U.S. patent application number 13/011986 was filed with the patent office on 2011-08-11 for variable displacement pump, oil jet and lublicating system using variable displacement pump.
This patent application is currently assigned to HITACHI AUTOMOTIVE SYSTEMS, LTD.. Invention is credited to Hideaki Ohnishi, Yasushi WATANABE.
Application Number | 20110194967 13/011986 |
Document ID | / |
Family ID | 44353869 |
Filed Date | 2011-08-11 |
United States Patent
Application |
20110194967 |
Kind Code |
A1 |
WATANABE; Yasushi ; et
al. |
August 11, 2011 |
VARIABLE DISPLACEMENT PUMP, OIL JET AND LUBLICATING SYSTEM USING
VARIABLE DISPLACEMENT PUMP
Abstract
A variable displacement pump arranged to supply a hydraulic
fluid to an oil jet arranged to inject the hydraulic fluid to a
piston of an internal combustion engine, the variable displacement
pump includes: a pump constituting section arranged to be driven
and rotated by the internal combustion engine, and thereby to
discharge the hydraulic fluid; a movable member arranged to
decrease the flow rate discharged from the discharge portion by
moving in one direction; a control section configured to move the
movable member in the one direction by a predetermined amount when
the discharge pressure becomes a first discharge pressure, and to
further move the movable member in the one direction when the
discharge pressure becomes a second discharge pressure larger than
the first discharge pressure, the first discharge pressure being
set smaller than the predetermined pressure at which the oil jet
starts to inject the hydraulic fluid.
Inventors: |
WATANABE; Yasushi;
(Aiko-gun, JP) ; Ohnishi; Hideaki; (Atsugi-shi,
JP) |
Assignee: |
HITACHI AUTOMOTIVE SYSTEMS,
LTD.
|
Family ID: |
44353869 |
Appl. No.: |
13/011986 |
Filed: |
January 24, 2011 |
Current U.S.
Class: |
418/138 |
Current CPC
Class: |
F04C 14/226 20130101;
F04C 2/3441 20130101 |
Class at
Publication: |
418/138 |
International
Class: |
F04C 15/00 20060101
F04C015/00 |
Foreign Application Data
Date |
Code |
Application Number |
Feb 9, 2010 |
JP |
2010-026335 |
Claims
1. A variable displacement pump arranged to supply a hydraulic
fluid to an oil jet arranged to inject the hydraulic fluid to a
piston of an internal combustion engine when a pressure of the
supplied hydraulic fluid becomes equal to or greater than a
predetermined pressure, the variable displacement pump comprising:
a pump constituting section arranged to be driven and rotated by
the internal combustion engine, and thereby to discharge the
hydraulic fluid entered from a suction portion to a plurality of
operation chambers, from a discharge portion by volume variations
of the operation chambers; a movable member arranged to decrease
the flow rate of the hydraulic fluid discharged from the discharge
portion by moving in one direction; and a control section
configured to move the movable member in the one direction by a
predetermined amount when the discharge pressure of the hydraulic
fluid becomes a first discharge pressure, and to further move the
movable member in the one direction when the discharge pressure of
the hydraulic fluid becomes a second discharge pressure larger than
the first discharge pressure, the first discharge pressure being
set smaller than the predetermined pressure at which the oil jet
starts to inject the hydraulic fluid.
2. The variable displacement pump as claimed in claim 1, wherein
the second discharge pressure is set larger than the predetermined
pressure at which the oil jet starts to inject the hydraulic
fluid.
3. The variable displacement pump as claimed in claim 1, wherein
the oil jet includes; a body including a hydraulic fluid supplying
portion to which the hydraulic fluid is supplied, a hydraulic fluid
introducing portion arranged to introduce the hydraulic fluid
supplied to the hydraulic fluid supplying portion, and a valve seat
formed between the hydraulic fluid supplying portion and the
hydraulic fluid introducing portion; a valve element arranged to be
seated on and released from the valve seat in accordance with the
pressure of the hydraulic fluid supplied to the hydraulic fluid
supplying portion, and thereby to open and close the hydraulic
fluid supplying portion; an urging member arranged to urge the
valve element in a valve closing direction in which the valve
element closes the hydraulic fluid supplying portion, and to set a
valve opening pressure of the valve element at which the valve
element opens the hydraulic fluid supplying portion, to a value
larger than the first discharge pressure; and an injection nozzle
connected on a downstream side of the hydraulic pressure
introducing portion, and arranged to inject the hydraulic fluid
from an injection opening toward the piston.
4. The variable displacement pump as claimed in claim 1, wherein
the movable member is a cam ring having a cam surface formed on an
inner circumference surface thereof; the pump forming section
includes a rotor arranged to be driven and rotated by the internal
combustion engine, and vanes disposed on an outer circumference
portion of the rotor, and arranged to be moved in a radially inward
direction or in a radially outward direction, and to be moved in
the radially outward direction toward the inner circumference
surface to separate the plurality of the operation chambers; and
the cam ring is arranged to move to vary an eccentric amount of the
cam ring with respect to a center of the rotor.
5. The variable displacement pump as claimed in claim 4, wherein
the discharged hydraulic fluid lubricates sliding portions of the
internal combustion engine.
6. The variable displacement pump as claimed in claim 1, wherein
the discharged hydraulic fluid activates a valve timing control
apparatus arranged to vary a relative rotational phase between a
driving rotational member and a cam shaft of the internal
combustion engine, and a lock mechanism of the valve timing control
apparatus; and the lock mechanism has a release pressure at which a
lock of the lock mechanism is released, and which is set smaller
than the first discharge pressure.
7. A lubricating system comprising: an oil jet arranged to inject a
hydraulic fluid to a piston of an internal combustion engine when a
pressure of a supplied hydraulic fluid becomes equal to or greater
than a predetermined pressure; and a variable displacement pump
arranged to supply the hydraulic fluid to the oil jet, the variable
displacement pump including; a pump constituting section arranged
to be driven and rotated by the internal combustion engine, and
thereby to discharge the hydraulic fluid entered from a suction
portion to a plurality of operation chambers, from a discharge
portion by volume variations of the operation chambers; a movable
member arranged to decrease the flow rate of the hydraulic fluid
discharged from the discharge portion by moving in one direction;
and a control section configured to move the movable member in the
one direction by a predetermined amount when the discharge pressure
of the hydraulic fluid becomes a first discharge pressure, and to
further move the movable member in the one direction when the
discharge pressure of the hydraulic fluid becomes a second
discharge pressure larger than the first discharge pressure, the
first discharge pressure being set smaller than the predetermined
pressure at which the oil jet starts to inject the hydraulic
fluid.
8. An oil jet comprising: a body including a hydraulic fluid
supplying portion to which a hydraulic fluid is supplied, a
hydraulic fluid introducing portion arranged to introduce the
hydraulic fluid supplied to the hydraulic fluid supplying portion,
and a valve seat formed between the hydraulic fluid supplying
portion and the hydraulic fluid introducing portion; a valve
element arranged to be seated on and released from the valve seat
in accordance with the pressure of the hydraulic fluid supplied to
the hydraulic fluid supplying portion, and thereby to open and
close the hydraulic fluid supplying portion; an urging member
arranged to urge the valve element in a valve closing direction in
which the valve element closes the hydraulic fluid supplying
portion, and to set a valve opening pressure of the valve element
at which the valve element opens the hydraulic fluid supplying
portion, to a value larger than the first discharge pressure; and
an injection nozzle connected on a downstream side of the hydraulic
pressure introducing portion, and arranged to inject the hydraulic
fluid from an injection opening toward the piston, the hydraulic
fluid supplying portion of body receiving the supply of the
hydraulic fluid from a variable displacement pump including a pump
constituting section arranged to be driven and rotated by an
internal combustion engine, and thereby to discharge the hydraulic
fluid entered from a suction portion to a plurality of operation
chambers, from a discharge portion by volume variations of the
operation chambers, a movable member arranged to decrease the flow
rate of the hydraulic fluid discharged from the discharge portion
by moving in one direction, and a control section configured to
move the movable member in the one direction by a predetermined
amount when the discharge pressure of the hydraulic fluid becomes a
first discharge pressure, and to further move the movable member in
the one direction when the discharge pressure of the hydraulic
fluid becomes a second discharge pressure larger than the first
discharge pressure, the valve element of the body being arranged to
open the hydraulic fluid supplying portion to inject the hydraulic
fluid to a piston of the internal combustion engine when a pressure
of the supplied hydraulic fluid becomes equal to or greater than a
predetermined pressure larger than the first discharge
pressure.
9. The oil jet as claimed in claim 8, wherein the oil jet is
arranged to start to inject the hydraulic fluid at the
predetermined pressure smaller than the second discharge pressure.
Description
BACKGROUND OF THE INVENTION
[0001] The present invention relates to a variable displacement
pump used in an internal combustion engine and so on for a vehicle,
an oil jet and a lubricating system using the variable displacement
pump.
[0002] U.S. Patent Application Publication No. 2009-0101092
(corresponding to Japanese Patent Application Publication No.
2009-97424) discloses a conventional variable displacement pump
including a first spring arranged to constantly act an urging force
to a cam ring, and a second spring arranged to provide an urging
force in a direction opposite to the urging force of the first
spring when the cam ring is moved by a predetermined distance or
more. In this conventional variable displacement pump, an eccentric
state of the cam ring is varied in two stages (steps) by the
relative urging forces of the springs, so that the discharge flow
rate characteristic is varied in the two stages.
[0003] Moreover, this variable displacement pump is arranged to
release a lock state of a valve timing control apparatus by the
first discharge pressure before the cam ring is moved against the
urging force of the first spring.
SUMMARY OF THE INVENTION
[0004] However, when this conventional variable displacement pump
is also used for supplying the oil to an oil jet arranged to cool a
piston of the internal combustion engine. When the oil in the first
stage of the discharge pressure of the variable displacement pump,
that is, the oil before the cam ring is moved is supplied, the
unnecessary energy may be consumed until the discharge pressure to
move the cam ring is obtained.
[0005] It is an object of the present invention to provide a
variable displacement pump devised to suppress an energy
consumption in an initial state of a discharge of an oil.
[0006] According to one aspect of the present invention, a variable
displacement pump arranged to supply a hydraulic fluid to an oil
jet arranged to inject the hydraulic fluid to a piston of an
internal combustion engine when a pressure of the supplied
hydraulic fluid becomes equal to or greater than a predetermined
pressure, the variable displacement pump comprises: a pump
constituting section arranged to be driven and rotated by the
internal combustion engine, and thereby to discharge the hydraulic
fluid entered from a suction portion to a plurality of operation
chambers, from a discharge portion by volume variations of the
operation chambers; a movable member arranged to decrease the flow
rate of the hydraulic fluid discharged from the discharge portion
by moving in one direction; and a control section configured to
move the movable member in the one direction by a predetermined
amount when the discharge pressure of the hydraulic fluid becomes a
first discharge pressure, and to further move the movable member in
the one direction when the discharge pressure of the hydraulic
fluid becomes a second discharge pressure larger than the first
discharge pressure, the first discharge pressure being set smaller
than the predetermined pressure at which the oil jet starts to
inject the hydraulic fluid.
[0007] According to another aspect of the invention, a lubricating
system comprises: an oil jet arranged to inject a hydraulic fluid
to a piston of an internal combustion engine when a pressure of a
supplied hydraulic fluid becomes equal to or greater than a
predetermined pressure; and a variable displacement pump arranged
to supply the hydraulic fluid to the oil jet, the variable
displacement pump including; a pump constituting section arranged
to be driven and rotated by the internal combustion engine, and
thereby to discharge the hydraulic fluid entered from a suction
portion to a plurality of operation chambers, from a discharge
portion by volume variations of the operation chambers; a movable
member arranged to decrease the flow rate of the hydraulic fluid
discharged from the discharge portion by moving in one direction;
and a control section configured to move the movable member in the
one direction by a predetermined amount when the discharge pressure
of the hydraulic fluid becomes a first discharge pressure, and to
further move the movable member in the one direction when the
discharge pressure of the hydraulic fluid becomes a second
discharge pressure larger than the first discharge pressure, the
first discharge pressure being set smaller than the predetermined
pressure at which the oil jet starts to inject the hydraulic
fluid.
[0008] According to still another aspect of the invention, an oil
jet comprises: a body including a hydraulic fluid supplying portion
to which a hydraulic fluid is supplied, a hydraulic fluid
introducing portion arranged to introduce the hydraulic fluid
supplied to the hydraulic fluid supplying portion, and a valve seat
formed between the hydraulic fluid supplying portion and the
hydraulic fluid introducing portion; a valve element arranged to be
seated on and released from the valve seat in accordance with the
pressure of the hydraulic fluid supplied to the hydraulic fluid
supplying portion, and thereby to open and close the hydraulic
fluid supplying portion; an urging member arranged to urge the
valve element in a valve closing direction in which the valve
element closes the hydraulic fluid supplying portion, and to set a
valve opening pressure of the valve element at which the valve
element opens the hydraulic fluid supplying portion, to a value
larger than the first discharge pressure; and an injection nozzle
connected on a downstream side of the hydraulic pressure
introducing portion, and arranged to inject the hydraulic fluid
from an injection opening toward the piston, the hydraulic fluid
supplying portion of body receiving the supply of the hydraulic
fluid from a variable displacement pump including a pump
constituting section arranged to be driven and rotated by an
internal combustion engine, and thereby to discharge the hydraulic
fluid entered from a suction portion to a plurality of operation
chambers, from a discharge portion by volume variations of the
operation chambers, a movable member arranged to decrease the flow
rate of the hydraulic fluid discharged from the discharge portion
by moving in one direction, and a control section configured to
move the movable member in the one direction by a predetermined
amount when the discharge pressure of the hydraulic fluid becomes a
first discharge pressure, and to further move the movable member in
the one direction when the discharge pressure of the hydraulic
fluid becomes a second discharge pressure larger than the first
discharge pressure, the valve element of the body being arranged to
open the hydraulic fluid supplying portion to inject the hydraulic
fluid to a piston of the internal combustion engine when a pressure
of the supplied hydraulic fluid becomes equal to or greater than a
predetermined pressure larger than the first discharge
pressure.
BRIEF DESCRIPTION OF THE DRAWINGS
[0009] FIG. 1 is an exploded perspective view showing a variable
displacement pump according to a first embodiment of the present
invention.
[0010] FIG. 2 is a front view showing the variable displacement
pump from which a cover member of the variable displacement pump is
detached.
[0011] FIG. 3 is a sectional view taken along a section line A-A of
FIG. 2.
[0012] FIG. 4 is a front view showing a pump housing of the
variable displacement pump of FIG. 1.
[0013] FIG. 5 is a view for illustrating an operation of the
variable displacement pump of FIG. 1.
[0014] FIG. 6 is a view for illustrating an operation of the
variable displacement pump of FIG. 1.
[0015] FIG. 7 is a characteristic view showing a relationship
between spring displacements and spring set loads of first and
second coil springs.
[0016] FIG. 8 is a characteristic view showing a relationship
between a discharge hydraulic pressure and an engine speed in a
conventional variable displacement pump.
[0017] FIG. 9 is a longitudinal sectional view showing an internal
combustion engine which employs an oil jet according to the first
embodiment of the present invention.
[0018] FIG. 10 is a perspective view showing the oil jet according
to the first embodiment of the present invention.
[0019] FIG. 11A is a longitudinal sectional view showing a valve
closed state of the oil jet. FIG. 11B is a longitudinal sectional
view showing a valve open state of the oil jet.
[0020] FIG. 12 is an overall sectional view showing a valve timing
control apparatus according to the first embodiment of the present
invention.
[0021] FIG. 13 is a sectional view showing the valve timing control
apparatus in which a vane member is rotated to a most retarded
position.
