., the openings start to be closed in a
state in which the volume of the pump chamber 13 increases and no
inlet ports 7, 7' exist. Thus, the oil can not sufficiently be
drawn, and the suction negative pressure becomes large again (shown
by b in FIG. 17). Because the insufficient oil suction can not be
resolved, the so-called cavitation occurs, and this causes
insufficient oil discharge quantity or an occurrence of pump
noise/vibration.
[0103] FIG. 18 shows a relationship between a pump rotation speed
(N) and a discharge quantity (Q) of the vane pump. As shown by a
solid line (the related art vane pump), the discharge quantity does
not increase in a high speed range of the pump. This is caused by
the occurrence of the cavitation.
[0104] Returning to FIG. 17, when the rotation angle of the rotor 4
exceeds 180.degree., both sides of the pump chamber 13 in which the
negative pressure has been increased open to the outlet port 8, 8'
(the inner-side port section 8b). Since both sides of the pump
chamber 13 open to the outlet port 8, 8' where the pressure is
high, the oil flows into the outlet port 8, 8' rapidly at once. The
oil then crashes into each other at almost circumferential middle
portion of the inner circumference surface 5a, and a spike pressure
(shown by c in FIG. 17) occurs.
[0105] Between the rotation angle 270.degree..about.360.degree.,
since shapes of the openings of both sides of the pump chamber 13
which face the outlet port 8, 8' are thin long crescent shape and
become thinner, the internal pressure of the pump chamber 13
increases to discharge the oil to the outlet port 8, 8'.
[0106] Around the end of the discharge stroke around rotation angle
360.degree., since no outlet port 8, 8' exist and the openings of
the pump chamber 13 are closed, a great closure pressure (shown by
d in FIG. 17) occurs.
[0107] Increase of the internal pressure of the pump chamber 13
accompanied by the spike pressure (c) and the closure pressure (d)
causes increase of friction of each part in pump and the occurrence
of pump noise/vibration.
[0108] In particular, increase of the closure pressure (d) results
in problem that lowers the actuation or operation pressure (moving
pressure) of the cam ring 5. That is, although the actuation
pressure of the cam ring 5 is essentially determined by the spring
forces of the first and second coil springs 20, 22 and the pump
discharge pressure in the control oil chamber 16 which acts on the
outer circumference surface of the cam ring 5, the closure pressure
(d) acts as a moving (rocking) torque in a direction in which the
eccentric amount of the cam ring 5 becomes small from the inner
circumference surface 5a of the cam ring 5, then lowers the
actuation pressure of the cam ring 5.
[0109] FIG. 19 shows a characteristic of a pump discharge pressure
(P) and a discharge quantity (Q) according to the pump rotation
speed. As is clear from this characteristic, in the case of the
related art pump (a solid line), the pump discharge pressure
sharply decreases with the increase in the pump rotation speed.
This is caused by the lowering of the actuation pressure of the cam
ring 5.
[0110] FIG. 20 shows a relationship between a pump rotation speed
and a pump discharge pressure. In the case of the related art pump
(a solid line), although the pump discharge pressure increases with
increase in the pump rotation speed, the pump discharge pressure
does not increase any more from about a pump discharge pressure P2.
This is also caused by the lowering of the actuation pressure of
the cam ring 5.
[0111] Since the cause of the insufficient oil suction in the
suction stroke is the insufficient opening area of the openings,
facing the inlet ports 7, 7', of the pump chamber 13, the oil is
the most insufficient at the middle in the axial direction in the
pump chamber 13, i.e. at almost circumferential middle portion of
the inner circumference surface 5a, and the suction negative
pressure becomes large. Although decrease of the negative pressure
becomes large in the pump chamber 13 at the suction start point
(around the rotor rotation angle 0.degree.) and at the suction end
point (around the rotor rotation angle 180.degree.), the decrease
of the negative pressure is cancelled around the rotor rotation
angle 90.degree..
