VALVE TIMING CONTROL APPARATUS OF INTERNAL COMBUSTION ENGINE

TECHNICAL FIELD

The present invention relates to a valve timing control apparatus of an internal combustion engine for variably controlling valve timing of an engine valve, such as an intake valve and/or an exhaust valve, depending on an engine operating condition.

BACKGROUND ART

In recent years, there have been proposed and developed various hydraulically-operated vane rotor equipped variable valve timing control (VTC) devices, capable of locking a vane rotor at a predetermined intermediate-phase angular position (simply, an intermediate position) between a maximum phase-advance position and a maximum phase-retard position by means of a lock pin, when an internal combustion engine stops. One such hydraulically-operated vane rotor equipped variable valve timing control device has been disclosed in Japanese Patent Provisional Publication No. 2010-261312 (hereinafter is referred to as “JP2010-261312”). In the valve timing control device disclosed in JP2010-261312, during an engine stopping period, the vane rotor rotates to the intermediate position by virtue of a fluttering motion of the vane rotor, caused by positive and negative alternating torque, acting on a camshaft due to spring forces of valve springs. Immediately when the vane rotor reaches the intermediate position, the vane rotor is locked and held at the intermediate position by engagement of the lock pin with a lock hole.

For instance, assume that the engine stalls under a specific condition where the lock pin is located nearer the phase-retard side than the intermediate position during a low-temperature engine operating condition that a viscosity of working fluid is high, and thus the engine is cranked for restarting. Under the previously-discussed specific condition, owing to a viscous resistance of working fluid in phase-retard hydraulic chambers and phase-advance hydraulic chambers, there is an increased tendency for a fluttering motion of the vane rotor to reduce. This undesirably lengthens a travelling time of the lock pin reaching its intermediate lock position, thus deteriorating a startability of the engine.

To avoid this, J22010-261312 teaches the provision of an auxiliary working-fluid discharge (exhaust) passage through which working fluid in each of the hydraulic chambers is released or exhausted to the outside of the VTC device, thereby increasing a fluttering motion of the vane rotor during the cranking and restarting period, and consequently causing a more smooth and rapid movement of the lock pin toward the intermediate lock position.

SUMMARY OF THE INVENTION

However, in the valve timing control device as disclosed in JP2010-261312, owing to a high viscous resistance of working fluid during a low-temperature engine operating condition as well as a large flow resistance that impedes the flow of working fluid through the auxiliary discharge (exhaust) passage, the system has the difficulty of rapidly discharging working fluid from each of the hydraulic chambers through the auxiliary discharge (exhaust) passage to the outside of the VTC device. Hence, it is difficult to ensure a rapid rotary motion of the vane rotor toward the intermediate lock position under the specific condition, in particular, during a low-temperature engine operating condition.

Accordingly, it is an object of the invention to provide a valve timing control apparatus of an internal combustion engine capable of more rapidly rotating a vane rotor to its lock position during a starting period of the engine.

In order to accomplish the aforementioned and other objects of the present invention, a valve timing control apparatus of an internal combustion engine, comprises a housing adapted to be driven by a crankshaft of the engine, and configured to define working-fluid chambers therein by partitioning an internal space by shoes protruding radially inward from an inner peripheral surface of the housing, a vane rotor having a rotor adapted to be fixedly connected to a camshaft and radially-extending vanes formed on an outer periphery of the rotor for partitioning each of the working-fluid chambers of the housing by the shoes and the vanes to define phase-advance hydraulic chambers and phase-retard hydraulic chambers, a lock mechanism configured to lock or unlock, depending on a condition on the engine, the vane rotor in a specified angular position between a maximum phase-retard angular position and a maximum phase-advance angular position of the vane rotor relative to the housing, and at least one recessed-groove fluid-communication passage formed in a portion of the housing being in sliding-contact with an'associated one of the vanes, a circumferential length of the fluid-communication passage being dimensioned to be greater than a circumferential width of the associated vane, wherein, at the maximum phase-retard angular position of the vane rotor relative to the housing, one circumferential end of the fluid-communication passage is formed in a position further displaced from the maximum phase-retard angular position of the associated vane in a phase-retard direction to face an associated one of the phase-advance hydraulic chambers and the other circumferential end of the fluid-communication passage is formed to face an associated one of the phase-retard hydraulic chambers, or at the maximum phase-advance angular position of the vane rotor relative to the housing one circumferential end of the fluid-communication passage is formed in a position further displaced from the maximum phase-advance angular position of the associated vane in a phase-advance direction to face the associated phase-retard hydraulic chamber and the other circumferential end of the fluid-communication passage is formed to face the associated phase-advance hydraulic chamber.

According to another aspect of the invention, a valve timing control apparatus of an internal combustion engine, comprises a housing adapted to be driven by a crankshaft of the engine, and configured to define working-fluid chambers therein by partitioning an internal space by shoes protruding radially inward from an inner peripheral surface of the housing, a vane rotor having a rotor adapted to be fixedly connected to a camshaft and radially-extending vanes formed on an outer periphery of the rotor for partitioning each of the working-fluid chambers of the housing by the shoes and the vanes to define phase-advance hydraulic chambers and phase-retard hydraulic chambers, a lock mechanism configured to lock or unlock, depending on a condition on the engine, the vane rotor in a specified angular position between a maximum phase-retard angular position and a maximum phase-advance angular position of the vane rotor relative to the housing, a control valve configured to control working-fluid supply-and-exhaust for each of the phase-advance hydraulic chambers and working-fluid supply-and-exhaust for each of the phase-retard hydraulic chambers, a controller configured to control operation of the control valve, and at least one recessed-groove passage formed in a portion of the housing being in sliding-contact with an associated one of the vanes, and configured to switch between a communicating state of an associated one of the phase-retard hydraulic chambers and an associated one of the phase-advance hydraulic chambers and a non-communicating state of the associated phase-retard hydraulic chamber and the associated phase-advance hydraulic chamber by relative rotation of the vane rotor with respect to the housing, wherein the recessed-groove passage is configured to permit the communicating state of the associated phase-retard hydraulic chamber and the associated phase-advance hydraulic chamber in at least one of the maximum phase-retard angular position and the maximum phase-advance angular position of the vane rotor relative to the housing, and configured to enable a transition from the communicating state to the non-communicating state when the vane rotor has rotated relatively to the housing by a specified angle or more in an opposite direction from the maximum phase-retard angular position or the maximum phase-advance angular position of the vane rotor relative to the housing.

According to a further aspect of the invention, a valve timing control apparatus of an internal combustion engine, comprises a driving rotary member adapted to be driven by a crankshaft of the engine, a driven rotary member adapted to be fixedly connected to a camshaft, and configured to partition an internal space of the driving rotary member into a phase-advance hydraulic chamber and a phase-retard hydraulic chamber, and configured to rotate the driven rotary member relative to the driving rotary member in a phase-advance direction by supplying working fluid to the phase-advance hydraulic chamber and exhausting working fluid from the phase-retard hydraulic chamber, and configured to rotate the driven rotary member relative to the driving rotary member in a phase-retard direction by supplying working fluid to the phase-retard hydraulic chamber and exhausting working fluid from the phase-advance hydraulic chamber, a lock mechanism configured to lock or unlock, depending on a condition on the engine, the driven rotary member in a specified angular position between a maximum phase-retard angular position and a maximum phase-advance angular position of the driven rotary member relative to the driving rotary member, at least one recessed-groove passage formed in a portion of the driving rotary member being in sliding-contact with the driven rotary member, and configured to switch between a communicating state and a non-communicating state of the phase-retard hydraulic chamber and the phase-advance hydraulic chamber by relative rotation of the driven rotary member with respect to the driving rotary member, wherein, the recessed-groove passage is configured to permit the communicating state of the phase-retard hydraulic chamber and the phase-advance hydraulic chamber in at least one of the maximum phase-retard angular position and the maximum phase-advance angular position of the driven rotary member relative to the driving rotary member, and configured to enable a transition from the communicating state to the non-communicating state when the driven rotary member has rotated relatively to the driving rotary member by a specified angle or more in an opposite direction from the maximum phase-retard angular position or the maximum phase-advance angular position of the driven rotary member relative to the driving rotary member.