[0022] FIG. 14 is a sectional view showing the valve timing control
apparatus in which the vane member is rotated in a most advanced
side.
[0023] FIG. 15 is a sectional view showing a lock mechanism of the
valve timing control apparatus of FIG. 12.
[0024] FIGS. 16A-16C are views showing operations of mechanisms at
a stop of the engine. FIG. 16A is an illustrative view showing a
state in which the vane member is controlled and rotated on the
most retarded position. FIG. 16B is an illustrative view showing a
state in which a lock piston is engaged in a lock hole. FIG. 16C is
an illustrative view showing a state in which a spool valve element
is held in a left side position.
[0025] FIGS. 17A-17C are views showing operations of the mechanism
when an ignition key is switched to an ON state. FIG. 17a is an
illustrative view showing a state in which the vane member is
controlled and rotated to the most retarded position. FIG. 17B is
an illustrative view showing a state in which the lock piston is
pulled out from the lock hole. FIG. 17C is an illustrative view
showing a state in which the spool valve element is held in the
left side position.
[0026] FIGS. 18A-18C are views showing operations of the mechanisms
when the engine is shifted to a middle engine speed. FIG. 18A is an
illustrative view showing a state in which the vane member is
controlled and rotated to the advance side. FIG. 18B is an
illustrative view showing a state in which the lock piston is
pulled out from the lock hole. FIG. 18C is an illustrative view
showing a state in which the spool valve element is held in the
right side position.
[0027] FIGS. 19A-19C are views showing operations of the mechanisms
when the engine is in the idling operation. FIG. 19A is an
illustrative view showing a state in which the vane member is
controlled and rotated to the retard side. FIG. 19B is an
illustrative view showing a state in which the lock piston is
pulled out from the lock hole. FIG. 19C is an illustrative view
showing a state in which the spool valve element is held in the
left side position.
[0028] FIG. 20 is a front view showing a variable displacement pump
according to a second embodiment of the present invention, from
which a cover member is detached.
[0029] FIG. 21 is an exploded perspective view showing a variable
displacement pump according to a third embodiment of the present
invention.
[0030] FIG. 22 is a front view showing the variable displacement
pump of FIG. 21 from which a cover member is detached, and in which
an eccentric amount of a cam ring is maximized.
[0031] FIG. 23 is a front view showing the variable displacement
pump of FIG. 21 from which the cover member is detached, and in
which the eccentric amount of the cam ring is minimized.
[0032] FIG. 24 is a longitudinal sectional view showing the
variable displacement pump of FIG. 21.
[0033] FIG. 25 is a front view showing an inside of a housing of
the variable displacement pump of FIG. 21.
[0034] FIG. 26 is a longitudinal sectional view showing a state in
which a solenoid valve is not energized in the variable
displacement pump of FIG. 21.
[0035] FIG. 27 is a longitudinal sectional view showing a state in
which the solenoid valve is energized in the variable displacement
pump of FIG. 21.
[0036] FIG. 28 is a view showing a hydraulic circuit in the
variable displacement pump of FIG. 21.
[0037] FIG. 29 is a graph showing a relationship between an engine
speed and a hydraulic pressure in the variable displacement pump of
FIG. 21.
[0038] FIG. 30 is a longitudinal sectional view showing a state in
which a solenoid valve is not energized in a variable displacement
pump according to a variation of the third embodiment of the
present invention.
[0039] FIG. 31 is a longitudinal sectional view showing a state in
which the solenoid valve is energized in the variable displacement
pump according to the variation of the third embodiment of the
present invention.
[0040] FIG. 32 is an exploded perspective view showing a variable
displacement pump according to a fourth embodiment of the present
invention.
[0041] FIG. 33 is a front view showing the variable displacement
pump of FIG. 32 from which a cover member is detached, and in which
an eccentric amount of the cam ring is maximized.
[0042] FIG. 34 is a front view showing the variable displacement
pump of FIG. 33 from which the cover member is detached, and in
which the eccentric amount of the cam ring is minimized.
[0043] FIG. 35 is a front view showing the cover member in the
variable displacement pump of FIG. 32.
[0044] FIG. 36 is a back view showing the cover member in the
variable displacement pump of FIG. 32.
[0045] FIG. 37 is a longitudinal sectional view showing a state in
which a hydraulic directional control valve is not activated in a
variable displacement pump according to a fifth embodiment of the
present invention.
[0046] FIG. 38 is a longitudinal sectional view showing a state in
which the hydraulic directional switching valve is activated in the
variable displacement pump according to the fifth embodiment of the
present invention.
[0047] FIG. 39 is a view showing a hydraulic circuit for the
variable displacement pump according to the fifth embodiment of the
present invention.
[0048] FIG. 40 is a graph showing a relationship between an engine
speed and a hydraulic pressure in the variable displacement pump
according to the fifth embodiment of the present invention.
DETAILED DESCRIPTION OF THE INVENTION
[0049] Hereinafter, variable displacement pumps according to
embodiments of the present invention are illustrated with reference
to drawings.
[0050] Variable displacement pump 01 according to embodiments are
arranged to supply a lubricating oil to sliding portions of an
internal combustion engine for a vehicle, to supply the lubricating
oil through an oil jet to a piston, and to supply the lubricating
oil to a valve timing control apparatus and a lock mechanism of the
valve timing control apparatus.
First Embodiment
[0051] As shown in FIGS. 1-3, variable displacement vane pump 01
includes a pump housing 1 which is provided at a front end portion
of a cylinder block of the internal combustion engine, and which
has an opening of a one end closed by a cover member 2; a drive
shaft 3 which penetrates a central portion within pump housing 1,
and which is a driven and rotated by a crank shaft of the engine; a
rotor 4 which is rotatably received within pump housing 1, and
which has a central portion connected with drive shaft 3; a cam
ring 5 which is a movable member swingably (pivotally) disposed
radially outside rotor 4; and a pair of vane rings 6 and 6 each of
which has a smaller diameter, and which is slidably disposed on
both side surfaces of rotor 4 on the inner circumference side.
[0052] Pump housing 1 is integrally formed from an aluminum alloy.
As shown in FIG. 4, pump housing 1 includes a bottom surface 1a
which is a recessed shape, and on which an axial one end surface of
cam ring 5 is arranged to be slid. Accordingly, bottom surface 1a
is formed to have a flatness, a surface roughness and so on which
have high accuracy. A sliding region (of bottom surface 1a) is
formed by a machine working (machining).
[0053] Moreover, pump housing 1 includes a hole which is formed at
a predetermined position on an inner circumference surface of pump
housing 1, and into which a one end portion of a pivot pin 9
serving as a pivot point of cam ring 5 about which cam ring 5 is
pivoted is inserted; and a pivot groove is which has a
semi-circular cross section, which is formed at a predetermined
position on an inner circumference surface of pump housing 1, and
into which a one end portion of pivot pin 9 is inserted.
Furthermore, pump housing 1 includes a seal surface 1s which has an
arc recessed shape, and which is formed on the inner circumference
surface of pump housing 1 on the left side of FIG. 2 at a position
above a line M (hereinafter, referred to as a cam ring reference
line) connecting a center of pivot pin 9 and a center of pump
housing 1 (a center of drive shaft 3).
[0054] A seal member 14 (described later) provided to cam ring 5 is
slid on seal surface 1s, so that seal surface 1s and also seal
member 14 seal one end of a control hydraulic chamber 16 (described
later) which is on upper end side in FIG. 2. As shown in FIG. 4,
this seal surface 1s is formed into an arc surface having a
predetermined radius.
[0055] As shown in FIG. 4, pump housing 1 includes a suction port 7
formed in bottom surface 1a on a left side of drive shaft 3, and a
discharge port 8 formed in bottom surface 1a on a right side of
drive shaft 3. Suction port 7 and discharge port 8 are disposed to
confront each other.
[0056] As shown in FIG. 4, suction port 7 is connected with a
suction opening 7a arranged to suck the oil within an oil pan (not
shown). Discharge port 8 is connected from a discharge opening 8a
through an oil main gallery (not shown) to the sliding portions of
the engine, the valve timing control apparatus which is a variable
valve actuating apparatus, and so on.
[0057] Suction port 7 includes an inside port portion 7b which has
an arc shape, and an outside port portion 7c which has a
substantially rectangular shape. Discharge port 8 includes an
inside port portion 8b which has an arc shape, and an outside port
portion 8c which is connected directly to discharge opening 8a.
[0058] Moreover, a bearing hole 1f for drive shaft 3 is formed at a
substantially central portion of bottom surface 1a. The oil
discharged from discharge port 8 is supplied to bearing hole 1f
through a tip end recessed groove 10a of a feeding (supplying)
groove 10 which has a small width, and which is formed into a
substantially L-shape. Furthermore, the oil is supplied from an
opening of feeding groove 10 to the both side surfaces of rotor 4,
and side surfaces of vanes 11 (described later) so as to ensure
lubricity.
[0059] As shown in FIGS. 1 and 3, cover member 2 is formed into a
plate shape with a large thickness. Cover member 2 includes an
inner side surface 2a which is a substantially flat surface. Cover
member 2 includes a suction port 7' and a discharge port 8' formed
in inner side surface 2a, like bottom surface 1a of pump housing 1.
Suction port 7' and discharge port 8' are connected, respectively,
with suction port 7 and discharge port 8. Cover member 2 includes a
pin hole 2b which is formed at an end portion (on the right side in
FIG. 3) of inner side surface 2a, and which receives the other end
portion of pivot pin 9. Furthermore, cover member 2 includes a
bearing hole 2c which is formed at a substantially central position
of cover member 2, which penetrates cover member 2, through which
drive shaft 3 is inserted, and which rotatably supports drive shaft
3.
[0060] This cover member 2 is positioned on pump housing 1 in the
circumferential direction through a plurality of positioning pins
1P shown in FIG. 1, and mounted on pump housing 1 by a plurality of
bolts B.
[0061] Drive shaft 3 is arranged to rotate rotor 4 in a clockwise
direction of FIG. 2 by the rotational force transmitted from the
crank shaft. A half on the left side of FIG. 2 around drive shaft 3
is a suction region. A half on the right side of FIG. 2 around
drive shaft 3 is a discharge region.
[0062] As shown in FIGS. 1 and 2, rotor 4 includes seven slots 4a
extending in the radial direction from the center side to the
outside. Each of seven vanes 11 is slidably held in one of seven
slots 4a, and arranged to be moved into and out of the one of seven
slots 4a. Rotor 4 includes back pressure chambers 12 each of which
is formed at a base end portion (radially inner end portion) of one
of slots 4a, each of which has a substantially circular cross
section, and each of which receives the discharge hydraulic fluid
discharged into discharge port 8. Furthermore, rotor 4 includes
recessed grooves 4b and 4b each of which has a circular (annular)
recessed shape, which are formed on both end surfaces of rotor 4 in
the axial direction, and which are arranged to hold vane rings 6
and 6 on the inner circumference side thereof so that vane rings 6
and 6 rotate in an eccentric state.
[0063] As shown in FIG. 2, each of vanes 11 includes an inside base
end (radially inner end portion) which is slidably abutted on outer
circumference surfaces of vane rings 6 and 6, and a tip end
(radially outer end) which is slidably abutted on an inner
circumference surface 5a of cam ring 5.
[0064] Moreover, a plurality of pump chambers 13 are liquid-tightly
separated between adjacent two of vanes 11, inner circumference
surface 5a of cam ring 5, the outer circumference surface of rotor
4, bottom surface 1a of pump housing 1, and inner side surface 2a
of cover member 2. Each of pump chambers 13 is an operation chamber
shaped like a sector. Each of vane rings 6 is arranged to push
vanes 11 in the radially outward direction.
[0065] Cam ring 5 is integrally formed into a substantially
cylindrical shape by an easily-worked sintered metal. Cam ring 5
includes a pivot raised portion 5b formed in a right side position
of cam ring 5 in FIG. 2, on the cam ring reference line, on the
outer circumference surface of cam ring 5. Cam ring 5 includes a
pivot support groove 5k which has a semi-circular cross section,
which extends in the axial direction, which is formed in a central
position of the outer surface of pivot raised portion 5b, and which
receives pivot pin 9 with pivot groove is so as to serve as an
eccentric swing point about which cam ring 5 is swung.
[0066] Moreover, cam ring 5 includes a boss portion 5c which has a
substantially inverse U-shape, and which is integrally formed with
cam ring 5 at a position above cam ring reference line X, that is,
an upper position on left side in FIG. 3. Cam ring 5 includes an
arc surface 5d which is an arc raised portion, which is formed on
an outside surface of boss portion 5c, and which confronts seal
surface 1s. Furthermore, cam ring 5 includes a holding groove 5e
which is a rectangular cross section, and which is formed in arc
surface 5d. Seal member 14 is mounted and fixed in holding groove
5e. Seal member 14 seals the one end side of control hydraulic
chamber 16. On the other hand, pivot support groove 5k of pivot
raised portion 5b of cam ring 5 and pivot pin 9 seal the other end
side of control hydraulic chamber 16. Arc surface 5d has a radius
of curvature which is identical to that of seal surface 1s to form
a minute constant clearance between arc surface 5d and seal surface
1s.
[0067] Seal member 14 is formed of, for example, a synthetic resin
with a low abrasion resistance (low abrasion quality). Seal member
14 has an elongated shape extending in the axial direction of cam
ring 5. Seal member 14 is pressed on seal surface 1s by a resilient
(elastic) force of a resilient (elastic) member 15 made of rubber,
and fixed on a bottom side of holding groove 5e. With this, good
liquid-tightness of control hydraulic chamber 16 is constantly
ensured.
[0068] As shown in FIGS. 1 and 3, cam ring 5 includes a pair of
suction side cutout grooves 18a and 18b which are formed, in the
circumferential direction, on the suction port 7's side, on both
axial end surfaces of cam ring 5. The pair of suction side cutout
grooves 18a and 18b introduce (flow) the oil in the suction region
into pump chambers 13. Moreover, cam ring 5 includes discharge side
cutout grooves 18c and 18d which are formed, in the circumferential
direction, on the discharge port 8's side, on the both axial end
surfaces of cam ring 5. Discharge side cutout grooves 18c and 18d
introduce (flow) the oil within pump chambers 13 in the discharge
region into discharge port 8.
[0069] Control hydraulic chamber 16 is separated in a substantially
arc shape between the outer circumference surface of cam ring 5,
pivot raised portion 5b and seal member 14. Control hydraulic
chamber 16 is arranged to act the discharge hydraulic pressure
introduced from discharge port 8, on a pressure receiving surface
5f in the outer circumference surface of cam ring 5 so that cam
ring 5 is swung (pivoted) about pivot pin 9 in the counterclockwise
direction of FIG. 2, and thereby to move cam ring 5 in a direction
in which an eccentric amount (eccentricity) of cam ring 5 with
respect to rotor 4 is decreased.
[0070] Cam ring 5 includes an arm 17 which is integrally formed
with cam ring 5 on the outer circumference surface of the
cylindrical body of cam ring 5 at a position opposite to pivot
raised portion 5b, and which protrudes in the radially outward
direction. As shown in FIGS. 1 and 2, this arm 17 includes an arm
body 17a which has a rectangular plate shape, and which extends in
the radial direction from the outer circumference surface of the
cylindrical body of cam ring 5; and a raised portion 17b integrally
formed on an upper surface of a tip end portion of arm body
17a.
[0071] Arm body 17a includes a protrusion 17c which has an arc
curved shape, and which is integrally formed with arm body 17a on a
lower surface of arm body 17a that is on a side opposite to raised
portion 17b. On the other hand, raised portion 17b extends in a
direction substantially perpendicular to arm body 17a. Raised
portion 17b includes an upper surface 17d formed into a curved
shape having a small radius of curvature.