[0112] In the present embodiment, the adjacent pump chambers 13,
each of which opens to the inlet ports 7, 7', communicate with each
other at almost circumferential middle portion of the inner
circumference surface 5a through the inlet side communication
groove 24. Hence, oil for canceling or resolving the insufficient
oil suction can be supplied through this inlet side communication
groove 24. As a consequence, the suction negative pressure can be
shared or equalized, and the insufficient oil suction can be
resolved.
[0113] This is shown by a broken line in FIG. 17. Around the rotor
rotation angle 0.degree. and the rotor rotation angle 180.degree.
in the suction stroke, internal pressures a' and b' (the broken
line) of the pump chamber 13 are negative pressures and are close
to 0, as compared with the suction negative pressure a and b (the
solid line). Further, between the rotation angle
0.degree..about.180.degree., the internal pressure are equalized,
as compared with that of the related art pump. This results from
the setting of the inlet side communication groove 24.
[0114] In particular, since the inlet side communication groove 24
is set at the circumferential middle portion, in the axial
direction of the cam ring 5, on the inner circumference surface 5a,
it is possible to effectively suppress the occurrence of the
cavitation at the middle portion where the negative pressure is apt
to occur.
[0115] Further, as shown by broken lines in FIG. 18, the pump
discharge quantity (Q) shows a proportional increase to the pump
rotation speed (N) even in the high speed range of the pump without
decreasing. Also consumption horsepower reduces. These result from
the effect of suppressing the cavitation by the inlet side
communication groove 24.
[0116] On the other hand, with respect to the discharge stroke, in
the case of the related art having no outlet side communication
groove 25, it is clear, as shown by the solid line, that the cause
of increase of the internal pressure of the pump chamber 13 is the
insufficient opening area of the openings, facing the outlet port
8, 8', of the pump chamber 13, and the internal pressure most
increases at the middle portion, in the axial direction, in the
pump chamber 13.
[0117] In contrast, in the present embodiment, since the outlet
side communication groove 25 is provided and then the adjacent pump
chambers 13 communicate with each other, as shown by the broken
line, it is possible to equalize the internal pressure of the pump
chamber 13. With this, as shown by broken lines in FIG. 19, an
effect of improving the lowering of the actuation pressure of the
cam ring 5 can be obtained.
[0118] In addition, as shown by the broken lines in FIG. 18, it is
clear that consumption horsepower in the high speed range reduces.
This results from the effect of equalizing the internal pressure of
the pump chamber 13 by the outlet side communication groove 25.
[0119] Moreover, as shown by a broken line in FIG. 20, also an
effect of suppressing decrease of the pump discharge pressure in
the pump high speed range results from the working of the outlet
side communication groove 25 together with the working of the inlet
side communication groove 24.
[0120] Next, in the embodiment, the starting-point portions 24a,
25a and the endpoint portions 24b, 25b of the inlet and outlet side
communication grooves 24 and 25 are set so that these portions 24a,
25a and 24b, 25b have the arc-shaped surface and their depths are
gradually shallower from the middle of the inlet and outlet side
communication grooves 24 and 25. Operation and Effect by these
structures will be explained.
[0121] As shown in FIG. 9, when the pump chamber 13 changes from
the discharge section (the discharge stroke) to the suction section
(the suction stroke) through a minimum volume section, since the
depth of the groove of the starting-point portion 24a of the inlet
side communication groove 24 is gradually deeper and a
cross-sectional area of the inlet side communication groove 24 is
gradually larger, a rapid release of the closure pressure (shown by
d in FIG. 17) can be suppressed. This can adequately suppress a
rapid change of the pressure and an occurrence of the
noise/vibration caused by this rapid pressure change.