The other objects and features of this invention will become understood from the following description with reference to the accompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a system diagram illustrating an embodiment of a valve timing control apparatus.

FIG. 2 is an exploded perspective view illustrating the valve timing control (VTC) apparatus of the embodiment, highlighting the essential part of the apparatus.

FIG. 3 is a cross-sectional view taken along the line A-A in FIG. 1 and showing a maximum phase-retard state where the vane rotor of the VTC apparatus of the embodiment has been rotated to an angular position corresponding to a maximum retarded phase.

FIG. 4 is a view taken in the direction of the arrow B in FIG. 1.

FIGS. 5A-5B are partial cross sections taken along the line C-C in FIG. 4, FIG. 5A illustrating a communicating state between a phase-advance hydraulic chamber and a phase-retard hydraulic chamber through a recessed-groove passage with the vane rotor held at the maximum phase-retard position, whereas FIG. 5B illustrating a non-communicating state between the phase-advance chamber and the phase-retard chamber with the vane rotor held at an angular position slightly displaced from the maximum phase-retard position to the phase-advance side.

FIG. 6 is a cross-sectional view taken along the line A-A in FIG. 1 and showing an intermediate phase state where the vane rotor of the VTC apparatus is held at an angular position corresponding to an intermediate phase.

FIG. 7 is a cross-sectional view taken along the line A-A in FIG. 1 and showing a maximum phase-advance state where the vane rotor of the VTC apparatus has been rotated to an angular position corresponding to a maximum advanced phase.

FIG. 8 is a development cross-sectional view illustrating an operation of each of lock pins with the vane rotor held at the maximum phase-retard position.

FIG. 9 is a development cross-sectional view illustrating another operation of each of the lock pins with the vane rotor slightly rotated from the maximum phase-retard position to the phase-advance side owing to alternating torque.

FIG. 10 is a development cross-sectional view illustrating a further operation of each of the lock pins with the vane rotor further rotated from the angular position of FIG. 9 to the phase-advance side.

FIG. 11 is a development cross-sectional view illustrating a still further operation of each of the lock pins with the vane rotor further rotated from the angular position of FIG. 10 to the phase-advance side.

FIG. 12 is a development cross-sectional view illustrating another operation of each of the lock pins with the vane rotor further rotated from the angular position of FIG. 11 to the phase-advance side.

FIG. 13 is a development cross-sectional view illustrating a further operation of each of the lock pins with the vane rotor further rotated from the angular position of FIG. 12 to the phase-advance side.

FIGS. 14A-14B are partial cross sections highlighting the essential part of the VTC apparatus of the second embodiment, FIG. 14A illustrating a communicating state between a phase-advance hydraulic chamber and a phase-retard hydraulic chamber through a recessed-groove passage of the VTC apparatus of the second embodiment with the vane rotor held at the maximum phase-retard position, whereas FIG. 14B illustrating a non-communicating state between the phase-advance chamber and the phase-retard chamber with the vane rotor held at an angular position slightly displaced from the maximum phase-retard position to the phase-advance side.

FIG. 15 is a front elevation, viewed from the front-plate side of the VTC apparatus of third embodiment.

FIGS. 16A-16B are partial cross sections highlighting the essential part of the VTC apparatus of the fourth embodiment, FIG. 16A illustrating a communicating state between a phase-advance hydraulic chamber and a phase-retard hydraulic chamber through a recessed-groove passage of the VTC apparatus of the fourth embodiment with the vane rotor held at the maximum phase-retard position, whereas FIG. 16B illustrating a non-communicating state between the phase-advance chamber and the phase-retard chamber with the vane rotor held at an angular position slightly displaced from the maximum phase-retard position to the phase-advance side.

FIG. 17 is a partial cross section highlighting the essential part of the VTC apparatus of the fifth embodiment, and illustrating a communicating state between a phase-advance hydraulic chamber and a phase-retard hydraulic chamber through a recessed-groove passage of the VTC apparatus of the fifth embodiment with the vane rotor held at the maximum phase-retard position.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

Referring now to the drawings, particularly to FIGS. 1-3, the valve timing control apparatus of the embodiment is exemplified in a phase control apparatus which is applied to an intake-valve side of an internal combustion engine of a hybrid electric vehicle (HEV), an idling-stop system equipped automotive vehicle, and the like.

As shown in FIGS. 1-3, the valve timing control apparatus includes a timing sprocket 1 driven by an engine crankshaft via a timing chain and serving as a driving rotary member, an intake-valve side camshaft 2 arranged in a longitudinal direction of the engine and configured to be relatively rotatable with the sprocket 1, a phase-change mechanism 3 installed between sprocket 1 and camshaft 2 to change a relative angular phase of camshaft 2 to sprocket 1 (the crankshaft), a lock mechanism 4 provided for locking or holding the phase-change mechanism 3 at a maximum phase-retard position as well as an intermediate-phase angular position between a maximum phase-advance position and the maximum phase-retard position, and a hydraulic circuit 5 provided for hydraulically operating phase-change mechanism 3 and lock mechanism 4 independently of each other.

Sprocket 1 is constructed as a rear cover that hermetically closes the rear end opening of a housing (described later). Sprocket 1 is formed into a thick-walled disc-shape. The outer periphery of sprocket 1 has a toothed portion 1a on which the timing chain is wound. Sprocket 1 is also formed with a supported bore 6 (a central through hole), which is rotatably supported on the outer periphery of one axial end 2a of camshaft 2. Also, sprocket 1 has circumferentially equidistant-spaced four female-screw threaded holes 1b formed on its outer peripheral side.

Camshaft 2 is rotatably supported on a cylinder head (not shown) via cam bearings (not shown). Camshaft 2 has a plurality of cams integrally formed on its outer periphery and spaced apart from each other in the axial direction of camshaft 2, for operating engine valves (i.e., intake valves). Camshaft 2 has a female-screw threaded hole 2b formed along the camshaft center at the axial end 2a.

As shown in FIGS. 1-3, phase-change mechanism 3 is comprised of a housing 7, a vane rotor 9, four phase-retard hydraulic chambers (simply, four phase-retard chambers) 11, 11, 11, 11 and four phase-advance hydraulic chambers (simply, four phase-advance chambers) 12, 12, 12, 12. Housing 7 is integrally connected to the sprocket 1 in the axial direction. Vane rotor 9 is fixedly connected to the axial end of camshaft 2 by means of a cam bolt 8 screwed into the female screw-threaded hole 2b of the axial end of camshaft 2, and serves as a driven rotary member rotatably enclosed in the housing 7. Housing 7 has radially-inward protruded four shoes (described later) integrally formed on the inner peripheral surface of housing 7. Four phase-retard chambers 11 and four phase-advance chambers 12 are defined by partitioning the working-fluid chamber (the internal space) of housing 7 by four shoes of housing 7 and four vanes (described later) of vane rotor 9.