[0072] A first spring receiving chamber 19 on a lower side of FIG.
2 and a second spring receiving chamber 21 on an upper side of FIG.
2 are formed at positions opposite to the position of pivot groove
1c of pump housing 1, that is, at upper and lower positions of arm
17. First spring receiving chamber 19 is formed coaxially with
second spring receiving chamber 21.
[0073] First spring receiving chamber 19 has a substantially
rectangular cross section. First spring receiving chamber 19
extends in the axial direction of pump housing 1. On the other
hand, second spring receiving chamber 21 has a length shorter than
a length of first spring receiving chamber 19. Like first spring
receiving chamber 19, second spring receiving chamber 19 has a
substantially rectangular cross section, and second spring
receiving chamber 21 extends in the axial direction of pump housing
1.
[0074] As shown in FIG. 5, in second spring receiving chamber 21,
there are provided a pair of retaining portions 23 and 23 which are
formed into an elongated rectangular plate, which are formed
integrally with pump housing 1 on inner side surfaces of second
spring receiving chamber 21 at a lower opening portion 21a of
second spring receiving chamber 21, and which extend in the inward
directions to confront each other in widthwise direction of lower
opening portion 21a. Raised portion 17b of arm 17 is arranged to be
moved into or out of second spring receiving chamber 21 through
opening portion 21a between retaining portions 23 and 23. Retaining
portions 23 and 23 are arranged to restrict a maximum expansion of
a second coil spring 22 described later.
[0075] A first coil spring 20 is received within first spring
receiving chamber 19. First coil spring 20 is arranged to urge cam
ring 5 through arm 17 in the clockwise direction of FIG. 2, that
is, to urge cam ring 5 in a direction to increase the eccentric
amount of the center of the inner circumference surface of cam ring
5 with respect to the center of the rotation of rotor 4.
[0076] First coil spring 20 is provided with a predetermined spring
load W3. First coil spring 20 includes a lower end abutted on a
bottom surface 19a of first spring receiving chamber 19, and an
upper end constantly abutted on arc protrusion 17c on the lower
surface of arm body 17a. First coil spring 20 is arranged to urge
cam ring 5 in the direction to increase the eccentric amount of the
center of the inner circumference surface of cam ring 5 with
respect to the center of the rotation of rotor 4, that is, in the
clockwise direction in FIG. 2.
[0077] A second coil spring 22 is received within second spring
receiving chamber 21. Second coil spring 22 is arranged to urge cam
ring 5 through arm 17 in the counterclockwise direction in FIG.
2.
[0078] This second coil spring 22 includes an upper end abutted on
an upper wall surface 21b of spring receiving chamber 21, and a
lower end abutted on raised portion 17b of arm 17 from the maximum
eccentric position of cam ring 5 in the clockwise direction, to a
position at which the lower end of this second coil spring 22 is
retained by retaining portions 23 and 23 so as to provide the
urging force in the counterclockwise direction of FIG. 2 to cam
ring 5.
[0079] That is, second coil spring 22 is also provided with a
predetermined spring set load in a direction to confront (opposite
to) first coil spring 20. This spring set load of second coil
spring 22 is set smaller than spring set load W3 of first coil
spring 20. With this, cam ring 5 is set to an initial position
(maximum eccentric position) by a spring load W1 which is a
difference of the spring set loads between first coil spring 20 and
second coil spring 22.
[0080] That is, first coil spring 20 and second coil spring 22
urges cam ring 5 through arm 17 in a direction in which cam ring 5
is constantly eccentric in the upward direction in a state in which
spring load W1 is provided, that is, in a direction in which
volumes of pump chambers 13 are increased. Spring load W1 is a load
at which cam ring 5 starts to move when the hydraulic pressure is
equal to or greater than a necessary hydraulic pressure P1
necessary for the valve timing control apparatus.
[0081] Second coil spring 22 is abutted on arm 17 when the
eccentric amount of cam ring 5 between the center of the inner
circumference surface of cam ring 5 and the center of the rotation
of rotor 4 is equal to or greater than a predetermined amount. When
the eccentric amount of cam ring 5 between the center of the inner
circumference surface 5a of cam ring 5 and the center of the
rotation of rotor 4 is smaller than the predetermined amount as
shown in FIG. 5, second coil spring 22 is retained by retaining
portions 23 and 23 to keep the compression state, and second coil
spring 22 almost does not contact arm 17. A spring load W2 of first
coil spring 20 at the swing movement amount of cam ring 5 at which
the spring load of second coil spring 22 to arm 17 becomes zero by
retaining second coil spring 22 by retaining portions 23 and 23 is
a load at which cam ring 5 starts to move when the hydraulic
pressure is a necessary hydraulic pressure P2 necessary for the
piston oil jet and so on, or the necessary hydraulic pressure P3 at
the maximum rotational speed of the crank shaft.
[0082] When cam ring 5 is pivoted in the clockwise direction by the
spring force of first coil spring 20 as shown in FIG. 2, the upper
surface of a join portion between arm body 17a and the cylindrical
body is abutted on a lower surface of one of retaining portions 23
and 23, so that the further rotation of cam ring 5 in the clockwise
direction is restricted. That is, the swing position of cam ring 5
is restricted to the initial set position (the maximum eccentric
position) by the spring force of first coil spring 20.
[0083] Hereinafter, a basic operation of variable displacement pump
01 according to the first embodiment is illustrated. Moreover, a
relationship between the control hydraulic pressure of the normal
variable displacement pump, and the necessary hydraulic pressure
necessary for the sliding portions of the engine, the valve timing
control apparatus, and the cooling of the piston is
illustrated.
[0084] When the valve timing control apparatus described later is
used for improving the fuel consumption and for countermeasure for
the exhaust air emission, the variable displacement pump is used as
the operation source for the valve timing control apparatus.
Accordingly, a high hydraulic pressure P1 shown by a broken line b
of FIG. 8 is needed from the low engine speed so as to improve the
operation responsiveness of the valve timing control apparatus.
[0085] Moreover, when oil jet 30 described later is used, high
hydraulic pressure P2 is needed at the middle engine speed. A
hydraulic pressure P3 is needed at the maximum engine speed, mainly
for lubricating the bearing portion of the crank shaft. Therefore,
the hydraulic pressure necessary for the entire of the internal
combustion engine is a characteristic of the entire of the broken
line connecting the broken lines b and c.
[0086] The relationship between middle engine speed necessary
hydraulic pressure P2 and high engine speed necessary hydraulic
pressure P3 is substantially P2<P3. Necessary hydraulic pressure
P2 is often near necessary hydraulic pressure P3. Accordingly, it
is desirable that the hydraulic pressure is not increased even when
the engine speed is increased from the middle engine speed to the
high engine speed in a region (D).
[0087] In this example, as shown by a solid line of FIG. 8, the
pump discharge pressure does not reach hydraulic pressure P1 from
the start of the internal combustion engine to the low engine speed
including the idling. Accordingly, arm body 17a of arm 17 of cam
ring 5 is abutted on the one of retaining portions 23 of housing 1
by the difference between the spring loads of first coil spring 20
and second coil spring 22, so that cam ring 5 is in the activation
stop state (cf. FIG. 2).
[0088] At this time, the eccentric amount of cam ring 5 is
maximized, and the pump capacity is maximized. The discharge
hydraulic pressure is suddenly increased in accordance with the
increase of the engine speed. Accordingly, the discharge hydraulic
pressure becomes a characteristic shown by (A) on the solid line of
FIG. 8.
[0089] Then, when the pump discharge hydraulic pressure is
increased in accordance with the increase of the engine speed and
reaches hydraulic pressure Pf shown in FIG. 8, the hydraulic
pressure introduced into control hydraulic chamber 16 is increased.
With this, cam ring 5 starts to compress first coil spring 20 acted
to arm 17, and cam ring 5 is pivoted about pivot pin 9 in the
counterclockwise direction to be eccentric. The hydraulic pressure
Pf is a first hydraulic pressure which is a cam activation pressure
for activating (starting the movement of) cam ring 5. The hydraulic
pressure Pf is set to be sufficiently greater than necessary
hydraulic pressure P1 of the valve timing control apparatus.
[0090] With this, the pump capacity is decreased. Accordingly, the
increase characteristic of the discharge hydraulic pressure is
decreased as shown in a region (B) of FIG. 8. Then, as shown in
FIG. 5, second coil spring 22 is retained by retaining portions 23
and 23 to keep the compression state. Cam ring 5 is swung in the
counterclockwise direction to a state in which the load of second
coil spring 22 is not acted to upper surface 17d of arm raised
portion 17b.
[0091] In a state shown in FIG. 5, the load of second coil spring
22 is not acted to cam ring 5 from this instant. Cam ring 5 can not
be swung and becomes a retained state until the discharge hydraulic
pressure reaches the hydraulic pressure P2 (hydraulic pressure P2
within control hydraulic chamber 16), and becomes greater than
spring load W2 of first coil spring 20. Accordingly, the discharge
hydraulic pressure becomes an increasing characteristic shown by
(C) of FIG. 8 in accordance with the increase of the engine speed.
However, the discharge hydraulic pressure does not become sudden
increase shown by (A) of FIG. 8 since the pump capacity is
decreased by the decrease of the eccentric amount of cam ring
5.
[0092] Moreover, when the discharge pressure increases equal to or
greater than hydraulic pressure Ps (P2) by the increase of the
engine speed, cam ring 5 is swung against the spring force of
spring load W2 of first coil spring 20 through arm 17 to compress
first coil spring 20, as shown in FIG. 6. Accordingly, the pump
capacity is further decreased in accordance with the swing movement
of cam ring 5, so that the increase of the discharge hydraulic
pressure becomes small. Then, the engine speed reaches the maximum
engine speed while the discharge hydraulic pressure keeps the state
of the characteristic shown by (D) of FIG. 8.
[0093] Accordingly, the discharge pressure (solid line) at the high
rotational speed of the pump sufficiently approaches the necessary
(requested) hydraulic pressure (broken line). Therefore, it is
possible to effectively suppress the power loss.
[0094] FIG. 7 shows a relationship between displacements of first
and second coil springs 20 and 22 or the swing angle of cam ring 5,
and spring loads W1 and W2. That is, the spring force of set load
W1 of first coil spring 20 is provided to cam ring 5 at the initial
state from the start of the internal combustion engine to the low
engine speed. Accordingly, cam ring 5 can not be moved until the
discharge pressure exceeds spring load W1. After the discharge
pressure exceeds spring load W1, first coil spring 20 is
compressed, so that the load of first coil spring 20 is increased.
On the other hand, second coil spring 22 approaches a free length
so that the load of second coil spring 22 is decreased.
Consequently, the spring load is increased. This inclination is a
spring constant.
[0095] At a position of cam ring 5 in FIG. 5, the load becomes
spring load W2 of first coil spring 20. The spring load W2 of first
coil spring 20 increases in a discontinuous manner. After the
discharge hydraulic pressure exceeds spring load W2, first coil
spring 20 is compressed, so that the load (of first coil spring 20)
is increased. However, only one coil spring is acted. Therefore,
the spring constant is decreased, so that the inclination is
varied.
[0096] As described above, when the discharge hydraulic pressure
reaches hydraulic pressure P1 by the increase of the internal
combustion engine speed, cam ring 5 starts to move to suppress the
increase of the discharge hydraulic pressure. When cam ring 5 is
moved by a predetermined movement amount in the counterclockwise
direction as shown in FIG. 5, the spring force of second coil
spring 22 is eliminated, and the spring constant becomes small.
Moreover, the spring load W2 of first coil spring 20 is increased
in the discontinuous manner. Accordingly, the swing movement of cam
ring 5 starts again after the discharge hydraulic pressure is
increased to the hydraulic pressure P2 (cf. FIG. 6). That is, the
relative spring load of first and second coil springs 20 and 22 are
acted, and the spring characteristic is non-linear state. With
this, cam ring 5 acts a special swing movement.
[0097] In this way, the characteristic of the discharge hydraulic
pressure becomes characteristics shown by (A)-(D) of FIG. 8 by the
non-linear characteristic of the spring forces of coil springs 20
and 22. Accordingly, the control hydraulic pressure (solid line)
can sufficiently approach the necessary (requested) hydraulic
pressure (broken line). Therefore, it is possible to sufficiently
decrease the power loss caused by the unnecessary increase of the
hydraulic pressure.
[0098] Next, an oil jet 30 according to the first embodiment of the
present invention is illustrated.
[0099] As shown in FIG. 9, oil jet 30 is provided to an internal
combustion engine 31. In the internal combustion engine 31, a crank
shaft 34 is rotatably supported by a bearing (not shown) within a
crank chamber 33 separated by a crank case of a cylinder block 32.
Moreover, a piston 36 is slidably disposed within a cylindrical
cylinder wall 37 formed at an upper portion of the crank case.
Piston 36 is connected through a con rod 35 to crank shaft 34.
[0100] Within a wall of cylinder wall 37, there is formed a water
jacket 37a in which the cooling water is circulated. Within a
partition wall 38 between the crank case and cylinder wall 37,
there is formed a main oil gallery 39 arranged to supply the oil
(the lubricating oil) discharged from variable displacement pump 01
to the sliding portions of the engine.
[0101] Within the lower portion of partition wall 38, there is
formed a connection passage 38a which is connected with main oil
gallery 39, and which extends in the upward and downward
directions, as shown in FIGS. 11A and 11B. In the lower portion of
connection passage 38a, there is formed a mounting hole including
an internal-thread portion 38b formed on an inner circumference
surface of the mounting hole.
[0102] Oil jet 30 is attached (mounted) at the lower portion of
partition wall 38. Oil jet 30 is arranged to inject the oil for the
lubrication and the cooling, to a portion between the inner
circumference surface of cylinder wall 37 and piston 36.
[0103] As shown in FIGS. 10, 11A and 11B, this oil jet 30 includes
a cylindrical holding member 40 made of an aluminum alloy; a valve
body 41 formed into a cylindrical shape, and inserted from the
below into an insertion hole 40a formed within holding member 40; a
protruding portion 42 which is integrally formed in the outside
portion of holding member 40, and which serves as a positioning
member; and a nozzle 43 formed on the outside portion of holding
member 40 at a position opposite to protruding portion 42.
[0104] Holding member 40 includes an annular passage portion 44
formed between insertion hole 40a and an outer circumference
surface of valve body 41; and a mounting groove 40b which is formed
in the outside portion of holding member 40, and in which a base
end portion 43a of nozzle 43 is mounted and fixed.
[0105] Valve body 41 is formed of an iron metal such as sintered
alloy. Valve body 41 includes an external thread portion 41a which
is formed on an outer circumference surface of the upper portion of
valve body 41, and which is screwed into internal thread portion
38b. Moreover, valve body 41 includes an oil supply hole 45 that is
a hydraulic fluid supplying portion which is formed within the
upper end portion of valve body 41, which extends in the axial
direction, and which is connected with connection passage 38a; an
oil introduction hole 47 which is formed on a lower side of oil
supply hole 45, which is connected with oil supply hole 45 in a
continuous manner, and which is arranged to movably hold a ball
valve element 46; and a seat surface 45a which is formed into an
annular shape, which is formed in a stepped portion between oil
supply hole 45 and oil introduction hole 47, and on which valve
ball element 46 is arranged to be seated.
[0106] Furthermore, valve body 41 includes a plurality of radial
holes 48 formed along the diameter direction in a circumferential
wall of the lower end portion of valve body 41. Radial holes 48
connect oil introduction hole 47 and passage portion 44. Valve body
41 includes a flange portion 41b integrally formed with the outer
circumference of the lower portion of valve body 41. This flange
portion 41 is arranged to press and fix base end portion 43a of
nozzle 43 and also holding member 40 on partition wall 38 when
valve body 41 is screwed and fixed in partition wall 38 through
external thread portion 41a and internal thread portion 38b.