[0122] As for the endpoint portion 24b side of the inlet side
communication groove 24, as shown in FIG. 10, since the depth of
the groove of the endpoint portion 24b is gradually shallower from
the middle of the inlet side communication groove 24, the oil
dragged in the inlet side communication groove 24 by the rotation
of the vane 11 flows into the pump chamber 13 that is positioned at
a rotation direction side by an inclined surface (the arc-shaped
surface) of the endpoint portion 24b. That is, a pressure-charging
(or supercharging) effect of the oil can be obtained. As a
consequence, the effect of reducing or suppressing the cavitation
can be further obtained.
[0123] On the other hand, with regard to the starting-point portion
25a side of the outlet side communication groove 25, as shown in
FIG. 11, when the pump chamber 13, which is in the negative
pressure state, changes from the suction stroke to the discharge
stroke, since the depth of the groove of the starting-point portion
25a is gradually deeper with the rotation of the rotor 4 (the vane
11), the high pressure from the adjacent pump chambers 13
positioned at the outlet port 8, 8' side in an opposite direction
to the rotation direction is not rapidly released. With this, the
occurrence of the great spike pressure (shown by c in FIG. 17) is
suppressed, thereby reducing the noise/vibration.
[0124] Further, the inlet and outlet side communication grooves 24
and 25 are formed at the circumferential middle portion, in the
axial direction, of the inner circumference surface 5a, and the
both circumferential edge side portions (circumferential edge side
surfaces) 5a', 5a' exist on both sides of each of the communication
grooves 24, 25. With this structure, the vanes 11 rotate by the
rotation of the rotor 4 with each top end edge of the vane 11
guided and supported by and making sliding contact with the both
circumferential edge side portions 5a', 5a'. Accordingly, each vane
11 is supported with stability even at the communication grooves
24, 25 sides without inclining or leaning, and a problem that the
vanes 11 strike against the bottom surface 1a of the pump housing 1
at edge portions of the inlet ports 7, 7' and the outlet port 8, 8'
then are broken can be prevented.
[0125] Furthermore, the inlet side cut grooves 18a, 18b are formed
on the axial direction both end surfaces of the cam ring 5.
Therefore, good oil flow (inflow) from the inlet ports 7, 7' into
the pump chamber 13 can be ensured. Likewise, the outlet side cut
grooves 18c, 18d are also formed on the axial direction both end
surfaces of the cam ring 5. Good oil flow (outflow) from each pump
chamber 13 to the outlet port 8, 8' can be therefore ensured.
[0126] In the present embodiment, although the inlet and outlet
side communication grooves 24 and 25 are formed at the
substantially middle portion of the axial direction width on the
inner circumference surface 5a, their positions could shift. That
is, since there is a case where an area where the negative pressure
is greatest and an area where the increase in the internal pressure
is highest in the pump chamber 13 shift from the middle portion of
the axial direction width to both sides depending on the depths of
the inlet ports 7, 7' and the outlet port 8, 8' provided on the
pump housing 1 side and the cover 2 side, in this case, the
positions where the communication grooves 24, 25 are formed could
shift to one side of the circumferential edge side portions 5a',
5a' on the inner circumference surface 5a. However, also in this
case, the communication grooves 24, 25 are formed in the position
so as to always include the middle of the axial direction width on
the inner circumference surface 5a.
[0127] Further, as shown in FIG. 21, the communication grooves 24,
25 and the starting-point portions 24a, 25a and the endpoint
portions 24b, 25b could be formed so that the communication grooves
24, 25 are formed into an arc shape having a large radius R of
curvature by cutting and depths of the rounded surfaces of the
starting-point portions 24a, 25a and the endpoint portions 24b, 25b
gradually change from the middle portion.
Second Embodiment
[0128] FIG. 2 shows a second embodiment. A basic structure such as
the main components is the same as the first embodiment. However,
in this embodiment, two control oil chambers 16a and 16b, which
move the cam ring 5 by hydraulic pressure (pressurized fluid or
oil) so that the eccentric amount of the cam ring 5 becomes large,
are formed at upper and lower sides of the pivot pin 9 with the
pivot pin 9 being a center.