Housing 7 includes a cylindrical housing body 10, a front plate 13, and the sprocket 1 serving as the rear cover for the rear opening end of housing 7. Housing body 10 is formed as a cylindrical hollow housing member, opened at both ends in the two opposite axial directions. Front plate 13 is produced by pressing. Front plate 13 is provided for hermetically covering the front opening end of housing body 10.

Housing body 10 is made of sintered alloy materials, such as iron-based sintered alloy materials. Housing body 10 has four radially-inward protruded shoes 10a, 10b, 10c, and 10d, integrally formed on its inner periphery. Four bolt insertion holes, namely axial through holes 10e, 10e, 10e, 10e are formed in respective shoes 10a-10d.

Front plate 13 is formed as a thin-walled metal disc. Front plate 13 is formed with a central through hole 13a. Also, front plate 13 has four circumferentially equidistant-spaced bolt insertion holes, namely axial through holes 13b, 13b, 13b, 13b.

Sprocket 1, housing body 10, and front plate 13 are integrally connected to each other by fastening them together with four bolts 14, 14, 14, 14 penetrating respective bolt insertion holes (i.e., four through holes 13b formed in the front plate 13 and four through holes 10e formed in respective shoes 10a-10d) and screwed into respective female-screw threaded holes 1b of sprocket 1.

In FIGS. 2-3, a pin denoted by reference sign 60 is a positioning pin attached onto the inner face 1c of sprocket 1, whereas an axially-elongated groove denoted by reference sign 61 is a positioning groove formed in the outer periphery of the first shoe 10a of housing body 10. When assembling, the positioning pin 60 of sprocket 1 is fitted into the positioning groove 61 of the first shoe 10a of housing body 10, thus ensuring easy positioning of housing body 10 relative to the sprocket 1.

Vane rotor 9 is formed of a metal material. Vane rotor 9 is comprised of a rotor 15 fixedly connected to the axial end of camshaft 2 by means of the cam bolt 8, and four radially-extending vane blades (simply, vanes) 16a, 16b, 16c, and 16d, formed on the outer periphery of rotor 15 and circumferentially spaced apart from each other by approximately 90 degrees.

Rotor 15 is formed into an axially-thick-walled, different-diameter deformed disc-shape. Rotor 15 is integrally formed with a central bolt insertion hole (an axial through hole) 15a. A substantially circular recessed bearing surface 15b, on which the head of cam bolt 8 is seated, is formed in the front end face of rotor 15.

Regarding the shape of rotor 15, in particular, the lateral cross-sectional configuration of rotor 15, the contour between the first vane 16a and the fourth vane 16d circumferentially adjacent to each other is configured as a small-diameter portion 15c, whereas the contour between the second vane 16b and the third vane 16c circumferentially adjacent to each other is also configured as a small-diameter portion 15d. The small-diameter pair (i.e., the first small-diameter portion 15c and the second small-diameter portion 15d) serves as a base circle. In contrast, the contour between the first vane 16a and the second vane 16b circumferentially adjacent to each other is configured as a first large-diameter portion 15e having an outside diameter greater than the first and second small-diameter portions 15c-15d. Also, the contour between the third vane 16c and the fourth vane 16d circumferentially adjacent to each other is configured as a second large-diameter portion 15f having an outside diameter greater than the first and second small-diameter portions 15c-15d.

First small-diameter portion 15c and second small-diameter portion 15d are arranged at angular positions circumferentially spaced apart from each other by approximately 180 degrees. That is, first and second small-diameter portions 15c-15d are arranged to be diametrically opposed to each other. The outer peripheral surface of each of first and second small-diameter portions 15c-15d is formed into a circular-arc shape having the same radius of curvature.

On the other hand, first and second large-diameter portions 15e-15f are arranged at angular positions circumferentially spaced apart from each other by approximately 180 degrees. That is, first and second large-diameter portions 15e-15f are also arranged to be diametrically opposed to each other. The outer peripheral surface of each of first and second large-diameter portions 15e-15f is formed into a circular-arc shape having the same radius of curvature. However, the outside diameter of the outer peripheral surfaces of large-diameter portions 15e-15f is configured to be one-size greater than that of small-diameter portions 15c-15d.

Therefore, the first shoe 10a, whose tip faces the outer peripheral surface of first small-diameter portion 15c, is formed as a comparatively long, radially-inward protruded partition wall having substantially rectangular side faces. In a similar manner, the second shoe 10b, whose tip faces the outer peripheral surface of second small-diameter portion 15d, is formed as a comparatively long, radially-inward protruded partition wall having substantially rectangular side faces. In contrast, the third shoe 10c, whose tip faces the outer peripheral surface of first large-diameter portion 15e, is formed as a comparatively short, radially-inward protruded partition wall having substantially circular-arc side faces. In a similar manner, the fourth shoe 10d, whose tip faces the outer peripheral surface of second large-diameter portion 15f, is formed as a comparatively short, radially-inward protruded partition wall having substantially circular-arc side faces.

Four shoes 10a-10d have respective axially-elongated seal retaining grooves, formed in their innermost ends (apexes) and extending in the axial direction. Each of four seal retaining grooves of the shoes is formed into a substantially rectangle. Four oil seal members (four apex seals) 17a, 17a, 17a, 17a, each having a substantially square lateral cross section, are fitted into respective seal retaining grooves of four shoes 10a-10d so as to bring the four apex seals 17a into sliding-contact with the respective outer peripheral surfaces of first and second small-diameter portions 15c-15d and first and second large-diameter portions 15e-15f. Leaf springs (not shown) are installed on the respective bottom faces of the seal retaining grooves of four shoes 10a-10d, for permanently biasing the four apex seals of four shoes 10a-10d toward the respective outer peripheral surfaces of first and second small-diameter portions 15c-15d and first and second large-diameter portions 15e-15f, thereby providing a sealing action between the different-diameter deformed outer peripheral surface of rotor 15 and the innermost ends (apexes) of shoes 10a-10d.

Regarding four vanes 16a-16d formed integral with the rotor 15 and radially extending outward from the outer peripheral surface of rotor 15, their entire lengths are dimensioned to be substantially identical to each other. Circumferential widths of four vanes 16a-16d are dimensioned to be substantially identical to each other, and thus each of vanes 16a-16d is formed into a thin-walled plate. Four vanes 16a-16d are disposed in respective internal spaces defined by four shoes 10a-10d. In a similar manner to the four shoes 10a-10d, four vanes 16a-16d have respective axially-elongated seal retaining grooves 17g, 17g, 17g, 17g, formed in their outermost ends (apexes) and extending in the axial direction. Each of four seal retaining grooves of the vanes is formed into a substantially rectangle. Four oil seal members (four apex seals) 17b, 17b, 17b, 17b, each having a substantially square lateral cross section, are fitted into respective seal retaining grooves 17g of four vanes 16a-16d so as to bring the four apex seals 17b into sliding-contact with the inner peripheral surface of housing body 10. Leaf springs (not shown) are installed in the respective seal retaining grooves 17g of four vanes 16a-16d, for permanently biasing the four apex seals of four vanes 16a-16d toward the inner peripheral surface of housing body 10, thereby providing a sealing action between the inner peripheral surface of housing body 10 and the outermost ends (apexes) of vanes 16a-16d.