[0107] Protruding portion 42 is mounted in a positioning hole 38c
formed in partition wall 38 when holding member 40 is fixed in
partition wall 38 through valve body 41, so as to position holding
member 40, and to prevent the rotation of holding member 40.
[0108] Nozzle 43 is raised (extends) in an inclined state from a
base end portion 43a located on the holding member 40's side to a
tip end portion 43b. Nozzle 43 is disposed so that tip end portion
43b is located in a lower portion within cylinder wall 37. Nozzle
43 includes an elongated hydraulic hole 43c which extends within
nozzle 43 in the axial direction, and which has a one end portion
opened to passage portion 44; and a nozzle portion 43d formed at
the tip end portion of hydraulic hole 43c, and arranged to direct
the lower portion of piston 36.
[0109] A valve spring 50 is held by a plug-shaped retainer 49 fit
in the lower end portion of oil introduction hole 47 by the press
fit. Valve spring 50 is an urging member having a coil shape. Ball
valve element 46 is urged by the urging force (the spring force) of
valve spring 50 in a direction in which ball valve element 46 is
seated on seat surface 45a, that is, in a direction to close the
opening of the lower end of oil supply hole 45.
[0110] The spring load of valve spring 50 (that is, the valve
opening pressure of ball valve element 46) is set to pressure P2
(cf. FIG. 8) which is sufficiently larger than first discharge
pressure Pf of variable displacement pump 01, and which is slightly
smaller than second discharge pressure Ps that is a working
(operating) pressure of cam ring 5.
[0111] Hereinafter, an operation of oil jet 30 is illustrated.
First, drive shaft 3 of variable displacement pump 01 is rotated in
response to the start of the engine, so that the pressurized oil is
supplied to main oil gallery 39 to lubricate the sliding portions
of the engine. In the initial state of the start of the engine, the
pump discharge pressure is first discharge pressure Pf, as shown in
FIG. 8. Accordingly, ball valve element 46 is seated on seat
surface 45a by the spring force of valve spring 50 to keep the
valve closed state, as shown in FIG. 11A.
[0112] Then, when the pump discharge pressure is increased and the
hydraulic pressure within oil supply hole 45 becomes equal to or
greater than the hydraulic pressure P2, that is, the spring load of
valve spring 50, valve spring 50 is compressed to open ball valve
element 46, as shown in FIG. 11B. With this, oil supply hole 45 is
connected with oil introduction hole 47, the oil supplied from main
oil gallery 39 through connection passage 38a to oil supply hole 45
flows (enters) from oil introduction hole 47 through radial holes
48 into connection passage 44. Moreover, the oil flows through
hydraulic hole 43c of nozzle 43, and the oil is injected from
nozzle portion 43d into the inside from the lower direction of
piston 36, as shown in FIG. 9.
[0113] In this way, the oil discharged from variable displacement
pump 01 is not injected from oil jet 30 to piston 36 until the
discharge pressure of the oil becomes equal to or greater than
first discharge pressure Pf, and reaches the hydraulic pressure P2
slightly smaller than second discharge pressure Ps. Therefore, it
is possible to effectively suppress the energy loss in the initial
stage of the pump discharge of variable displacement pump 01.
[0114] Moreover, as described above, first discharge pressure Pf is
set smaller than the valve opening pressure of ball valve element
46 of oil jet 30. Accordingly, oil jet 30 does not inject the oil
in the engine speed region (normal region) which is used at the
normal running of the vehicle. Therefore, the oil supply amount to
the sliding portions of the engine is increased while the excessive
oil discharge amount of the pump is suppressed. Therefore, it is
possible to decrease the friction of the pump and the internal
combustion engine, and to improve the fuel consumption.
[0115] Moreover, at the cold state of internal combustion engine
31, it is possible to suppress the injection of the low temperature
oil by oil jet 30 to piston 36. Therefore, it is possible to
improve the warm-up characteristic, and to decrease the exhaust
emission.
[0116] The structure of oil jet 30 is not limited to the structure
of the above-described embodiment. Holding member 40 may be formed
integrally with nozzle 43. Nozzle 43 may be fixed to valve body 41
by the brazing. Moreover, a plunger may be employed as the valve
element, in place of the ball.
[0117] Next, the valve timing control apparatus is illustrated
below.
[0118] This valve timing control apparatus is applied to the intake
side. As shown in FIGS. 12-15, this valve timing control apparatus
includes a timing sprocket 51 which is a driving rotational member
driven and rotated through a timing chain by the crank shaft (not
shown) of the engine; a cam shaft 52 arranged to rotate relative to
timing sprocket 51; a vane member 53 which is a driven rotational
member fixed at an end portion of cam shaft 52, and rotatably
received within timing sprocket 51, and a hydraulic pressure supply
and discharge mechanism 54 arranged to rotate vane member 53 in a
normal direction or in a reverse direction by the hydraulic
pressure.
[0119] Timing sprocket 51 includes a housing 55 having a teeth
portion 55a integrally formed on an outer circumference of housing
55, and engaged with the timing chain, and which rotatably receives
vane member 53; a front cover 56 closing an opening of a front end
of housing 55; and a rear cover 57 closing an opening of the rear
end of housing 55. These housing 55, front cover 56 and rear cover
57 are integrally fixed by four small diameter bolts 58 from the
axial direction of the cam shaft.
[0120] Housing 55 has a cylindrical shape having front and rear
both ends each having an opening. Housing 55 includes four
partition wall portions 60 which are shoes that are arranged on the
inner circumference surface at regular intervals of 90 degrees in
the circumferential direction, and that protrudes in the radially
inward direction. Each of partition wall portions 60 has a
substantially trapezoid cross section. Each of partition wall
portions 60 extends in the axial direction of housing 55. Each of
partition wall portions 60 has both axial end surfaces which are
same planes with both end surfaces of housing 55. Moreover, each of
four partition wall portions 60 includes a bolt insertion hole 61
which is located at a substantially central position of the each of
partition wall portions 60, and into which one of bolts 58 is
inserted. Each of partition wall portions 60 includes an inner end
surface (radially inner surface) formed into an arc shape to
correspond to an outer circumference of vane rotor 64 (described
later) of vane member 53, and a holding groove which is formed on
the inner end surface, and which extends in the axial direction. A
U-shaped seal member 62 and a plate spring (not shown) arranged to
press seal member 62 in the inward direction are fit in and held by
the holding groove of the each of partition wall portions 60.
[0121] Front cover 56 includes a bolt insertion hole 56a which has
a relatively large diameter, and which is formed at a substantially
central portion of front cover 56; and four bolt holes which are
formed in the outer circumference portion of front cover 56, and
each of which is connected with one of bolt insertion holes 61 of
housing 55.
[0122] Rear cover 57 includes a bearing hole 57a which is formed at
a substantially central portion of rear cover 57, and which
rotatably supports a front end portion 52a of cam shaft 52; and
four internal thread holes which are formed in an outer
circumference portion of rear cover 57, and into which one of bolts
58 is screwed.
[0123] Cam shaft 52 is rotatably supported by an upper end portion
of the cylinder head through a cam bearing (not shown). Cam shaft
52 includes a cam which is integrally formed on the outer
circumference surface of cam shaft 52 at a predetermined position,
and which is arranged to open an intake valve (not shown) through a
valve lifter.
[0124] Vane member 53 is integrally formed by the sintered alloy.
Vane member 53 includes an annular (circular) vane rotor 64 located
at a central portion, and fixed at a front end portion of cam shaft
52 by cam bolt 63; four vanes 65 integrally formed with vane rotor
64, and arranged on the outer circumference surface of vane rotor
64 at intervals of 90 degrees in the circumferential direction.
Vane rotor 64 includes an axial hole 64a which is located at a
substantially central position of vane rotor 64, and into which cam
bolt 63 is inserted; and a mounting groove 64b in which front end
portion 52a of cam shaft 52 is inserted and mounted. Vane rotor 64
is fixed on front end portion 52a of cam shaft 52 by cam bolt 63
from the axial direction.
[0125] One of four vanes 65 has a substantially trapezoid shape
having a large circumference width in the substantially
circumferential direction. Each of the other three of the four
vanes 65 has an elongated rectangular shape. These four vanes 65
are disposed in the circumferential direction at predetermined
angular positions to attain the weight balance of the entire of
vane member 53. Moreover, each of vanes 65 is disposed between
adjacent two of partition wall portions 60. Each of vanes 65 has a
holding groove formed at a central portion of the outer
circumference surface. A U-shaped seal member 66 and a plate spring
66a are fit and mounted in each of the holding grooves. Seal member
66 is slidably abutted on the inner circumference surface of
housing 55. Plate spring 66a is arranged to push seal member 66
toward the inner circumference surface of housing 55.
[0126] Moreover, an advance fluid pressure chamber 67 and a retard
fluid pressure chamber 68 are formed, respectively, on both sides
of each vane 65. Accordingly, four advance fluid pressure chambers
67 and four retard fluid pressure chambers 68 are separated between
vanes 65 and partition wall portions 60.
[0127] As shown in FIG. 12, hydraulic pressure supply and discharge
mechanism 54 includes two hydraulic passage systems including a
first hydraulic passage 69 arranged to supply and discharge the
hydraulic pressure of the lubricating oil, to and from advance
fluid pressure chambers 67; a second hydraulic passage 70 arranged
to supply and discharge the hydraulic pressure to and from retard
fluid pressure chambers 68. First and second hydraulic passages 69
and 70 are connected through a flow passage switching valve 73 to a
supply passage 71 and a drain passage 72 which are main oil
galleries for supplying the lubricating oil for the engine. One-way
variable displacement pump 01 arranged to pressurize the hydraulic
fluid within oil pan 74, and to supply this pressurized hydraulic
fluid is provided to supply passage 71. Moreover, a lower end of
drain passage 72 is connected with oil pan 74.
[0128] As shown in FIGS. 12 and 13, first hydraulic passage 69 is
formed between flow passage switching valve 73 and each of advance
fluid pressure chambers 67. First hydraulic passage 69 includes a
first passage portion 69a formed from the inside of the cylinder
head to the inside of the cam bearing and cam shaft 52 in the axial
direction; and four bifurcated passages 69b which are formed by
bifurcating within vane rotor 64 in the substantially radial
directions from the grooves on the front end side of cam shaft 52,
and each of which connects first passage portion 68a and one of
advance fluid pressure chambers 67.
[0129] On the other hand, second hydraulic passage 70 is formed
between flow passage switching valve 73 and each of retard fluid
pressure chambers 68. Second hydraulic passage 70 includes a second
passage portion 70a formed from the inside of the cylinder head to
the insides of the cam bearing and cam shaft 52 in the axial
direction; and four second bifurcated passages 70b which are formed
by bifurcating in the radial direction from the radial hole of cam
shaft front end portion 52a to the inside of vane rotor 64, and
each of which connects second passage portion 70a and one of retard
fluid pressure chambers 68.
[0130] A phase varying mechanism is constituted by vane member 53,
housing 55, advance fluid pressure chambers 67, retard fluid
pressure chambers 68, and hydraulic pressure supply and discharge
mechanism 54.
[0131] As shown in FIG. 12, flow passage switching valve 73 is a
solenoid valve which is 4-port and 2-position type. Flow passage
switching valve 73 includes a valve body 77 which has a cylindrical
shape with a bottom, and which is fixed within valve hole 76 formed
within the cylinder head; a solenoid 78 integrally fixed to one end
portion of valve body 77; and a spool valve element 79 slidably
disposed within valve body 77.
[0132] Valve body 77 includes a supply port 80 located at a
substantially central position in the axial direction, and arranged
to connect supply passage 71 and the inside of valve body 77; and
first and second ports 81 and 82 which are located on both sides of
supply port 80 in the axial direction, which are arranged to
connect, respectively, the end portions of first hydraulic passage
69 and second hydraulic passage 70, and the inside of valve body
77, and which extend in the radial direction. Moreover, valve body
77 includes first and second drain ports 83 and 84 which are
formed, respectively, on both sides of first and second ports 81
and 82, and which connect, respectively, the inside of valve body
77 and drain passage 72.
[0133] Solenoid 78 includes an electromagnetic coil 78b provided
within a solenoid casing 78a; a fix core 78c arranged to be excited
by an energization to electromagnetic coil 78b; a movable plunger
78d arranged to be slid by the excitation of fix core 78c, and
thereby to push and move spool valve element 79. Electromagnetic
coil 78b is connected through a harness (not shown) to an
electronic controller 86.
[0134] Spool valve element 79 includes a first land portion 79a
located at a substantially central portion of spool valve element
79, and arranged to open and close supply port 80 in accordance
with the sliding position of spool valve element 79 in the axial
direction; and second and third land portions 79b and 79c disposed
on both sides of first land portion 79a in the axial direction, and
arranged to relatively open and close first and second ports 81 and
82 and drain ports 83 and 84. Moreover, this spool valve element 79
is urged to a maximum left position, that is, a position to connect
supply port 80 and second port 82, and to connect first port 81 and
drain port 83, by a spring force of a return spring 85 mounted
between a spring retainer 77a provided on the other end side of
valve body 77, and an outer end surface of third land portion 79c.
Furthermore, spool valve element 79 is arranged to be controlled to
move against the spring force of return spring 85, to a maximum
right position or a predetermined central position, by a control
current from electronic controller 86.
[0135] Electronic controller 86 is configured to sense a current
driving state by signals from a crank angle sensor (not shown)
arranged to sense the engine speed, and an air flow meter arranged
to sense an intake air amount, and various sensors such as a
throttle opening sensor and a water temperature sensor arranged to
sense a water temperature of the engine.
[0136] This electronic controller 86 is configured to perform a
switching control of the flow passages by applying or breaking
(cutting off) a pulse control current to electromagnetic coil 78a
of flow passage switching valve 73, in accordance with the driving
state of the engine.
[0137] Moreover, between vane 65 with the maximum width and housing
55, there is provided a lock mechanism 87 arranged to restrict the
rotation of vane member 53 with respect to housing 55, or to
release the restriction of the rotation of vane member 53.
[0138] As shown in FIGS. 12 and 15, this lock mechanism 87 includes
a sliding hole 88 formed between vane 65 with the large width and
rear cover 57, and formed within vane 65 in the axial direction of
cam shaft 52; a lock piston 89 which is a cylindrical shape with a
cover, and which is slidably provided within sliding hole 88; a
lock hole 90a which is formed in an engagement hole forming section
90 that has a cup-shaped cross section, and that is fixed in a
fixing hole formed in rear cover 57, and which is an abutment
portion which a taper tip end portion 89a of lock piston 89 is
engaged with or disengaged from; a coil spring 92 which is held by
a spring retainer 91 fixed on the bottom surface side of sliding
hole 88, and which is a third urging member arranged to urge lock
piston 89 toward lock hole 90a.
[0139] Lock piston 89 includes a large diameter flange 89b which is
integrally formed with the outer circumference on the rear end side
of lock piston 89, and which is arranged to receive the pressure;
and a tip end portion 89a arranged to be engaged with lock hole 90a
by the spring force of coil spring 92 at a position at which vane
member 53 is rotated to the most retarded position, and thereby to
lock the relative rotation between timing sprocket 51 and cam shaft
52.
[0140] As shown in FIG. 15, lock piston 89 is moved in a backward
direction against the spring force of coil spring 92 by the
hydraulic pressure supplied from advance fluid pressure chambers 67
into lock hole 90a through a first hydraulic hole 93a formed in
vane member 53, or by the hydraulic pressure supplied from retard
fluid pressure chambers 68 into a pressure receiving chamber 89c
between the large diameter flange portion 89a and the stepped
portion of sliding hole 88, through a second hydraulic hole 93b
formed in vane member 53, so as to release the engagement with lock
hole 90a.