[0129] That is, the control oil chamber 16 of the first embodiment
corresponds to the first control oil chamber 16a of the second
embodiment. The second control oil chamber 16b is formed by a
substantially L-shaped concave trench 30 at the lower side of the
pivot pin 9 inside the pump housing 1.
[0130] A second seal surface 30a is provided at a lower portion of
the concave trench 30. More specifically, the second seal surface
30a is formed into an arc-shaped surface having a certain radius
with the pivot pin 9 being a center.
[0131] On the other hand, a substantially triangular convex portion
31 which faces the concave trench 30 is formed integrally with the
cam ring 5. The convex portion 31 has, in an opposing position to
the second seal surface 30a, an arc-shaped surface 31a having the
certain radius with the pivot pin 9 being the center. Further, a
holding groove having a rectangular shape in longitudinal cross
section is formed on a top end side of the arc-shaped surface 31a.
A sealing member 32 making sliding contact with the second seal
surface 30a and an elastic member 33 having a rectangular shape in
longitudinal cross section for pressing the sealing member 32
against the second seal surface 30a are provided in the holding
groove.
[0132] An arc length of the second seal surface 30a is set so that
the sealing member 32 can make sliding contact with the second seal
surface 30a even when the cam ring 5 moves and its eccentric amount
with respect to the rotor 4 varies from the maximum eccentric
position (FIG. 2) to the minimum eccentric position (FIG. 14).
[0133] The second control oil chamber 16b communicates with the
outlet port 8 via a communication groove 1g that is formed on the
bottom surface 1a of the pump housing 1. Therefore, the same
discharge pressure as the discharge pressure which the
pressure-receiving surface 5f receives in the first control oil
chamber 16a acts on a second pressure-receiving surface 5g formed
on an outer circumference surface of the cam ring 5, which faces
the second control oil chamber 16b.
[0134] The radius of curvature of the second arc-shaped surface 31a
is set to be smaller than the radius of curvature of the first arc
surface 5d on the first sealing member 14 side. Thus, a surface
area of the second pressure-receiving surface 5g is smaller than
that of the first pressure-receiving surface 5f. When the discharge
pressures in the first and second control oil chambers 16a, 16b act
on the pressure-receiving surfaces 5f, 5g respectively, the moving
(rocking) torque in the counterclockwise direction is generated to
the cam ring 5, same as the first embodiment. However, since the
hydraulic pressure torque from the second control oil chamber 16b,
which acts on only the second pressure-receiving surface 5g, is the
torque in the clockwise direction, part of the torque from the
first control oil chamber 16a is cancelled. As a result, in a case
where the discharge pressures in the first and second control oil
chambers 16a, 16b are the same, the moving torque of the cam ring 5
is small as compared with the first embodiment.
[0135] Thus, since the spring forces of the first and second coil
springs 20, 22 can be set to be small, each coil diameter of the
coil springs 20, 22 can be set to be small. This leads to reduction
in overall size of the vane pump.
[0136] However, since the moving torque generated to the
pressure-receiving surfaces 5f, 5g of the cam ring 5 becomes small,
an effect of the increase of the internal pressure of each pump
chamber 13 on the actuation pressure of the cam ring 5 becomes
large. The effect of improving the lowering of the actuation
pressure of the cam ring 5 by the inlet and outlet side
communication grooves 24, 25 therefore works well.
[0137] The present invention is not limited to the above
embodiments. For example, each axial direction width of the inlet
and outlet side communication grooves 24, 25 could be set to be
small. Or each of the inlet and outlet side communication grooves
24, 25 might be formed by a plurality of narrow long bands arranged
parallel to each other.
[0138] Further, instead of the inlet and outlet side communication
grooves 24, 25 of the communication portions, a hole or a conduit
could be provided in the cam ring 5 as a communication channel.