As discussed above, apex seals 17a of shoes 10a-10d and apex seals 17b of vanes 16a-16d are cooperated with each other to ensure a fluid-tight sealing structure between phase-retard chamber 11 and phase-advance chamber 12.

As shown in FIG. 3, when vane rotor 9 rotates relative to the housing 7 (or the sprocket 1) in the phase-retard direction, one side face (an anticlockwise side face 16e, viewing FIG. 3) of the first vane 16a is brought into abutted-engagement with a radially-inward protruding surface formed on one side face (a clockwise side face, viewing FIG. 3) of the opposed first shoe 10a, and thus a maximum phase-retard angular position of vane rotor 9 is restricted. Conversely, as shown in FIG. 7, when vane rotor 9 rotates relative to the housing 7 (or the sprocket 1) in the phase-advance direction, the other side face (a clockwise side face, viewing FIG. 7) of the first vane 16a is brought into abutted-engagement with a radially-inward protruding surface formed on one side face (an anticlockwise side face, viewing FIG. 7) of the opposed third shoe 10c, and thus a maximum phase-advance angular position of vane rotor 9 is restricted. That is, the third shoe 10c cooperates with the first vane 16a to provide a stopper function (i.e., a maximum phase-advance side stopper) for restricting a maximum phase-advance angular position of vane rotor 9 (in other words, rotary motion of vane rotor 9 relative to sprocket 1 in the phase-advance direction). In a similar manner, the first shoe 10a cooperates with the first vane 16a to provide a stopper function (i.e., a maximum phase-retard side stopper) for restricting a maximum phase-retard angular position of vane rotor 9 (in other words, rotary motion of vane rotor 9 relative to sprocket 1 in the phase-retard direction).

With the first vane 16a kept in its maximum phase-retard angular position (see FIG. 3) or with the first vane 16a kept in its maximum phase-advance angular position (see FIG. 7), both side faces of each of the other vanes 16b-16d are kept in a spaced, contact-free relationship with respective side faces of the associated shoes. Hence, the accuracy of abutment between the vane rotor 9 and the shoe (i.e., the first shoe 10a) can be enhanced, and additionally the speed of hydraulic pressure supply to each of hydraulic chambers 11 and 12 can be increased, thus a responsiveness of normal-rotation/reverse-rotation of vane rotor 9 can be improved.

By the way, during normal relative-rotation control of vane rotor 9 to housing 7, rotary motion of vane rotor 9 relative to housing 7 is controlled within a somewhat narrow relative-rotation angular range between an angular position slightly displaced circumferentially inside of (clockwise from) a maximum retarded phase that the first vane 16a is kept in abutted-engagement with the associated phase-retard side shoe (i.e., the first shoe 10a) and an angular position slightly displaced circumferentially inside of (anticlockwise from) a maximum advanced phase that the first vane 16a is kept in abutted-engagement with the associated phase-advance side shoe (i.e., the third shoe 10c).

The previously-discussed four phase-retard chambers 11 and four phase-advance chambers 12 are defined by both side faces of each of vanes 16a-16d and both side faces of each of shoes 10a-10d. Regarding volumetric capacities of phase-retard chambers 11 and phase-advance chambers 12, by virtue of the different-diameter deformed outer peripheral surface of rotor 15, the total volumetric capacity of hydraulic chambers 11a and 12a, located in the area corresponding to the small-diameter portion (each of first and second small-diameter portions 15c-15d) of rotor 15, is set to be greater than the total volumetric capacity of hydraulic chambers 11b and 12b, located in the area corresponding to the large-diameter portion (each of first and second large-diameter portions 15e-15f). Thus, the pressure-receiving surface area of each of side faces 16e-16h of vanes 16a-16d, facing hydraulic chambers 11a and 12a located in the area corresponding to the small-diameter portion (each of first and second small-diameter portions 15c-15d), is set to be greater than that of each of side faces of vanes 16a-16d, facing hydraulic chambers 11b and 12b located in the area corresponding to the large-diameter portion (each of first and second large-diameter portions 15e-15f).

Each of phase-retard chambers 11 is configured to communicate with the hydraulic circuit 5 (described later) via the first communication hole 11c formed in the rotor 15. In a similar manner, each of phase-advance chambers 12 is configured to communicate with the hydraulic circuit 5 via the second communication hole 12c formed in the rotor 15.

Lock mechanism 4 is provided for holding or locking an angular position of vane rotor 9 relative to housing 7 either at an intermediate-phase angular position, corresponding to the angular position (an intermediate lock position) of vane rotor 9 in FIG. 6 between the maximum phase-retard angular position (see FIG. 3) and the maximum phase-advance angular position (see FIG. 7), or at the maximum phase-retard angular position, depending on whether the engine is stopped manually by turning an ignition switch OFF or automatically stopped by means of an idling-stop system.

That is, as shown in FIGS. 2 and 8-13, lock mechanism 4 is comprised of a first lock hole 24, a second lock hole 25, a third lock hole 26, a first lock pin 27, a second lock pin 28, a third lock pin 29, and a lock-unlock passage (simply, a lock passage) 20. First, second and third lock holes 24-26 (serving as first, second and third lock recessed portions) are disposed in the inner face 1c of sprocket 1, and arranged at respective given circumferential positions. The first lock pin 27 (serving as a substantially cylindrical locking member engaged with the associated recessed portion) is operably disposed in the first large-diameter portion 15e of rotor 15 such that movement of first lock pin 27 into and out of engagement with the first lock hole 24 is permitted. The second lock pin 28 (serving as a substantially cylindrical locking member) is operably disposed in the first large-diameter portion 15e of rotor 15 such that movement of second lock pin 28 into and out of engagement with the second lock hole 25 is permitted. In a similar manner, the third lock pin 29 (serving as a substantially cylindrical locking member) is operably disposed in the second large-diameter portion 15f of rotor 15 such that movement of third lock pin 29 into and out of engagement with the third lock hole 26 is permitted. First, second and third lock pins 27-29 are arranged at respective given circumferential positions of rotor 15. Lock passage 20 is provided for disengagement of the first lock pin 27 from the first lock hole 24 and for disengagement of the second lock pin 28 from the second lock hole 25 and for disengagement of the third lock pin 29 from the third lock hole 26.

As seen in FIGS. 2 and 8-13, the first lock hole 24 is arranged on the side of first large-diameter portion 15e. The first lock hole 24 is formed into a cylindrical-hollow shape having an inside diameter greater than an outside diameter of the tip 27a of first lock pin 27 so as to permit a slight circumferential movement of the tip 27a of first lock pin 27 engaged with the first lock hole 24. Also, the first lock hole 24 is formed in the inner face 1c of sprocket 1 and arranged at an intermediate position somewhat displaced toward the phase-advance side with respect to the maximum phase-retard angular position of vane rotor 9. Additionally, the depth of the bottom face 24a of first lock hole 24 is dimensioned or set to be almost the same depth as the second bottom face 25b of second lock hole 25 and also dimensioned to be almost the same depth as the second bottom face 26b of third lock hole 26. Hence, in the presence of movement of first lock pin 27 into engagement with the first lock hole 24 owing to rotary motion of the vane rotor 9 in the phase-advance direction, the tip 27a of first lock pin 27 is brought into abutted-engagement with the bottom face 24a of first lock hole 24. At the same time, the outer periphery (the edge) of the tip 27a of first lock pin 27 is brought into abutted-engagement with the upstanding inner face 24b of first lock hole 24, and whereby rotary motion of vane rotor 9 in the phase-retard direction is restricted (see FIG. 13).