[0141] Coil spring 92 serves as a lock state holding mechanism to
hold the lock state between vane member 53 and housing 55. Coil
spring 92 has a spring force which is set to a value by which the
air accumulated in retard fluid pressure chambers 68 at the start
of the engine is not largely compressed and deformed by the
pressure compressed by the pressurized hydraulic pressure supplied
from variable displacement pump 01, and which is set to a value by
which the air accumulated in retard fluid pressure chambers 68 is
compressed and deformed when the discharged hydraulic pressure
reaches the hydraulic pressure Px in the initial state (A) shown in
FIG. 8.
[0142] Hereinafter, the operation of the valve timing control
apparatus is illustrated with reference to FIGS. 16-19. The
operation of variable displacement pump 01 is stopped at the stop
of the engine, so that the supply of the hydraulic pressure to
advance fluid pressure chambers 68 and retard fluid pressure
chambers 68 is stopped. As shown in FIGS. 13 and 16A, vane member
53 is rotated in a direction opposite to the rotational direction
of cam shaft 52 (shown by an arrow) to the most retarded side, by
the alternating torque generated in advance in cam shaft 52
immediately after the stop of the engine.
[0143] At this instant, tip end portion 89a of lock piston 89 of
lock mechanism 87 is engaged with lock hole 90a by the spring force
of coil spring 92 as shown in FIG. 16B, so as to restrict the free
rotation of vane member 53.
[0144] Moreover, the energization from electronic controller 86 to
flow passage switching valve 73 is cut off (shut off). Accordingly,
spool valve element 79 is urged to the maximum left side position
by the spring force of return spring 85, as shown in FIG. 16C.
[0145] Next, when the ignition key is switched to the ON state to
start the engine, the control current from electronic controller 86
is not outputted to electromagnetic coil 78b for a few seconds from
the start of the cranking. Accordingly, spool valve element 79 is
urged to the maximum left side position by the spring force of
return spring 85, as shown in FIG. 17C. Therefore, supply port 80
and second port 82 are connected with each other, and second land
portion 79b closes second drain port 84. At the same time, first
and third land portions 79a and 79c connect first port 81 and first
drain port 83.
[0146] Accordingly, the hydraulic pressure (discharge pressure)
discharged from variable displacement pump 01 flows from supply
passage 71 through supply port 80 into valve body 77, as shown by
arrows in FIG. 17C. Then, this hydraulic pressure flows from second
port 82 directly to second hydraulic passage 70. This hydraulic
pressure flows through second bifurcated passage 70b into retard
fluid pressure chambers 68.
[0147] Accordingly, as shown in FIG. 17A, vane member 53 is held,
by the low hydraulic pressure supplied into retard fluid pressure
chambers 68, to a state in which vane member 53 is positioned in
the most retarded side. Consequently, it is possible to improve the
performance of the start of the engine.
[0148] In this case, the air accumulated in retard fluid pressure
chambers 68 is pressurized by the low hydraulic pressure, so that
the air accumulated in retard fluid pressure chambers 68 pushes
vane member 53 to the most retarded side with the low hydraulic
pressure.
[0149] On the other hand, when the internal pressure of retard
fluid pressure chambers 68 is increased, this hydraulic pressure is
supplied from second hydraulic hole 93b to pressure receiving
chambers 89c, and acted to the pressure receiving surface of large
diameter flange 89b. With this, as shown in FIG. 17B, lock piston
89 is moved in the backward direction against the spring force of
coil spring 92, and pulled out from lock hole 90. Consequently,
vane member 53 is released from the lock state to allow the free
rotation. However, vane member 53 is held to the maximum retarded
position like the stop of the engine since the hydraulic pressure
within retard fluid pressure chambers 68 is high.
[0150] This timing at which end portion 89a of lock piston 89 is
pulled out from lock hole 90a is a timing at which the discharge
hydraulic pressure characteristic of variable displacement pump 01
becomes discharge pressure Px which is lower than first discharge
pressure Pf, and which is at the sudden increase before the
compression of first coil spring 20 in the region (A) of FIG. 8,
that is, a timing at which two or three seconds elapsed from the
switching of the ignition key to the ON state.
[0151] Then, when the engine speed becomes, for example, the middle
engine speed region after the start of the cranking, electronic
controller 86 energizes electromagnetic coil 78b of flow passage
switching valve 73 so as to excite fix core 78c. With this, spool
valve element 79 is moved in the right direction from the position
shown in FIG. 17C through movable plunger 78d to the maximum right
side position shown in FIG. 18C. Consequently, spool valve element
79 closes the connection between first port 81 and first drain port
83, and connects supply port 80 and first port 81. At the same
time, spool valve element 79 connects second port 82 and second
drain port 84.
[0152] Accordingly, as shown in FIG. 18C, the discharge hydraulic
pressure of variable displacement pump 01 flows from supply passage
71 into supply port 80 and valve body 77. Then, this discharge
hydraulic pressure flows from first port 81 into first passage
portion 69a of first hydraulic passage 69. This discharge hydraulic
pressure is supplied through bifurcated passage 69b into advance
fluid pressure chambers 67 to increase the pressure in the advance
fluid pressure chambers 67. On the other hand, the hydraulic fluid
within retard fluid pressure chambers 68 is returned from second
hydraulic passage 70 and so on through second drain port 84 to oil
pan 74 so as to decrease the pressure within retard fluid pressure
chambers 68.
[0153] Accordingly, in lock piston 89, the hydraulic pressure of
pressure receiving chamber 89c is lowered. However, as shown in
FIG. 18B, lock piston 89 is held to a state in which lock piston 89
is pulled out from lock hole 90a against the spring force of coil
spring 92 by the high hydraulic pressure supplied into lock hole
90a from first hydraulic hole 93a in accordance with the increase
of the hydraulic pressure of advance fluid pressure chambers 67.
With this, as shown in FIG. 18A, vane member 53 is rotated by the
high hydraulic pressure of advance fluid pressure chambers 67 from
the position shown in FIG. 13 in the rightward direction, that is,
in a direction identical to the rotational direction of cam shaft
52, so that the relative rotational phase between the crank shaft
and cam shaft 52 is rapidly varied to the advanced side.
[0154] Accordingly, the valve overlap between the intake valve and
the exhaust valve is slightly increased. Therefore, it is possible
to decrease the discharge amount of HC in the exhaust gas by the
effect of the internal EGR, as described later.
[0155] Moreover, when the engine is shifted, for example, to the
high engine speed region, the energization from electronic
controller 86 to electromagnetic coil 78b is held, the hydraulic
pressure is constantly supplied to advance fluid pressure chambers
67. Accordingly, vane member 53 is further rotated in the same
direction, and held in the maximum rotation position as shown in
FIG. 14. With this, the relative rotational phase between the crank
shaft and cam shaft 52 is varied to the most advanced position.
Consequently, the valve overlap becomes large, and it is possible
to improve the output (power) of the engine.
[0156] Moreover, when the operation of the engine is shifted to the
idling operation, the control current from electronic controller 86
to electromagnetic coil 78b is cut off. Accordingly, as shown in
FIG. 19C, spool valve element 79 is moved in the maximum left
direction by the spring force of return spring 85. Consequently,
supply port 80 and second port 82 are connected with each other,
and second land portion 79b closes second drain port 84. At the
same time, first and third land portions 79a and 79c connect first
port 81 and first drain port 83.
[0157] Accordingly, the hydraulic pressure discharged from variable
displacement pump 01 flows from supply passage 71 through supply
port 80 into valve body 77, as shown by an arrow of FIG. 19C. This
hydraulic pressure flows from second port 82 directly to second
hydraulic passage 70. Then, the hydraulic pressure is supplied
through second bifurcated passages 70b into retard fluid pressure
chambers 68. On the other hand, the hydraulic pressure of advance
fluid pressure chambers 67 is discharged from first hydraulic
passage 69 through first port 81, first drain port 83, and drain
passage 72 into oil pan 74, so that advance fluid pressure chambers
67 are brought into the low pressure state.
[0158] In this case, lock piston 89 is held to a pulled-out state
in which lock piston 89 is pulled out from lock hole 90e, by the
hydraulic pressure in pressure receiving chamber 89c which receives
the high pressure within retard fluid pressure chambers 68, as
shown in FIG. 19B. Accordingly, the free rotation of vane member 53
is allowed. Consequently, this vane member 53 is pivoted in the
most retarded side by the high hydraulic pressure supplied into the
retard angle chambers 68, as shown in FIG. 19A. With this, it is
possible to improve the combustion, and to improve the stability of
the idle operation of the engine.
[0159] As described above, in this example, it is possible to
improve the responsiveness of the operation of the valve timing
control apparatus at the start of the engine by the special
structures by using first and second coil springs 20 and 22 of
variable displacement pump 01.
[0160] That is, variable displacement pump 01 supplies the
lubricating oil discharged from the discharge opening through
discharge port 8, to the sliding parts of the engine. Moreover,
variable displacement pump 01 is also used as the source of the
operation of the valve timing control apparatus. In variable
displacement pump 01, it is possible to improve the initial
increase of the initial discharge hydraulic pressure (region (A))
shown in FIG. 8, as described above. Accordingly, it is possible to
improve, for example, the responsiveness of the relative rotational
phase between timing sprocket 55 and cam shaft 52 to the retarded
side immediately after the start of the engine.
Second Embodiment
[0161] FIG. 20 shows a variable displacement pump according to a
second embodiment of the present invention. The variable
displacement pump according to the second embodiment has a basic
structure substantially identical to the basic structure such as
the pump constituting section of the variable displacement pump
according to the first embodiment. However, in the second
embodiment, there are provided two control hydraulic chambers 16
which are arranged to push cam ring 5 to increase the eccentric
amount, and which are located on upper and lower sides of pivot pin
9.
[0162] That is, the control hydraulic chamber in the first
embodiment represents first hydraulic chamber 16a. Moreover, in
this example, pump housing 1 includes a recessed groove 24 which is
formed on the lower side of pivot pin 9, and which has a
substantially L-shape. This recessed groove 24 constitutes (forms)
a second control hydraulic chamber 16b. Furthermore, in the lower
portion of recessed groove 24, there is formed a second seal
surface 24a. This second seal surface 24a is formed into an arc
shape around a center of pivot pin 9.
[0163] On the other hand, cam ring 5 includes a raised portion 25
which is formed integrally with cam ring 5 at a portion to confront
recessed groove 24, and which has a substantially triangular shape.
Raised portion 25 includes a second arc surface 25a which is
located in a portion to confront second seal surface 24a, and which
has an arc shape around the center of pivot pin 9. Second arc
surface 25a includes a holding groove which is located at a tip end
portion of second arc surface 25a, and which has a substantially
rectangular cross section. The holding groove of second arc surface
25a receives a seal member 26 slidably abutted on seal surface 24a,
and a resilient (elastic) member 27 which has a rectangular cross
section, and which pushes seal member 26 toward second seal surface
24a.
[0164] Seal surface 24a has a length of the arc by which seal
member 26 can be slidably abutted on seal surface 24a when the
eccentric amount of cam ring 5 with respect to the center of rotor
4 is swung from the maximum eccentric amount shown in FIG. 2 to the
minimum eccentric amount shown in FIG. 14.
[0165] Second control hydraulic chamber 16b is connected with
discharge port 8 through a connection groove 1g formed in bottom
surface 1a of pump housing 1. Accordingly, the discharge pressure
is acted on second pressure receiving surface 5g of cam ring 5 on
the outer circumference to confront second control hydraulic
chamber 16b, like first pressure receiving surface 5f receiving the
discharge pressure of first control hydraulic chamber 16a.
[0166] Second arc surface 25a has a radius of curvature smaller
than a radius of curvature of first arc surface 5d on first seal
member 14's side. Second pressure receiving surface 5g has a
surface area smaller than a surface area of first pressure
receiving surface 5f. Accordingly, when the discharge pressures of
first and second control hydraulic chambers 16a and 16b are acted,
respectively, on pressure receiving surfaces 5f and 5g, a swing
torque in the counterclockwise direction of FIG. 20 is generated in
cam ring 5, like the first embodiment. However, the hydraulic
torque from second control hydraulic chamber 16b acted only on
second pressure receiving surface 5g is in the clockwise direction.
Accordingly, a part of the hydraulic torque generated in cam ring 5
is canceled. Therefore, the swing torque of cam ring 5 is smaller
than that of cam ring 5 in the first embodiment when the discharge
pressures are identical to each other.
[0167] Accordingly, the spring forces of coil springs 20 and 22 can
be set to the smaller values. Consequently, it is possible to
decrease radii of coil springs 20 and 22, and thereby to decrease
the entire size of the vane pump.
Third Embodiment
[0168] FIGS. 21-29 show a variable displacement pump according to a
third embodiment of the present invention. The variable
displacement pump 01 according to the third embodiment is
substantially identical to the variable displacement pump 01
according to the first and second embodiments. Accordingly,
repetitive illustrations are omitted.
[0169] That is, this variable displacement pump 01 includes a pump
housing 111 which has a U-shaped cross section, and which includes
a pump receiving chamber 113; a cover member 112 which closes an
opening of an one end side of pump housing 111; a drive shaft 114
which penetrates a substantially central portion of pump chamber
113, and which is driven and rotated by the crank shaft of the
engine; a rotor 115 which is rotatably received within pump
receiving chamber 113, and which includes a central portion
connected with drive shaft 114; seven vanes 116 each of which is
moved into and out of one of slots 115a that are formed in an outer
circumference portion of rotor 115, and that extend in the radial
directions; a cam ring 117 which is disposed within pump housing
111, and which is arranged to be swung to be eccentric with respect
to a center of the rotation of rotor 115; a single coil spring 118
which is received within pump housing 111, and which is an urging
member arranged to constantly urge cam ring 117 in a direction to
increase the eccentric amount (eccentricity) of cam ring 117 with
respect to the center of the rotation of rotor 115; and vane rings
119 and 119 slidably disposed on the inner circumference portion of
the both axial side surfaces of rotor 115. A pump constituting
(forming) section is constituted (formed) by drive shaft 114, rotor
115, vanes 116 and cam ring 117.
[0170] As shown in FIGS. 24 and 25, pump housing 111 includes a
bearing hole 111a which is formed at a substantially central
portion of a bottom surface 113a of pump receiving chamber 113,
which rotatably supports one end portion of drive shaft 114, and
which penetrates bottom surface 113a (pump housing 111). Moreover,
as shown in FIG. 25, pump housing 111 includes a support groove
111b which has a semi-circular cross section, which is formed at a
predetermined position of an inner circumference wall of pump
receiving chamber 113 that is an inside surface of pump housing
111, and which swingably supports cam ring 117.
[0171] Moreover, on the inner circumference wall of pump receiving
chamber 113, there are formed first and second sliding surfaces
111c and 111d located on both sides of a cam ring reference line M
connecting a center of bearing hole 111a and a center of support
groove 111b to sandwich cam ring reference line M, and on which
seal members 130 and 130 (described later) disposed on the outer
circumference surface of cam ring 117 are slidably abutted. These
seal sliding surfaces 111c and 111d have, respectively, arc
surfaces which are formed about the center of support groove 111b,
and which have predetermined radii R1 and R2. These seal sliding
surfaces 111c and 111d have, respectively, circumferential lengths
by which seal members 130 and 130 can be constantly slidably
abutted on these seal sliding surfaces 111c and 111d in the
eccentric swing region of cam ring 117. Accordingly, cam ring 117
is slid on and introduced by seal sliding surfaces 111c and 111d
when cam ring 117 is swung to be eccentric. Consequently, it is
possible to obtain smooth activation (eccentric swing movement) of
cam ring 117.