[0139] Furthermore, although the sealing member 14 is provided to
ensure the liquid-tightness of the control oil chamber 16; if a
required pressure characteristic for the internal combustion engine
is satisfied, the sealing member could be eliminated for cost
reduction.
[0140] Arrangement of the first and second spring holders 19, 21
could be changed. The spring loads of the first and second coil
springs 20, 22 can be freely set according to specifications of the
pump or a pump size, also their coil diameters and lengths could be
changed.
[0141] As for the variable valve timing control apparatus, a
variable valve control apparatus is not limited to the variable
valve timing control apparatus. For instance, it could be a
valve-lift control apparatus driven by the fluid pressure and
controlling an operating angle and a lift amount of the engine
valve. Then the present invention can be used as the actuating
source of the valve-lift control apparatus.
[0142] Moreover, the vane pump of the present invention can be used
for fluid pressure-driven devices except the internal combustion
engine.
[0143] From the foregoing, the present invention has the following
effects.
The groove 24; 25 is formed so that a depth of the groove 24; 25 is
gradually shallower from a middle, in the circumferential
direction, of the groove 24; 25 to both ends of the groove 24;
25.
[0144] Since the depth of the both ends, in the circumferential
direction, of the groove 24; 25 is gradually shallower, the oil
coming from the middle of the groove 24; 25 into one of the
adjacent pump chambers 13 can smoothly flow into the other of the
adjacent pump chambers 13.
[0145] The groove 24; 25 is formed by cutting. Since the groove 24;
25 is formed only by cutting, production cost can be reduced.
[0146] The grooves 24; 25 are provided in both of the oil suction
section and the oil discharge section on the inner circumference
surface 5a of the cam ring 5. With this, the occurrence of the
cavitation in the suction section can be suppressed, and also the
pressure change in the discharge section can be suppressed, then
the stable movement of the cam ring 5 can be achieved.
[0147] The vane pump is a vane pump in which the oil is drawn from
both sides, in the axial direction, of the cam ring 5 and is
discharged from both sides, in the axial direction, of the cam ring
5, and the groove 24; 25 is formed at a circumferential portion
except both circumferential edge sides 5a', 5a' of the axial
direction width on the inner circumference surface 5a of the cam
ring 5.
[0148] Widths of the both circumferential edge sides 5a', 5a' of
the axial direction width on the inner circumference surface 5a,
where no groove 24; 25 exists, are set to be substantially the
same.
[0149] The communication portion 24 is formed by a groove 24 that
extends in a circumferential direction of the inner circumference
surface 5a of the cam ring 5.
[0150] The groove 24 is formed so that depths of both ends of the
groove 24 is shallower than that of a middle, in the
circumferential direction, of the groove 24.
[0151] The cam ring 5 moves with respect to the rotor 4 and an
eccentric amount of the cam ring 5 with respect to the rotor 4
changes, and an oil amount discharged from the outlet port 8 is
varied by the change of the eccentric amount.
[0152] The cam ring 5 is forced in a direction in which the
eccentric amount with respect to the rotor 4 becomes large by a
forcing member 20, and
the cam ring 5 is moved in an opposite direction against the force
of the forcing member 20, for controlling the oil discharge
amount.
[0153] The cam ring 5 is configured to move in the opposite
direction by receiving a pressure of the outlet port 8.
[0154] The communication portion 24 is formed by a communication
conduit.
[0155] From the foregoing, the stable action of the cam ring can be
achieved all the time.
[0156] The entire contents of Japanese Patent Application No.
2010-18201 filed on Jan. 29, 2010 are incorporated herein by
reference.
[0157] Although the invention has been described above by reference
to certain embodiments of the invention, the invention is not
limited to the embodiments described above. Modifications and
variations of the embodiments described above will occur to those
skilled in the art in light of the above teachings. The scope of
the invention is defined with reference to the following
claims.
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