The second lock hole 25 is arranged on the side of first large-diameter portion 15e, in a similar manner to the first lock hole 24. The second lock hole 25 is formed into an elliptic or oval shape (a circumferentially-elongated groove) extending in the circumferential direction of sprocket 1. That is, the second lock hole 25 is formed as a two-stage stepped hole whose bottom face lowers stepwise from the phase-retard side to the phase-advance side. The second lock hole 25 (i.e., the two-stage stepped groove) is configured to serve as a second lock guide groove. That is, assuming that the inner face 1c of sprocket 1 is regarded as the uppermost level, the second lock guide groove (the two-stage stepped groove) 25 is configured to gradually lower from the first bottom face 25a to the second bottom face 25b, in that order. Each of inner faces, vertically extending from respective bottom faces 25a-25b on the phase-retard side, is formed as an upstanding wall surface (viewing FIGS. 8-13). The inner face 25c, vertically extending from the second bottom face 25b on the phase-advance side, is also formed as an upstanding wall surface (viewing FIGS. 8-13).

The second bottom face 25b is formed as a somewhat circumferentially-elongated recessed groove extending to the phase-advance side. With the tip 28a of second lock pin 28 engaged with the second bottom face 25b, the somewhat circumferentially-elongated second bottom face 25b permits a slight movement of second lock pin 28 in the phase-advance direction (see FIGS. 12-13).

The third lock hole 26 is arranged on the side of second large-diameter portion 15f and formed into a cocoon shape (or a circular-arc circumferentially-elongated groove) extending in the circumferential direction of sprocket 1 and dimensioned to be longer than the second lock hole 25. The third lock hole 26 is formed in the inner face 1c of sprocket 1 and arranged at an intermediate position somewhat displaced toward the phase-advance side with respect to the maximum phase-retard angular position of vane rotor 9. Additionally, the third lock hole 26 is formed as a two-stage stepped hole whose bottom face lowers stepwise from the phase-retard side to the phase-advance side. The third lock hole 26 (i.e., the two-stage stepped groove) is configured to serve as a lock guide groove.

That is, as seen in FIGS. 8-13, assuming that the inner face 1c of sprocket 1 is regarded as an uppermost level, the third lock guide groove (the two-stage stepped groove) 26 is configured to gradually lower from the first bottom face 26a to the second bottom face 26b, in that order. Each of inner faces, vertically extending from respective bottom faces 26a-26b on the phase-retard side, is formed as an upstanding wall surface (viewing FIGS. 8-13). The inner face 26c, vertically extending from the second bottom face 26b on the phase-advance side, is also formed as an upstanding wall surface (viewing FIGS. 8-13).

As best seen in FIGS. 2 and 8-13, the first lock pin 27 is slidably disposed in a first lock-pin hole 31a (an axial through hole) formed in the first large-diameter portion 15e of rotor 15. The first lock pin 27 is contoured as a stepped shape, comprised of the comparatively small-diameter tip 27a, a comparatively large-diameter cylindrical-hollow basal portion 27b integrally formed continuously with the rear end of small-diameter tip 27a, and a stepped pressure-receiving surface 27c defined between the tip 27a and the large-diameter cylindrical-hollow basal portion 27b. The end face of tip 27a is formed as a flat face, which can be brought into abutted-engagement (exactly, into wall-contact) with each of bottom faces 24a and 24b.

The first lock pin 27 is permanently biased in a direction of movement of first lock pin 27 into engagement with the first lock hole 24 by a spring force of a first spring 36 (a first biasing member or first biasing means). The first spring 36 is disposed between the bottom face of an axial spring bore formed in the large-diameter cylindrical-hollow basal portion 27b in a manner so as to axially extend from the rear end face and the inner wall surface of front cover 13 under preload.

The first lock pin 27 is also configured such that hydraulic pressure from a first unlocking pressure-receiving chamber 32, which chamber is formed in the rotor 15, is applied to the stepped pressure-receiving surface 27c. The applied hydraulic pressure causes a backward movement of first lock pin 27 against the spring force of first spring 36, and thus the first lock pin 27 is disengaged from the first lock hole 24.

In a similar manner to the first lock pin 27, the second lock pin 28 is slidably disposed in a second lock-pin hole 31b (an axial through hole) formed in the first large-diameter portion 15e of rotor 15. The second lock pin 28 is contoured as a stepped shape, comprised of the comparatively small-diameter tip 28a, a comparatively large-diameter cylindrical-hollow basal portion 28b integrally formed continuously with the rear end of small-diameter tip 28a, and a stepped pressure-receiving surface 28c defined between the tip 28a and the large-diameter cylindrical-hollow basal portion 28b. The end face of tip 28a is formed as a flat face, which can be brought into abutted-engagement (exactly, into wall-contact) with each of bottom faces 25a and 25b.

The second lock pin 28 is permanently biased in a direction of movement of second lock pin 28 into engagement with the second lock hole 25 by a spring force of a second spring 37 (a second biasing member or second biasing means). The second spring 37 is disposed between the bottom face of an axial spring bore formed in the large-diameter cylindrical-hollow basal portion 28b in a manner so as to axially extend from the rear end face and the inner wall surface of front cover 13 under preload.

The second lock pin 28 is also configured such that hydraulic pressure from a second unlocking pressure-receiving chamber 33, which chamber is formed in the rotor 15, is applied to the stepped pressure-receiving surface 28c. The applied hydraulic pressure causes a backward movement of second lock pin 28 against the spring force of second spring 37, and thus the second lock pin 28 is disengaged from the second lock hole 25.

The third lock pin 29 is slidably disposed in a third lock-pin hole 31c (an axial through hole) formed in the second large-diameter portion 15f of rotor 15. The third lock pin 29 is contoured as a stepped shape, comprised of the comparatively small-diameter tip 29a, a comparatively large-diameter cylindrical-hollow basal portion 29b integrally formed continuously with the rear end of small-diameter tip 29a, and a stepped pressure-receiving surface 29c defined between the tip 29a and the large-diameter cylindrical-hollow basal portion 29b. The end face of tip 29a is formed as a flat face, which can be brought into abutted-engagement (exactly, into wall-contact) with the bottom face 26a.

The third lock pin 29 is permanently biased in a direction of movement of third lock pin 29 into engagement with the third lock hole 26 by a spring force of a third spring 38 (a third biasing member or third biasing means). The third spring 38 is disposed between the bottom face of an axial spring bore formed in the large-diameter cylindrical-hollow basal portion 29b in a manner so as to axially extend from the rear end face and the inner wall surface of front cover 13 under preload.

The third lock pin 29 is also configured such that hydraulic pressure from a third unlocking pressure-receiving chamber 34, which chamber is formed in the rotor 15, is applied to the stepped pressure-receiving surface 29c. The applied hydraulic pressure causes a backward movement of third lock pin 29 against the spring force of third spring 38, and thus the third lock pin 29 is disengaged from the third lock hole 26.