[0172] As shown in FIGS. 22 and 25, there are formed a suction port
121 which is formed on bottom surface 113a of pump receiving
chamber 113 in the outer circumference region of bearing hole 111a,
and which is a suction portion that has a substantially arc
recessed shape, and that is opened in a region (suction region) in
which the inside volumes of pump chambers 120 are increased in
accordance with the pump operation; and a discharge port 122 which
is formed on bottom surface 113a of pump receiving chamber 113 in
the outer circumference region of bearing hole 111a, and which is a
discharge portion that is a substantially arc recessed shape, and
that is opened in a region (discharge region) in which the inside
volumes of pump chambers 120 are decreased in accordance with the
pump operation. Suction port 121 and discharge port 122 are
positioned at positions to sandwich bearing hole 111a, and to
confront each other.
[0173] Suction port 121 is connected with an introduction passage
124 extending from a substantially central position of suction port
121 toward spring receiving chamber 128. In introduction passage
124, there is formed a suction hole 121a which penetrates a bottom
wall of pump housing 111, and which is opened to the outside. With
this, as shown in FIG. 28, the lubricating oil stored in oil pan
152 of the engine is sucked through suction hole 121a and suction
port 121 to pump chambers 120 in the suction region, based on the
negative pressure generated in accordance with the pump operation
of the pump constituting member.
[0174] Suction hole 121a and also introduction passage 124 confront
the outer circumference region of cam ring 117 on the pump suction
side. Suction hole 121a is arranged to introduce the suction
pressure to the outer circumference region of cam ring 117 on the
pump suction side. With this, the pressure of the outer
circumference region of cam ring 117 on the pump suction side which
is adjacent to pump chambers 120 in the suction region becomes the
suction pressure or the atmospheric pressure. Accordingly, it is
possible to suppress the leakage of the lubricating oil from pump
chambers 120 in the suction region to the outer circumference
region of cam ring 117 on the pump suction side. In this case, the
pump suction side is a region on a left side of a cam ring
eccentric direction line N (described later) in FIG. 22.
[0175] Discharge port 122 is connected with an introduction passage
125 extending from a start end portion of discharge port 122 to
confront a first control hydraulic chamber 131 (described later)
defined on the outer circumference side of cam ring 117. At a
terminal end portion of introduction passage 125, there is formed a
discharge hole 122a which penetrates the bottom wall of pump
housing 111, and which is opened to the outside.
[0176] This discharge hole 122a is connected through main oil
gallery 39 to the sliding portions of the engine, the valve timing
control apparatus, and oil jet 30.
[0177] By the thus-constructed configuration, the lubricating oil
which is pressurized by the pump operation of the pump constituting
section, and which is discharged from pump chambers 120 in the
discharge region is supplied through discharge port 122 and
discharge hole 122a to the sliding portions within the engine and
the valve timing control apparatus.
[0178] Discharge hole 122a and also introduction passage 125
confront the outer circumference region of cam ring 117 on the pump
discharge side. Discharge hole 122a is arranged to introduce the
discharge pressure to the outer circumference region of cam ring
117 on the pump discharge side. In this case, the pump discharge
side represents a region on a right side of the cam ring eccentric
direction lien N (described later) in FIG. 22.
[0179] At a portion near the start end portion of discharge port
122, there is formed a connection groove 123 connecting discharge
port 122 and bearing hole 111a. The lubricating oil is supplied
through connection groove 123 to bearing hole 111a. Moreover, the
lubricating oil is supplied to the side portions of rotor 115 and
vanes 116 to ensure the lubricity of the sliding portions.
[0180] This connection groove 123 is formed so as not to correspond
to the direction in which each of the vanes 116 is moved into and
out of one of slots 115a. With this, vanes 116 are suppressed to
drop into connection groove 123 when vanes 116 are moved into and
out of slots 115a.
[0181] Cover member 112 is formed into a substantially plate shape.
Cover member 112 has a portion which is located on the outer side
surface (on the left side in FIG. 24), which corresponds to bearing
hole 111a of pump housing 111, and which has a slightly large
thickness. Cover member 112 includes a bearing hole 112a which is
formed in this large thickness portion, which penetrates the large
thickness portion, and which rotatably supports the other end
portion of drive shaft 114. This cover member 112 has a
substantially flat inner side surface (on the right side in FIG.
24). This cover member 112 is mounted on the open end surface of
pump housing 111 by a plurality of bolts 126.
[0182] Drive shaft 114 is arranged to rotate rotor 115 in the
clockwise direction of FIG. 22 by the rotational force transmitted
from the crank shaft. A left half of FIG. 22 with respect to a line
N (hereinafter, referred to as a cam ring eccentric direction line)
perpendicular to cam ring reference line M at the center of drive
shaft 114 is the pump suction side. A right half of FIG. 22 with
respect to cam ring eccentric direction line N is the pump
discharge side.
[0183] As shown in FIGS. 21 and 22, rotor 115 includes the
plurality of slots 115a formed to extend in the radially outward
directions from the center side; a plurality of back pressure
chambers 115b each of which has a substantially circular cross
section, each of which is formed at a radially inside end (inside
base end) of one of slots 115a, and which is arranged to receive
the discharge fluid discharged to discharge port 122. With this,
vanes 116 are pushed in the radially outward directions by the
centrifugal force caused by the rotation of rotor 115, and the
hydraulic pressures of back pressure chambers 115b.
[0184] Each of vanes 116 includes a tip end portion (radially
outside end) slidably abutted on an inner circumference surface of
cam ring 117, and a base end portion (radially inside end) slidably
abutted on an outer circumference surfaces of vane rings 119 and
119. With this, pump chambers 120 are liquid-tightly separated by
the outer circumference surface of rotor 115, the inner side
surfaces of adjacent two of vanes 116 and 116, the inner
circumference surface of cam ring 117, bottom surface 113a of pump
receiving chamber 113 of pump housing 111, and the inner side
surface of cover member 112, even when the engine speed is low and
the centrifugal force and the hydraulic pressures of back pressure
chambers 115b are small.
[0185] Cam ring 117 is integrally formed into a substantially
cylindrical shape by the sintered metal. Cam ring 117 includes a
pivot portion 117a which is formed into a substantially arc raised
shape, which is mounted in support groove 111b of pump housing 111,
and which serves as an eccentric swing support portion about which
cam ring 117 is swung; and an arm portion 117b which is located at
a position opposite to pivot portion 117a with respect to the
center of cam ring 117, which is liked (connected) with coil spring
118, and which extends in the axial direction to protrude.
[0186] Within pump housing 111, there is formed a spring receiving
chamber 128 located at a position opposite to support groove 111b,
and connected with pump receiving chamber 113 through connection
portion 127 having a predetermined width L. Coil spring 118 is
received in this spring receiving chamber 128.
[0187] This coil spring 118 is resiliently held between the bottom
surface of spring receiving chamber 128 and the lower surface of
the tip end portion of arm portion 117b extending through
connection portion 127 to spring receiving chamber 128. Coil spring
118 has a predetermined set load W. Arm portion 117b includes a
support protrusion 117i which is formed into a substantially arc
shape, which is formed on a lower surface of the tip end portion of
arm portion 117b, and which is engaged with the inner circumference
side of coil spring 118. One end portion of coil spring 118 is
supported by support protrusion 117i.
[0188] Coil spring 118 constantly urges cam ring 117 through arm
portion 117b in a direction in which the eccentric amount of cam
ring 117 is increased (in the clockwise direction of FIG. 22) by
the resilient force based on set load W. With this, cam ring 117 is
brought to a state in which the upper surface of arm portion 117b
is pressed by the urging force of coil spring 118 on a stopper
portion 128a protruding on a cover portion of spring receiving
chamber 128, in the non-activation state of cam ring 117 shown in
FIG. 22. Cam ring 117 is restricted at the position at which the
eccentric amount of cam ring 117 is maximized.
[0189] In this way, arm portion 117b extends in a direction
opposite to pivot portion 117a. The tip end portion of arm portion
117b is urged by coil spring 118. With this, the maximum torque can
be generated in cam ring 117. Accordingly, it is possible to
decrease the size of coil spring 118, and thereby to decrease the
size of the pump itself.
[0190] Moreover, cam ring 117 includes a pair of first and second
seal constituting portions 117c and 117d which are formed on the
outer circumference portion of cam ring 117, which have
substantially triangular cross section, which extend in the axial
direction to protrude, and which include first and second seal
surfaces 117g and 117h that have arc surfaces concentric with seal
sliding surfaces 111c and 111d, and that confront first and second
seal sliding surfaces 111c and 111d. First and second seal surfaces
117g and 117h of first and second seal constituting portions 117c
and 117d include first and second holding grooves 117e and 117f
having a substantially rectangular cross section, and extending in
the axial direction. Seal holding grooves 117e and 117f receive,
respectively, seal members 130 and 130 slidably abutted on seal
sliding surfaces 111c and 111d at the eccentric swing movement of
cam ring 117.
[0191] Seal surfaces 117g and 117h are formed about the center of
pivot portion 117a. Seal surfaces 117g and 117h have predetermined
radii R3 and R4 slightly smaller than radii R1 and R2 of the
corresponding seal sliding surfaces 111c and 111d. Between seal
surfaces 117g and 117h and seal sliding surfaces 111c and 111d,
there are formed, respectively, minute clearances C.
[0192] Seal members 130 and 130 are made of, for example, fluorine
resin with a low frictional characteristic. Each of seal members
130 and 130 has an elongated shape extending linearly in the axial
direction of cam ring 117. Seal members 130 and 130 are pressed on
seal sliding surfaces 111c and 111d by the resilient forces of
resilient members 129 and 129 made from rubber, and disposed on the
bottom portions of seal holding grooves 117e and 117f. With this,
it is possible to constantly ensure the good liquid-tightness of
pressure chambers 131 and 132 described later.
[0193] In the outer circumference region of cam ring 117 on the
pivot portion 117a's side of cam ring eccentric direction line N
which is the pump discharge side (on the right side in FIG. 22) in
the non-activation state of cam ring 117, there are separated a
first control hydraulic chamber 131 and a second control hydraulic
chamber 132 defined on both sides of pivot portion 117a to sandwich
pivot portion 117a, by the outer circumference surface of cam ring
117, pivot portion 117a, seal members 130 and 130, and the inside
surface of pump housing 111.
[0194] In this example, the entire of first and second control
hydraulic chambers 131 and 132 are set within the region on the
pump discharge side, on the outer circumference region of cam ring
117. It is preferable to set the entire of first and second control
hydraulic chambers 131 and 132 in a region overlapped with the
discharge region which is the pressurized region in the radial
direction, that is, a region which confront pump chambers 120 that
become constantly the positive pressure to sandwich the
circumference wall of cam ring 117.
[0195] The discharge pressure discharged to discharge port 122 is
constantly introduced through introduction passage 125 to first
control hydraulic chamber 131. The discharge pressure is acted to a
first pressure receiving surface 133 which is constituted by the
outer circumference surface of cam ring 117 that confronts first
control hydraulic chamber 131, and which receives the force acted
to counteract (block) the urging force of coil spring 118. With
this, cam ring 117 receives the swing force (movement force) in a
direction (the counterclockwise direction) to decrease the
eccentric amount of cam ring 117.
[0196] That is, this first control hydraulic chamber 131 is
arranged to urge cam ring 117 through first pressure receiving
surface 133 in a direction in which the center of cam ring 117
approaches (become concentric to) the center of the rotation of
rotor 115. With this, first control hydraulic chamber 131 serves
for the movement amount control in the concentric direction of cam
ring 117.
[0197] On the other hand, second control hydraulic chamber 132 is
arranged to receive the discharge pressure through an introduction
hole 135 which penetrates the bottom wall of pump housing 111, and
which is connected with discharge hole 122a through a solenoid
valve 140 (described later) which is controlled in accordance with
the driving state of the engine. Accordingly, the discharge
pressure is acted to second pressure receiving surface 134 which is
constituted by the outer circumference surface of cam ring 117 that
confronts second hydraulic chamber 132, and which receives the
force acted in a direction to assist the urging force of coil
spring 118. With this, cam ring 117 receives the swing force in a
direction (in the clockwise direction of FIG. 22) in which the
eccentric amount of cam ring 117 is increased.
[0198] As shown in FIG. 22, second pressure receiving surface 134
has a pressure receiving area S2 smaller than a pressure receiving
area S1 of first pressure receiving area 133. The urging force in
the eccentric direction of cam ring 117 by the urging force based
on the internal pressure of second control hydraulic chamber 132
and the urging force of coil spring 118, and the urging force by
first control hydraulic chamber 131 are balanced to keep a
predetermined force relationship. The urging force by second
control hydraulic chamber 132 assists the urging force of coil
spring 118.
[0199] Second control hydraulic chamber 132 is arranged to act the
discharge pressure supplied through solenoid valve 140 to second
pressure receiving surface 134, and thereby to assist the urging
force of coil spring 118. With this, second control hydraulic
chamber 132 serves for the movement amount control of cam ring 117
in the eccentric direction.
[0200] As shown in FIG. 28, this variable displacement pump 01 is
provided with solenoid valve 140 which is a component different
from variable displacement pump 01, and which is operated in
accordance with the driving state of the engine based on the
excitation current from ECU 151 mounted on the vehicle. Discharge
hole 122a and introduction hole 135 are connected with each other
through this solenoid valve 140. With this, first control hydraulic
chamber 131 and second control hydraulic chamber 132 are connected
with each other in the open state of solenoid valve 140.
[0201] As shown in FIGS. 26 and 27, this solenoid valve 140
includes a valve body 141 which has a cylindrical shape, and which
includes an one end (on the right side in FIGS. 26 and 27) opened,
and the other end (on the left side in FIGS. 26 and 27) closed; a
valve element 142 which is disposed within valve body 141 to be
slid in the axial direction, and which includes first and second
land portions 142a and 142b located on both side end portions of
valve element 142, and arranged to be slid on an inner
circumference surface of valve body 141; a spring 143 which is
received in a back pressure chamber 145 separated on the other end
side of valve body 141 by second land portion 142b of valve element
142, and which is arranged to urge valve element 142 toward the one
end side of valve body 141; an electromagnetic unit 144 mounted on
the opening of valve body 143, and which is arranged to move a rod
144b in accordance with the energization, and thereby to move valve
element 142 toward the other end side of valve body 141 in the
axial direction against the urging force of spring 143.
[0202] Valve body 141 includes an IN port 141a which is formed in
the circumference wall, which penetrates the circumference wall,
and which is connected with discharge hole 122a; an OUT port 141b
which is formed in the circumference wall, which penetrates the
circumference wall, and which is connected with introduction hole
135; a drain port 141c which is formed in the circumference wall,
which penetrates the circumference wall, and which is connected
with suction port 121 or the outside. Moreover, valve body 141
includes a back pressure port 141d which is formed in a side wall
of the other end, which penetrates the side wall of the other end,
which is connected with suction port 121 or the outside, and which
is constantly opened to back pressure chambers 145.
[0203] Valve element 142 includes a substantially central portion
in the axial direction which has a smaller diameter, and land
portions 142a and 142b which define an annular space 146 between
the central portion of valve element 142 and valve body 141. OUT
port 141b, and IN port 141a or drain port 141c are connected
through annular space 146.
[0204] Electromagnetic unit 144 has a conventional structure.
Electromagnetic unit 144 includes a coil unit 144a having a bobbin
around which a coil is wound, and a yoke mounted on the
thus-constructed bobbin; an armature (not shown) which is made of a
magnetic material, which is disposed radially inside coil unit
144a, and which is moved into and out of coil unit 144a in the
axial direction; and a rod 144b which is connected with the
armature, and which is moved into and out of (in the forward and
rearward directions) with the armature in accordance with the
energization state.