The relative-position relationship of first, second, and third lock holes 24-26 formed in the inner face 1c of sprocket 1 and first, second, and third lock pins 27-28 located and installed in the rotor 15 is as follows.

That is, as seen in FIG. 8, when vane rotor 9 has rotated relative to sprocket 1 and reached the maximum phase-retard position, the first lock pin 27 is brought into engagement with the second lock hole 25, and thus the axial end face of the tip 27a of first lock pin 27 is brought into abutted-engagement with the second bottom face 25b of second lock hole 25 and simultaneously the outer periphery (the edge) of the tip 27a of first lock pin 27 is also brought into abutted-engagement with the phase-advance side upstanding inner face 25c.

Thereafter, with the first lock pin 27 sliding out of engagement with the second lock hole 25, suppose that vane rotor 9 somewhat rotates in the phase-advance direction from the maximum phase-retard position. In a phase wherein the third lock pin 29 is brought into engagement with the first bottom face 26a of third lock hole 26 (see FIG. 9), and in a phase just after the third lock pin 29 has been brought into engagement with the second bottom face 26b (see FIG. 10), the axial end face of the tip 27a of first lock pin 27 and the axial end face of the tip 28a of second lock pin 28 are still kept in abutted-engagement with the inner face 1c of sprocket 1.

Thereafter, as seen in FIG. 11, when, owing to a slight rotary motion of vane rotor 9 in the phase-advance direction, the axial end face of the tip 29a of third lock pin 29 slides along the second bottom face 26b of third lock hole 26 and then reaches a substantially midpoint of the second bottom face 26b, the tip 28a of second lock pin 28 is brought into abutted-engagement with the first bottom face 25a of second lock hole 25.

As seen in FIG. 12, when the tip 29a of third lock pin 29 further moves in the phase-advance direction, while being kept in sliding-contact with the second bottom face 26b, the tip 28a of second lock pin 28 slides out of engagement with the first bottom face 25a of second lock hole 25 but slides into abutted-engagement with the second bottom face 25b. At this time, the axial end face of the tip 27a of first lock pin 27 slides in the phase-advance direction, while being still kept in abutted-engagement with the inner face 1c of sprocket 1.

Thereafter, when, owing to a further rotary motion of vane rotor 9 in the phase-advance direction, the second lock pin 28 kept in abutted-engagement with the second bottom face 25b and the third lock pin 29 kept in abutted-engagement with the second bottom face 26b further move in the same phase-advance direction, the tip 27a of first lock pin 27 slides into engagement with the first lock hole 24 (see FIG. 13). In this manner, the relative-position relationship among first, second and third lock holes 24-26 and first, second and third lock pins 27-29 is preset. With three lock pins 27-29 engaged with respective lock holes 24-26, the circumferentially-opposed outer peripheral edges of first and second lock pins 27-28, circumferentially opposed to each other, abut with the circumferentially-opposed upstanding inner faces 24b and 25c of first and second lock holes 24-25, respectively, such that the specified area of the inner face 1c of sprocket 1, ranging between the two upstanding inner faces 24b and 25c, is sandwiched with the two lock pins 27-28.

At this time, as best seen in FIG. 13, a further movement of third lock pin 29 in the phase-advance direction is restricted by a combined locking action of first and second lock pins 27-28 (that is, by abutment of the outer periphery (the edge) of the tip 27a of first lock pin 27 with the upstanding inner face 24b and by abutment of the outer periphery (the edge) of the tip 28a of second lock pin 28 with the upstanding inner face 25c) under a specified state where the outer periphery of the tip 29a of third lock pin 29 is slightly spaced apart from the upstanding inner face 26c vertically extending from the second bottom face 26b.

Briefly speaking, as can be seen from the cross sections of FIGS. 8-13, according to rotary motion of vane rotor 9 relative to sprocket 1 from the phase-retard position toward the phase-advance position, the third lock pin 29 is brought into abutted-engagement with the first and second bottom faces 26a-26b, one-by-one (in a stepwise manner) and further moves in the phase-advance direction, while being kept in sliding-contact with the second bottom face 26b. From the middle of sliding movement of the tip 29a of third lock pin 29 along the second bottom face 26b, the second lock pin 28 slides into engagement with the second lock hole 25 and then brought into abutted-engagement with the first and second bottom faces 25a-25b, one-by-one (in a stepwise manner). Thereafter, the first lock pin 27 is sequentially brought into engagement with the first lock hole 24. As discussed above, the third and second lock guide groove structures (i.e., third and second holes 26-25) and the first lock hole 24 permit normal rotation of vane rotor 9 relative to sprocket 1 in the phase-advance direction, but restrict or prevent reverse-rotation (counter-rotation) of vane rotor 9 relative to sprocket 1 in the phase-retard direction by virtue of a five-stage ratchet action in total. Finally, the angular position of vane rotor 9 relative to sprocket 1 is held or locked at the intermediate-phase angular position (see FIG. 6) between the maximum phase-retard angular position (see FIG. 3) and the maximum phase-advance angular position (see FIG. 7).

Returning to FIG. 1, the rear end of each of first, second, and third lock-pin holes 31a-31c is configured to be opened to the atmosphere via a breather 39, thereby ensuring a smooth sliding movement of each of lock pins 27, 28 and 29.

As shown in FIG. 1, hydraulic circuit 5 includes a phase-retard passage 18, a phase-advance passage 19, lock passage 20, an oil pump 40 (serving as a fluid-pressure supply source), and a single electromagnetic directional control valve 41. Phase-retard passage 18 is provided for fluid-pressure supply-and-exhaust for each of phase-retard chambers 11 via the first communication hole 11c. Phase-advance passage 19 is provided for fluid-pressure supply-and-exhaust for each of phase-advance chambers 12 via the second communication hole 12c. Lock passage 20 is provided for fluid-pressure supply-and-exhaust for each of first, second, and third unlocking pressure-receiving chambers 32-34. Oil pump 40 is provided for supplying working fluid pressure to at least one of phase-retard passage 18 and phase-advance passage 19, and also provided for supplying working fluid pressure to lock passage 20. Single electromagnetic directional control valve 41 is provided for switching between phase-retard passage 18 and phase-advance passage 19, and also provided for switching between working-fluid supply to lock passage 20 and working-fluid exhaust from lock passage 20.

One end of phase-retard passage 18 and one end of phase-advance passage 19 are connected to respective ports (not shown) of electromagnetic directional control valve 41. The other end of phase-retard passage 18 is configured to communicate with each of phase-retard chambers 11 via an axial passage portion 18a formed in the camshaft 2 and the first communication hole 11c formed in the rotor 15. The other end of phase-advance passage 19 is configured to communicate with each of phase-advance chambers 12 via an axially-extending but partly-radially-bent passage portion 19a formed in the camshaft 2 and the second communication hole 12c formed in the rotor 15.

As shown in FIGS. 1-2, one end of lock passage 20 is connected to a lock port (not shown) of electromagnetic directional control valve 41. The other end of lock passage 20, serving as a fluid-passage portion 20a, is formed in the camshaft to be bent from the radial direction to the axial direction. The fluid-passage portion 20a of lock passage 20 is configured to communicate with respective unlocking pressure-receiving chambers 32-34 via branch oil holes 20b-20c formed in the rotor 15 and branching away.