[0205] As shown in FIG. 26, solenoid valve 140 is a normally open
type. When the excitation current is applied to coil unit 144a, IN
port 141a and OUT port 141b are connected with each other through
annular space 146. With this, the discharge pressure is introduced
into second control hydraulic chamber 132. In this case, drain port
141c is opened to back pressure chamber 145.
[0206] On the other hand, when the excitation current is applied to
coil unit 144a, valve element 142 is pressed and returned toward
the other end side of valve body 141 against the urging force of
spring 143 by the pressing force of rod 144b, as shown in FIG. 27.
With this, IN port 141a is closed by first land portion 142a of
valve element 142. OUT port 141b is connected through annular space
146 with drain port 141c. With this, second control hydraulic
chamber 132 is opened to the suction pressure or the atmospheric
pressure.
[0207] By the thus-constructed structure, in variable displacement
pump 01, a relationship of relative forces acted to cam ring 117
between the internal pressure of first control hydraulic chamber
131, the urging force of coil spring 118, and the internal pressure
of second control hydraulic chamber 132 that is controlled by
solenoid valve 140 is controlled so as to control the eccentric
amount of cam ring 117. With this, the variation of the inside
volumes of pump chambers 120 are controlled at the pump operation
by controlling the eccentric amount, so that the discharge pressure
characteristic of variable displacement pump 01 is controlled.
[0208] Hereinafter, the operation of variable displacement pump 01
according to the third embodiment, that is, the discharge pressure
control of the pump based on the eccentric amount control of cam
ring 117 is illustrated with reference to FIGS. 22, 23 and 29.
[0209] When the valve timing control apparatus is activated, the
requested (necessary) pressure of the discharge pressure of
variable displacement pump 01 becomes hydraulic pressure P1 in FIG.
29, as described above. That is, the valve timing control apparatus
is set to activate (start) by the low hydraulic pressure P1
immediately after the start of the engine.
[0210] The requested hydraulic pressure of the crank metal at the
high engine speed is a hydraulic pressure P2 in FIG. 29 at the low
load or the low hydraulic oil temperature. On the other hand, the
requested hydraulic pressure of the crank metal at the high engine
speed is a hydraulic pressure P4 in FIG. 29 at the high load or the
high hydraulic oil temperature.
[0211] Moreover, at the high load of the engine, oil jet 30 is used
for cooling the piston. The valve opening pressure of ball valve
element 46 of this oil jet 30 is set to a hydraulic pressure P3 in
FIG. 29 at a predetermined engine speed n which is the middle
engine speed.
[0212] At the low load or the low hydraulic oil temperature,
variable displacement pump 01 is set to a low pressure
characteristic X which is a first discharge pressure characteristic
that satisfies one of hydraulic pressure P1 and hydraulic pressure
P2 of FIG. 29, or both of the hydraulic pressure P1 and hydraulic
pressure P2. At the high load or the high hydraulic oil
temperature, variable displacement pump 01 is set to a high
pressure characteristic Y which is a second discharge pressure
characteristic that satisfies one of the hydraulic pressure P3 and
the hydraulic pressure P4, or both of hydraulic pressure P3 and
hydraulic pressure P4.
[0213] The operation characteristic of cam ring 117, that is, first
and second operation hydraulic pressures Px and Py which are needed
for the activation of cam ring 117 are varied by switching of
ON/OFF states of solenoid valve 140. The appropriate hydraulic
pressure characteristic is selected from both hydraulic pressure
characteristics X and Y in accordance with the driving state of the
engine, so as to satisfy the requested hydraulic pressure of the
engine.
[0214] In this example, as shown in FIG. 29, low pressure
characteristic X is set to the hydraulic pressure characteristic
shown by a broken line connecting requested hydraulic pressure P1
of the valve timing control apparatus and requested hydraulic
pressure P2 at the high engine speed at the low load or the low
hydraulic fluid temperature. On the other hand, high pressure
characteristic Y is set to the hydraulic pressure characteristic
shown by a solid line connecting requested hydraulic pressure P3
which is the valve opening pressure of oil jet 30 at the middle
engine speed at the high load or the high hydraulic fluid
temperature, and requested hydraulic pressure P4 at the high engine
speed at the high load or the high hydraulic fluid temperature.
[0215] That is, in variable displacement pump 01, spring load W of
coil spring 118 is set to first activation hydraulic pressure Px.
IN port 141a is closed (shut off) by applying the excitation
current from ECU 151 to solenoid valve 140 at the low load or the
low oil temperature. With this, the discharge pressure is
introduced only to first control hydraulic chamber 131.
[0216] Accordingly, the eccentric amount of cam ring 117 is held in
the maximum state until the internal pressure of the first control
hydraulic chamber 131 reaches first activation hydraulic pressure
Px (cf. FIG. 22). The discharge pressure is suddenly increased in
accordance with the increase of the engine speed.
[0217] When the internal pressure of first control hydraulic
chamber 131 reaches first activation hydraulic pressure Px by the
increase of the discharge pressure, cam ring 117 is swung about
pivot portion 117a in the downward direction of the cam ring
eccentric direction line N, that is, in a direction in which the
eccentric amount is decreased (cf. FIG. 23). Consequently, the
variations of the volumes of pump chambers 120 are decreased at the
operation of the pump. Therefore, the increase of the discharge
pressure according to the increase of the engine speed becomes
gentle, so that low pressure characteristic X shown in FIG. 29 can
be obtained.
[0218] On the other hand, when it is shifted from the low load or
the low oil temperature state to the high load or the high oil
temperature state, the excitation current from ECU 151 to solenoid
valve 140 is shut off. With this, IN port 141a and OUT port 141b
are connected with each other. Accordingly, the discharge pressure
is introduced to first control hydraulic chamber 131 and also
second control hydraulic chamber 132.
[0219] The pressure acted on second pressure receiving surface 134
of second control hydraulic chamber 132 serves to assist the urging
force of coil spring 118. Accordingly, cam ring 117 is not
activated even when the internal pressure of first control
hydraulic chamber 131 reaches first activation hydraulic pressure
Px of FIG. 29. Cam ring 117 is held in a state in which the
eccentric amount of cam ring 117 is maximized, until the difference
of the hydraulic pressures which are acted on first pressure
receiving surface 133 and second pressure receiving surface 134 by
the internal pressure of first control hydraulic chamber 131 and
the internal pressure of second control hydraulic chamber 132
reaches the urging force of coil spring 118 (cf. FIG. 22).
[0220] As shown in FIG. 29, the eccentric amount of cam ring 117 is
held in the maximum state at the high load or the high oil
temperature state, until the discharge pressure reaches second
activation hydraulic pressure Py so that the difference of the
hydraulic pressures acted on first pressure receiving surface 133
and second pressure receiving surfaces 134 by the internal pressure
of first control hydraulic chamber 131 and the internal pressure of
second control hydraulic chamber 132 becomes equal to the urging
force of coil spring 118. Accordingly, the discharge pressure is
largely increased in accordance with the increase of the engine
speed.
[0221] Then, cam ring 117 is swung in a direction in which the
eccentric amount is decreased when the internal pressure of first
control hydraulic chamber 131 reaches second activation hydraulic
pressure Py (cf. FIG. 23). With this, the variations of the volumes
of pump chambers 120 become small at the pump operation, and the
increase of the discharge pressure according to the increase of the
engine speed becomes gentle. Consequently, high pressure
characteristic Y can be obtained as shown in FIG. 29.
[0222] In variable displacement pump 01, in principle, the pump
discharge pressure characteristic is shifted to high pressure
characteristic Y when ECU 151 judges that the high pressure is
needed by the engine speed, the load, the oil temperature and so
on.
[0223] Normally, the pump discharge pressure characteristic is
shifted to high pressure characteristic Y when the load of the
engine, the oil temperature and so on are high. In the
above-described illustration, the high pressure characteristic Y is
used at the high load of the engine or the high oil temperature.
However, for example, the valve timing control apparatus may need
the hydraulic pressure higher than requested hydraulic pressure P1.
In this case, ECU 151 switches solenoid valve 140 in accordance
with the activation signal of the valve timing control apparatus.
Even at the low load of the engine, the low oil temperature or so
on, the pump discharge pressure characteristic is shifted to high
pressure characteristic Y.
[0224] That is, in this example, requested hydraulic pressure P1 is
set to the normal requested hydraulic pressure of the valve timing
control apparatus. Requested hydraulic pressure P1 is set to
minimum requested hydraulic pressure in the valve timing control
apparatus, in accordance with the specification and so on of the
vehicle employing this valve timing control apparatus.
[0225] Moreover, when it is shifted again from the high load or the
high oil temperature state to the low load or the low oil
temperature state, ECU 151 applies the excitation current to
solenoid valve 140 again, so that solenoid valve 140 becomes the
energization state shown in FIG. 27. With this, second control
hydraulic chamber 132 is opened to the atmospheric pressure or the
suction pressure. Accordingly, the activation of cam ring 117
depends on the relationship of the force between the internal
pressure of first control hydraulic chamber 131 and the urging
force of coil spring 118. With this, the discharge pressure
characteristic of the pump is varied to low pressure characteristic
X. Consequently, it is possible to decrease the discharge pressure
which is not needed for shifting to the low load or the low oil
temperature state, and to suppress the power loss of the
engine.
[0226] In this variable displacement pump 01, ECU 151 switches
solenoid valve 140 in accordance with the driving information such
as the engine speed, the load of the engine, and the oil
temperature, so that the activation characteristic of cam ring 117
is varied. Accordingly, it is possible to select the discharge
pressure characteristic suitable for the engine speed, the load of
the engine, the oil temperature and so on. With this, it is
possible to cut waste of the work of the pump, and to suppress
(minimize) the power loss of the engine.
[0227] Moreover, in this variable displacement pump 01, the
operation control of cam ring 117 does not need the complex control
such as the duty control. Furthermore, variable displacement pump
01 does not need a high-precision work (processing) of the shape of
the port, the tuning of the valve opening characteristic and so on
of solenoid valve 140. Accordingly, it is possible to readily
attain the operation control of cam ring 117 by the simple control
by the switching of ON/OFF of solenoid valve 140, and by the simple
structure using the general solenoid valve 140. Therefore, it is
possible to decrease the manufacturing cost of the pump.
[0228] In variable displacement pump 01, the internal pressures of
pump chambers 120 in the discharge region are acted on the inner
circumference surface of cam ring 117 on pivot portion 117a's side,
as shown by a bold solid arrow of FIG. 23. Accordingly, cam ring
117 is pressed in the right direction of FIG. 23 along the cam ring
reference line M, that is, toward support groove 111b, so that
pivot portion 117a is pressed against support groove 111b.
[0229] However, in variable displacement pump 01 of this example,
control hydraulic chambers 131 and 132 are disposed radially
outside cam ring 117 on the pump discharge side, that is, so as to
confront these pump chambers 120 in the discharge region to
sandwich the circumference wall of cam ring 117. As shown by bold
broken arrows in FIG. 23, the internal pressures of both of control
hydraulic chambers 131 and 132 are acted, respectively, on cam ring
117 so as to push cam ring 117 in a direction opposite to support
groove 111b. Consequently, it is possible to decrease the tight
abutment (engagement) of pivot portion 117a against support groove
111b. Therefore, it is possible to decrease the friction between
pivot portion 117b and support groove 111b at the eccentric swing
movement of cam ring 117.
[0230] Accordingly, it is possible to suppress the abrasion of
pivot portion 117a and support groove 111b, in particular, the
abrasion of support groove 111b made of a material having a low
rigidity relative to cam ring 117. Therefore, it is possible to
improve the endurance of the pump.
[0231] By this function, the forces acted on the inner and outer
circumference sides of cam ring 117 on the pump discharge side are
canceled. On the other hand, the atmospheric pressure or the
suction pressure is acted through introduction passage 124 on the
outer circumference region of cam ring 117 on the pump suction side
which is opposite to support groove 111b. Pivot portion 117a is
pressed into support groove 111b by the atmospheric pressure or the
suction pressure. Accordingly, pivot portion 117a is not apart from
(disengaged from) the inside surface of support groove 111b. With
this, pivot portion 117a is appropriately abutted and slid on
support groove 111b, and it is possible to obtain an appropriate
activation of cam ring 117.
[0232] Moreover, the both of pressure chambers 131 and 132 are
disposed in the region on the pump discharge side, so as to
confront pump chambers 120 in the discharge region, as described
above. With this, in this region, the pressure acted on the inner
circumference side and the pressure acted on the outer
circumference side of cam ring 117 become the discharge pressure.
Accordingly, the pressure acted on the inner circumference side of
cam ring 117 is substantially identical to the pressure acted on
the outer circumference side of cam ring 117. Therefore, it is
possible to suppress (minimize) the pressure difference between the
inner circumference and the outer circumference of cam ring 117 in
the discharge region. Consequently, it is possible to suppress
(minimize) the leakage of the lubricating oil in the discharge
region through the minute clearances between the both side surfaces
of cam ring 117, bottom wall 113a of pump receiving chamber 113,
and the inner side surface of cover member 112. Therefore, it is
possible to cut the waste of the work of variable displacement pump
01, and to improve the efficiency of variable displacement pump
01.
[0233] As described above, in variable displacement pump 01, first
and second pressure chambers 131 and 132 are disposed on the both
sides of pivot portion 117a to sandwich pivot portion 117a. With
this, the internal pressure of second control hydraulic chamber 132
serves to assist the urging force of coil spring 118. Accordingly,
it is possible to set the urging force of coil spring 118 to a
small value.
[0234] That is, by the disposition of second control hydraulic
chamber 132, coil spring 118 only needs to have the urging force
which ensures low pressure characteristic X, and which balances
with first activation hydraulic pressure Px. Accordingly, it is
possible to use the coil spring with the low load which has a
spring constant smaller than that of the conventional coil spring.
Consequently, it is possible to decrease the space for the
disposition of coil spring 118 in pump housing 111, and to decrease
the size and the weight of variable displacement pump 01.
Therefore, it is possible to improve the mounting characteristic of
variable displacement pump 01 on the engine.
[0235] Moreover, second pressure receiving surface 134 has the
pressure receiving area smaller than the pressure receiving area of
first pressure receiving surface 133. The activation hydraulic
pressure of cam ring 117 is set to the two stages (steps) by second
control hydraulic chamber 132. With this, it is possible to improve
degree of freedom of the discharge pressure characteristic of the
pump.
[0236] Furthermore, the operation of the valve timing control
apparatus and the lock release hydraulic pressure of the lock
mechanism are set to hydraulic pressure P1 in low pressure
characteristic X. Accordingly, it is possible to improve the
responsiveness of the operation of the valve timing control
apparatus, like the first and second embodiments.
[0237] Moreover, first discharge pressure X is set smaller than
valve opening pressure P3 of oil jet 30. Accordingly, oil jet 30
does not inject the hydraulic fluid in the engine speed which is
used in the normal running of the vehicle.
[0238] Therefore, like the first and second embodiments, it is
possible to suppress the discharge amount of variable displacement
pump 01, to decrease the friction of various portions, and to
improve the fuel consumption.
[0239] Moreover, oil jet 30 does not inject the oil of the low
temperature in the cold state of the engine. Accordingly, it is
possible to improve the warm-up characteristic.