In the shown embodiment, an internal gear rotary pump, such as a trochoid pump having inner and outer rotors, is used as the oil pump 40 driven by the engine crankshaft. During operation of oil pump 40, when the inner rotor is driven, the outer rotor also rotates in the same rotational direction as the inner rotor by mesh between the outer-rotor inner-toothed portion and the inner-rotor outer-toothed portion. Working fluid in an oil pan 42 is introduced through a suction passage into the pump, and then discharged through a discharge passage 40a. Part of working fluid discharged from oil pump 40 is delivered through a main oil gallery M/G to sliding or moving engine parts. The remaining working fluid discharged from oil pump 40 is delivered to electromagnetic directional control valve 41. An oil filter (not shown) is disposed in the downstream side of discharge passage 40a. Also, a flow control valve (not shown) is provided to appropriately control an amount of working fluid discharged from oil pump 40 into discharge passage 40a, thus enabling surplus working fluid discharged from oil pump 40 to be directed via a drain passage 43 to the oil pan 42.

As seen in FIG. 1, electromagnetic directional control valve 41 is an electromagnetic-solenoid operated, six-port, six-position, spring-offset, proportional control valve. Electromagnetic directional control valve 41 is comprised of a substantially cylindrical-hollow, axially-elongated valve body (a valve housing), a valve spool (an electrically-actuated valve element) slidably installed in the valve body in a manner so as to axially slide in a very close-fitting bore of the valve body, a valve spring installed inside of one axial end of the valve body for permanently biasing the valve spool in an axial direction, and an electromagnetic solenoid attached to the valve body so as to cause axial sliding movement of the valve spool against the spring force of the valve spring.

Electromagnetic directional control valve 41 is configured to move the valve spool to either one of six axial positions by the two opposing pressing forces, produced by a spring force of the valve spring and a control current generated from a controller 35 and flowing through the electromagnetic solenoid coil, so as to change a state of fluid-communication between the discharge passage 40a of oil pump 40 and each of three passages (that is, phase-retard passage 18, phase-advance passage 19, and lock passage 20) and simultaneously change a state of fluid-communication between the drain passage 43 and each of the three passages 18, 19, and 20, depending on a selected one of the six positions of the valve spool.

As discussed above, electromagnetic directional control valve 41 is configured to change the path of flow through the directional control valve 41 by selective switching among the ports depending on a given axial position of the valve spool, determined based on latest up-to-date information about an engine operating condition (e.g., engine speed and engine load), thereby changing a relative angular phase of vane rotor 9 (camshaft 2) to sprocket 1 (the crankshaft) and also enabling selective switching between locked and unlocked states of lock mechanism 4, in other words, selective switching between a locked (engaged) state of lock pins 27-29 with respective lock holes 24-26 and an unlocked (disengaged) state of lock pins 27-29 from respective lock holes 24-26. Accordingly, by means of electromagnetic directional control valve 41 as previously discussed, free rotation of vane rotor 9 relative to sprocket 1 can be enabled (permitted) or disabled (restricted) depending on the engine operating condition.

Controller (ECU) 35 generally comprises a microcomputer. Controller 35 includes an input/output interface (I/O), memories (RAM, ROM), and a microprocessor or a central processing unit (CPU). The input/output interface (I/O) of controller 35 receives input information from various engine/vehicle switches and sensors, namely a crank angle sensor (a crank position sensor), an airflow meter, an engine temperature sensor (e.g., an engine coolant temperature sensor), a throttle opening sensor (a throttle position sensor), a cam angle sensor, an oil-pump discharge pressure sensor, and the like. The crank angle sensor is provided for detecting revolution speeds of the engine crankshaft and for calculating an engine speed. The airflow meter is provided for generating an intake-air flow rate signal indicating an actual intake-air flow rate or an actual air quantity. The engine temperature sensor is provided for detecting an actual operating temperature of the engine. The cam angle sensor is provided for detecting latest up-to-date information about an angular phase of camshaft 2. The discharge pressure sensor is provided for detecting a discharge pressure of working fluid discharged from the oil pump 40. Within controller 35, the central processing unit (CPU) allows the access by the I/O interface of input informational data signals from the previously-discussed engine/vehicle switches and sensors, so as to detect the current engine operating condition, and also to generate a control pulse current, determined based on latest up-to-date information about the detected engine operating condition and the detected discharge pressure, to the electromagnetic coil of the solenoid of electromagnetic directional control valve 41, for controlling the axial position of the sliding valve spool, thus achieving selective switching among the ports depending on the controlled axial position of the valve spool.

As hereunder described in detail, output control for pulse current, applied to electromagnetic directional control valve 41, is classified into a so-called manual-engine-stop pulse-current output control, executed when the engine is stopped by manually turning the ignition switch OFF, and a so-called automatic-engine-stop pulse-current output control, executed when the engine is automatically temporarily stopped by means of an idling-stop system, for instance in accordance with idle-reduction (idle-stop) action.

Referring now to FIGS. 4, and 5A-5B, there are shown the position of formation of four recessed-groove passages 50 (serving as fluid-communication passages) and the communicating/non-communicating operation of recessed-groove passage 50.

As seen in FIGS. 4 and 5A-5B, the circumferentially-spaced four recessed-groove passages 50 are formed in the inside end face of front plate 13 of housing 7, and located on the side of phase-advance chamber 12. Each of recessed-groove passages 50 is simple in fluid-communication passage structure. This eliminates the necessity of separate fluid-communication conduits.

Concretely, each of recessed-groove passages 50 is formed as a substantially rectangular fluid-communication groove, cut in the inside end face of front plate 13. More concretely, each of recessed-groove fluid-communication passages 50 is formed into a somewhat circumferentially-elongated, circular-arc shape having a specified circumferential length L. A depth D from the inside end face of front plate 13 to a flat bottom face 50c of recessed-groove passage 50 is set to a substantially uniform depth. A radial length as well as the circumferential length L of recessed-groove passage 50 is dimensioned to be greater than the depth D.

Regarding the position of formation of each of recessed-groove passages 50, arranged concentrically with each other, one circumferential end (an anticlockwise end 50a) of each of recessed-groove passages 50 is formed in a position further displaced slightly from the maximum phase-retard position of each of vanes 16a-16d toward the phase-retard side, when vane rotor 9 has rotated relative to housing 7 and held at its maximum phase-retard position, that is, when vanes 16a-16d are held at their maximum phase-retard positions under a state where the maximum phase-retard angular position of vane rotor 9 has been restricted by abutment of the anticlockwise side face 16e of the first vane 16a with one side face (a clockwise side face 10f, viewing FIG. 4) of the opposed first shoe 10a. For instance, regarding the first vane 16a, as seen from the partial cross sections of FIGS. 5A-5B, the anticlockwise end 50a of the associated recessed-groove passage 50 is located in a position displaced slightly from the clockwise side face 10f of the first shoe 10a toward the phase-retard side.

Regarding the position of formation of each of recessed-groove passages 50 in the radial direction, the radially-inward end of each of recessed-groove passages 50 is formed or contoured along the outer peripheral surface of rotor 15, in particular, the large-diameter portions 15e-15f. On the other hand, the radially-outward ends of recessed-groove passages 50 are formed inside of the respective seal retaining grooves of vanes 16a-16d, in other words, inside of the respective apex seals 17b of vanes 16a-16d, such that recessed-groove passages 50 are not communicated with the respective seal retaining grooves.