[0240] Heretofore, there are proposed various pump such as the
variable displacement pump for the power steering apparatus, which
is arranged to control and swing the cam ring by the pressure
difference between the two pressure chambers. These conventional
pump is arranged to generate the pressure difference based on the
pressure loss by the orifice and so on. This pressure loss
decreases the pump efficiency. On the other hand, in variable
displacement pump 01 according to the third embodiment, the
discharge pressure is introduced into first control hydraulic
chamber 131 and second control hydraulic chamber 132 without the
pressure loss. Variable displacement pump 01 is arranged to
generate the activation torque of cam ring 117 by the difference of
the areas of the pressure receiving surfaces of first control
hydraulic chamber 131 and second control hydraulic chamber 132,
that is, by the difference between the areas of first pressure
receiving surface 133 and second pressure receiving surface 134.
Accordingly, in variable displacement pump 01, the decrease of the
pump efficiency is not generated, unlike the conventional variable
displacement pump. Therefore, it is possible to improve the pump
efficiency relative to the conventional variable displacement pump
since the pressure loss is not generated.
[0241] Moreover, variable displacement pump 01 is set to the high
pressure characteristic when solenoid valve 140 is not energized.
Variable displacement pump 01 has a fail-safe function to ensure
necessary discharge pressure in the entire region which is used by
the engine when solenoid valve 140 is in the failure state.
[0242] FIGS. 30 and 31 show variation of the third embodiment. In
this variation of the third embodiment, solenoid valve 140 is a
normally-closed type, unlike solenoid valve 140 of the third
embodiment.
[0243] That is, solenoid valve 140 according to this variation of
the third embodiment is the normally-closed type which is opposite
to solenoid valve 140 of the third embodiment. As shown in FIG. 30,
in the non-energization state, IN port 141a is shut off, and OUT
port 141b is connected with drain portion 141c. As shown in FIG.
31, in the energization state, IN port 141a is connected with OUT
port 141b. With this, variable displacement pump 01 has low
pressure characteristic X when solenoid valve 140 is not energized.
Variable displacement pump 01 has high pressure characteristic Y
when solenoid valve 140 is energized.
[0244] By the thus-constructed configuration, when the frequency of
high pressure characteristic Y is less than the frequency of low
pressure characteristic X, it is possible to decrease the time
duration of the energization to solenoid valve 140, and thereby to
suppress the time degradation of solenoid valve 140.
Fourth Embodiment
[0245] FIGS. 32-36 show a variable displacement pump according to a
fourth embodiment of the present invention. In this fourth
embodiment, the dispositions of seal members 130 and 130 are
varied, and solenoid valve 140 is integrally formed with the
housing, unlike the third embodiment.
[0246] In this fourth embodiment, seal holding grooves 117e and
117f formed in seal constituting sections 117c and 117d of cam ring
117 are omitted, unlike the third embodiment. Alternatively, seal
sliding surface 111c and 111d include seal holding grooves 111e and
111f which are identical to seal holding grooves 117e and 117f of
the third embodiment, and which are formed, respectively, at
positions to confront the omitted seal holding grooves 117e and
117f of the third embodiment. Seal holding grooves 111e and 111f
receive, respectively, resilient members 129 and 129 and seal
members 130 and 130.
[0247] Moreover, in this fourth embodiment, valve body 141 of
solenoid valve 140 is integrally formed in outside surface 112b of
cover member 112 to be substantially parallel with cam ring
eccentric direction line N, as shown in FIGS. 32, 35 and 36.
Solenoid valve 140 and the housing are integrally formed.
[0248] Solenoid valve 140 has a structure identical to that of the
solenoid valve of the third embodiment. Valve element 142 is
slidably received within valve body 141 integrally formed with
cover member 112. An electromagnetic unit 144 is mounted in an
opening portion of an one end portion which is an upper end portion
of valve body 141 in FIG. 35.
[0249] By the change of these structure, on the inner side surface
112c of cover member 112, there are formed suction port 121,
discharge port 122, connection groove 123 connecting discharge port
122 and bearing hole 112a, and introduction passage 125 extending
from discharge port 122, like pump housing 111, as shown in FIG.
36.
[0250] This cover member 112 includes an IN port 141a which is
formed at a predetermined position of introduction passage 125, and
which connects the inside (pump receiving chamber 113) of pump
housing 111 and the inside of valve body 141; and an OUT port 141b
which is formed at a predetermined position substantially
symmetrical to IN port 141a with respect to cam ring reference line
M, and which serves as introduction hole 135. Valve body 111
integrally formed with cover member 112 includes a drain port 141c
and a back pressure port 141d formed at predetermined positions on
the circumference wall and the bottom wall of valve body 111.
[0251] Accordingly, at the swing eccentric movement of cam ring
117, seal members 130 and 130 are slidably abutted on seal surfaces
117g and 117h of cam ring 117 which is made of sintered material of
iron, and which has a hardness higher than a hardness of pump
housing 111 made of aluminum alloy. Therefore, it is possible to
suppress the abrasion of counterpart by seal members 130 and 130.
Consequently, it is possible to improve the endurance (durability)
of variable displacement pump 01, relative to the third
embodiment.
[0252] Moreover, in this fourth embodiment, solenoid valve 140 is
integrally formed with cover member 112, that is, the housing. The
entire of the hydraulic circuit of variable displacement pump 01 is
formed in variable displacement pump 01. Accordingly, it is
possible to decrease the size of the hydraulic pressure supply
system including variable displacement pump 01 as a main
device.
Fifth Embodiment
[0253] FIGS. 37-39 show a fifth embodiment of the present
invention. The fifth embodiment has a basic structure identical to
that of the fourth embodiment. In this fifth embodiment, a
hydraulic directional control valve 150 activated by the discharge
pressure is arranged to vary the discharge pressure characteristic
of the pump, in place of solenoid valve 140 according to the fourth
embodiment.
[0254] In this fifth embodiment, known hydraulic directional
switching valve 150 of a spool type is used in place of solenoid
valve 140. As shown in FIGS. 37 and 38, this hydraulic directional
control valve 150 includes a valve body 151 which is formed into a
substantially cylindrical shape, and which has one end opened, and
the other end closed; a plug 152 which closes the opening of the
one end of valve body 151; a valve element 153 which is received
within valve body 151, which is arranged to be slid within valve
body 151 in the axial direction, and which includes first and
second land portions 153a and 153b that are located in the both end
portions of valve element 153, and that define a pressure chamber
155 and a back pressure chamber 156 within valve body 151; a spring
154 which is received in back pressure chamber 156, and which is
arranged to urge valve element 153 toward pressure chamber 155.
When an internal pressure of pressure chamber 155 exceeds a
predetermined set pressure Pz which is larger than required
hydraulic pressure P1, and smaller than required hydraulic pressure
P2, valve element 153 is moved toward back pressure chamber 156
against the urging force of spring 154, as shown in FIG. 38.
[0255] Valve body 151 includes an IN port 151a which is formed in
the circumference wall of valve body 151 at a predetermined axial
position, which penetrates the circumference wall of valve body
151, and which is connected with discharge hole 122a; an OUT port
151b which is formed in the circumference wall of valve body 151 at
a predetermined axial position, which penetrates the circumference
wall of valve body 151, and which is connected with introduction
hole 135; and a drain port 151c which is formed in the
circumference wall of valve body 151 at a predetermined axial
position, which penetrates the circumference wall of valve body
151, and which is connected with suction port 121 or the outside.
Moreover, valve body 151 includes a back pressure port 151d which
is formed in a side wall of valve body 151 that is on the back
pressure chamber 155's side, which penetrates the side wall of
valve body 151, and which is connected with suction port 121 or the
outside to constantly open (connect) back pressure chamber 145 to
the suction pressure or the atmospheric pressure.
[0256] Plug 152 is screwed in an internal thread portion formed on
an inner circumference surface of the opening of the one end
portion of valve body 151. Plug 152 includes an introduction port
152a extending in the axial direction, and penetrating plug 152.
The discharge pressure is constantly introduced through
introduction port 152a into pressure chamber 155.
[0257] Valve element 153 includes an axially central portion which
is located at a central portion of valve element 153 in the axial
direction, and which has a smaller diameter; and land portions 153a
and 153b which are located on both sides of the central portion,
and which define an annular space 157 with valve body 151. OUT port
151b, and IN port 151a or drain port 151c are connected with each
other through annular space 157.
[0258] That is, in the non-activation state of valve element 153,
first land portion 153a closes IN port 151a, and OUT port 151b and
drain port 151c are connected with each other through annular space
157. In the activation state of valve element 153, second land
portion 153b closes drain port 151c, and IN port 151a and OUT port
151b are connected with each other through annular space 157.
[0259] Accordingly, in variable displacement pump 01 of the fifth
embodiment, when the engine speed is low, IN port 151a of hydraulic
directional control valve 150 is closed, so that the discharge
pressure is acted only to first control hydraulic chamber 131.
Consequently, when the discharge pressure reaches first activation
hydraulic pressure Px as shown in FIG. 40, cam ring 117 is swung in
a direction in which the eccentric amount is decreased to obtain
low pressure characteristic X in which the increase of the
discharge pressure becomes gradual (region T1 in FIG. 40).
[0260] Then, when the internal pressure of pressure chamber 155
reaches set pressure Pz by the increase of the discharge pressure,
valve element 153 starts to be moved by the internal pressure of
pressure chamber 155 in the axial direction toward back pressure
chamber 156 against the urging force of spring 153. Accordingly,
drain port 151c is closed by second land portion 153b, and IN port
151a is gradually opened to annular space 157. With this, IN port
151a and OUT port 151b are gradually connected with each other
through annular space 157, so that the discharge pressure is
gradually introduced into second control hydraulic chamber 132.
Consequently, the internal pressure of second control hydraulic
chamber 132 is increased, and cam ring 117 is moved in a direction
in which the eccentric amount of cam ring 117 is increased.
Therefore, high pressure characteristic Y to further increase the
discharge pressure is attained (region T2 in FIG. 40).
[0261] In this way, by the fifth embodiment, it is possible to
obtain an oil pump having a discharge pressure characteristic
corresponding to the engine speed, by a lower manufacturing
cost.
[0262] Moreover, the activation pressure of the valve timing
control apparatus is set to the hydraulic pressure P1 in low
pressure characteristic X. The valve opening pressure of oil jet 30
is set to the hydraulic pressure P3 in the hydraulic pressure
characteristic Y. In this case, first activation hydraulic pressure
Px is set to a hydraulic pressure sufficiently smaller than the
hydraulic pressure P3. Accordingly, it is possible to decrease the
consumption energy, like the first to fourth embodiments.
[0263] Moreover, in the above-described embodiments, the operation
of cam ring 117 is controlled by balancing the urging force of coil
spring 118 and the internal pressure of second control hydraulic
chamber 132 with respect to the internal pressure of first control
hydraulic chamber 131. The pressure receiving area of first
pressure receiving surface 133 may be set larger than the pressure
receiving area of second pressure receiving surface 134 in
accordance with the specifications of the pump, and thereby coil
spring 118 may be omitted. With this, the activation of cam ring
117 may be controlled only by the internal pressures (pressure
difference) of pressure chambers 131 and 132.
[0264] Moreover, in the above-described embodiments, the pressure
receiving area of second pressure receiving surface 134 is set
smaller than the pressure receiving area of first pressure
receiving surface 133. However, the pressure receiving area of
second pressure receiving surface 134 may be set equal to the
pressure receiving area of first pressure receiving surface 133 in
accordance with the requirement of the internal combustion
engine.
[0265] Moreover, the seal members are disposed to ensure the
sealing ability of the control hydraulic chamber. The seal members
may be omitted for the cost saving as long as the requested
hydraulic pressure characteristic of the internal combustion engine
is satisfied.
[0266] Moreover, the disposition of the spring receiving chamber
may be varied. Set loads of the coil springs may be varied in
accordance with the specifications and the size of the pump.
Furthermore, the coil diameters and lengths of the coil springs may
be varied.
[0267] The variable valve actuating apparatus is not limited to the
valve timing control apparatus. Moreover, the present invention is
applicable to, for example, a lift varying mechanism to vary a
working angle (operation angle) and a lift amount of valve of the
engine.
[0268] Moreover, this variable displacement pump is applicable to
hydraulic equipments and so on which are other than the internal
combustion engine.
[0269] In the present invention, a variable displacement pump
arranged to supply a hydraulic fluid to an oil jet arranged to
inject the hydraulic fluid to a piston of an internal combustion
engine when a pressure of the supplied hydraulic fluid becomes
equal to or greater than a predetermined pressure, the variable
displacement pump includes: a pump constituting section arranged to
be driven and rotated by the internal combustion engine, and
thereby to discharge the hydraulic fluid entered from a suction
portion to a plurality of operation chambers, from a discharge
portion by volume variations of the operation chambers; a movable
member arranged to decrease the flow rate of the hydraulic fluid
discharged from the discharge portion by moving in one direction; a
control section configured to move the movable member in the one
direction by a predetermined amount when the discharge pressure of
the hydraulic fluid becomes a first discharge pressure, and to
further move the movable member in the one direction when the
discharge pressure of the hydraulic fluid becomes a second
discharge pressure larger than the first discharge pressure, the
first discharge pressure being set smaller than the predetermined
pressure at which the oil jet starts to inject the hydraulic
fluid.
[0270] Accordingly, it is possible to suppress the energy
consumption at the initial stage of the discharge of the hydraulic
fluid.
[0271] In the variable displacement pump according to the present
invention, the second discharge pressure is set larger than the
predetermined pressure at which the oil jet starts to inject the
hydraulic fluid.
[0272] Accordingly, the second discharge pressure is set larger
than the predetermined pressure at which the oil jet starts to
inject the hydraulic fluid. Therefore, it is possible to ensure the
injection from the oil jet to the piston without being influenced
by the increase of the oil temperature and the variation of the
cooling state of the internal combustion engine.
[0273] In the variable displacement pump according to the present
invention, the oil jet includes; a body including a hydraulic fluid
supplying portion to which the hydraulic fluid is supplied, a
hydraulic fluid introducing portion arranged to introduce the
hydraulic fluid supplied to the hydraulic fluid supplying portion,
and a valve seat formed between the hydraulic fluid supplying
portion and the hydraulic fluid introducing portion; a valve
element arranged to be seated on and released from the valve seat
in accordance with the pressure of the hydraulic fluid supplied to
the hydraulic fluid supplying portion, and thereby to open and
close the hydraulic fluid supplying portion; an urging member
arranged to urge the valve element in a valve closing direction in
which the valve element closes the hydraulic fluid supplying
portion, and to set a valve opening pressure of the valve element
at which the valve element opens the hydraulic fluid supplying
portion, to a value larger than the first discharge pressure; and
an injection nozzle connected on a downstream side of the hydraulic
pressure introducing portion, and arranged to inject the hydraulic
fluid from an injection opening toward the piston.
[0274] In the variable displacement pump according to the present
invention, the movable member is a cam ring having a cam surface
formed on an inner circumference surface thereof; the pump forming
section includes a rotor arranged to be driven and rotated by the
internal combustion engine, and vanes disposed on an outer
circumference portion of the rotor, and arranged to be moved in a
radially inward direction or in a radially outward direction, and
to be moved in the radially outward direction toward the inner
circumference surface to separate the plurality of the operation
chambers; and the cam ring is arranged to move to vary an eccentric
amount of the cam ring with respect to a center of the rotor.
[0275] In the variable displacement pump according to the present
invention, the discharged hydraulic fluid lubricates sliding
portions of the internal combustion engine.
[0276] In the variable displacement pump according to the present
invention, the discharged hydraulic fluid activates a valve timing
control apparatus arranged to vary a relative rotational phase
between a driving rotational member and a cam shaft of the internal
combustion engine, and a lock mechanism of the valve timing control
apparatus; and the lock mechanism has a release pressure at which a
lock of the lock mechanism is released, and which is set smaller
than the first discharge pressure.
[0277] The entire contents of Japanese Patent Application No.
2010-26335 filed Feb. 9, 2010 are incorporated herein by
reference.
[0278] 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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