The circumferential length L of each of recessed-groove passages 50 is dimensioned to be slightly greater than a circumferential width W of each of vanes 16a-16d. When vane rotor 9 is held at the maximum phase-retard position, the anticlockwise end 50a of recessed-groove passage 50 faces the associated phase-advance chamber 12, and the other circumferential end (a clockwise end 50b) of recessed-groove passage 50 faces the associated phase-retard chamber 11, thereby establishing fluid-communication between phase-retard chamber 11 and phase-advance chamber 12 through the recessed-groove passage 50.

Operation of Valve Timing Control Apparatus of Embodiment

Details of operation of the valve timing control apparatus of the embodiment are hereunder described.

[Manual-Engine-Stop]

For instance, when an ignition switch has been turned OFF after normal vehicle traveling and thus the engine has stopped rotating, a supply of control current from controller 35 to the electromagnetic coil of electromagnetic directional control valve 41 is stopped and thus the solenoid is de-energized. Thus, the valve spool is positioned at the maximum rightward axial position (i.e., the “first position”, in other words, the spring-loaded or spring-offset position or the initial valve position) by the spring force of the valve spring. Hence, the discharge passage 40a communicates with both of the phase-retard passage 18 and the phase-advance passage 19, whereas the lock passage 20 communicates with the drain passage 43.

At the same time, oil pump 40 is placed into an inoperative state, and thus working-fluid supply to phase-retard chamber 11 or phase-advance chamber 12 becomes stopped, and also working-fluid supply to each of first, second, and third unlocking pressure-receiving chambers 32-34 becomes stopped.

That is, during idling before the engine is brought into a stopped state, vane rotor 9 is placed into the maximum phase-retard angular position shown in FIG. 3 by the working-fluid pressure supply to each of phase-retard chambers 11. At this time, as seen in FIG. 8, the second and third lock pins 28-29 are kept out of engagement with the respective lock holes 25-26 but kept in abutted-engagement with the inner face 1c of sprocket 1 under preload. On the other hand, the first lock pin 27 is kept in engagement with the second lock hole 25.

Under these conditions, when the ignition switch becomes manually turned OFF, there is a pulse current output to the solenoid of electromagnetic directional control valve 41 immediately before the engine stops during the initial part of turning-OFF action of the ignition switch, and thus there is a working-fluid supply from oil pump 40 to each of unlocking pressure-receiving chambers 32-34 responsively to the pulse current output. Hence, as indicated by the one-dotted line in FIG. 8, a backward movement of first lock pin 27 against the spring force of first spring 36 occurs. As a result, the first lock pin 27 slides out of engagement with the second lock hole 25.

Also, immediately before the engine stops, alternating torque, acting on camshaft 2, occurs. In particular, when rotary motion of vane rotor 9 relative to sprocket 1 in the phase-advance direction occurs owing to the negative torque of alternating torque acting on camshaft 2 and thus the angular position of vane rotor 9 relative to sprocket 1 reaches the intermediate-phase angular position (see FIG. 6), the tip 27a of first lock pin 27, the tip 28a of second lock pin 28, and the tip 29a of third lock pin 29 are brought into engagement with respective lock holes 24-26 by the spring forces of first, second, and third springs 36-38 (see FIG. 13). As a result of this, the angular position of vane rotor 9 relative to sprocket 1 is held or locked at the intermediate-phase angular position (see FIG. 6) between the maximum phase-retard angular position (see FIG. 3) and the maximum phase-advance angular position (see FIG. 7).

More concretely, at a point of time when a slight rotary motion of vane rotor 9 relative to sprocket 1 in the phase-advance direction (see the direction indicated by the arrow in FIG. 8) from the angular position of FIG. 8 to the angular position of FIG. 9 occurs owing to the negative torque of alternating torque acting on camshaft 2, a pulse current output from controller 35 to the electromagnetic coil of electromagnetic directional control valve 41 is stopped, and thus a working-fluid supply from oil pump 40 to each of unlocking pressure-receiving chambers 32-34 is also stopped.

Thus, as seen in FIG. 9, the tip 27a of first lock pin 27 is kept in abutted-engagement with the inner face 1c of sprocket 1 under preload (by the spring force of first spring 36), and the tip 29a of third lock pin 29 is brought into abutted-engagement with the first bottom face 26a of third lock hole 26 by the spring force of third spring 38. At this time, even when vane rotor 9 tends to rotate relative to sprocket 1 in the opposite direction (i.e., in the phase-retard direction) owing to the positive torque of alternating torque acting on camshaft 2, such a rotary motion of vane rotor 9 in the phase-retard direction (see the direction indicated by the arrow in FIG. 9) can be restricted by abutment of the outer periphery (the edge) of the tip 29a of third lock pin 29 with the upstanding stepped inner face vertically extending from the first bottom face 26a.

Thereafter, when a further rotary motion of vane rotor 9 relative to sprocket 1 in the phase-advance direction occurs owing to the negative torque acting on camshaft 2, as shown in FIGS. 9-10, third lock pin 29 lowers from the first bottom face 29a to the second bottom face 29b stepwise in the phase-advance direction and thus the tip 29a of third lock pin 29 is brought into abutted-engagement with the second bottom face 26b. Then, by virtue of the ratchet action, the tip 29a of third lock pin 29 moves along the second bottom face 26b in the phase-advance direction, and then reaches a substantially midpoint of the second bottom face 26b. At this time, as shown in FIG. 11, the tip 28a of second lock pin 28 slides into abutted-engagement with the first bottom face 25a of second lock hole 25 by the spring force of second spring 37. Thereafter, when vane rotor 9 further rotates in the phase-advance direction, as shown in FIGS. 11-12, the tip 29a of third lock pin 29 moves to the vicinity of the upstanding inner face 26c of third lock hole 26. At the same time, the tip 28a of second lock pin 28 is brought into abutted-engagement with the second bottom face 25b by virtue of the ratchet action.

When vane rotor 9 still further rotates in the phase-advance direction owing to the negative torque, as shown in FIGS. 12-13, the tip 27a of first lock pin 27 slides into engagement with the first lock hole 24, while second and third lock pins 28-29 slide in the same direction. Under these conditions, as previously discussed, the circumferentially-opposed outer peripheral edges of first and second lock pins 27-28, circumferentially opposed to each other, abut with the circumferentially-opposed upstanding inner faces 24b and 25c of first and second lock holes 24-25, respectively, such that the specified area of the inner face 1c of sprocket 1, ranging between the two upstanding inner faces 24b and 25c, is sandwiched with the two lock pins 27-28. Hence, vane rotor 9 can be stably surely held or locked at the intermediate-phase angular position (see FIG. 6) between the maximum phase-retard angular position and the maximum phase-advance angular position.

Thereafter, immediately after the ignition switch has been turned ON to start up the engine, due to initial explosion (the start of cranking) oil pump 40 begins to operate. Thus, the discharge pressure of working fluid discharged from oil pump 40 is delivered to each phase-retard chamber 11 and each phase-advance chamber 12 via respective passages 18 and 19. On the other hand, the lock passage 20 is kept in a fluid-communication relationship with the drain passage 43. Thus, first, second, and third lock pins 27-29 are kept in engagement with respective lock holes 24-26 by the spring forces of first, second, and third springs 36-38.

As previously discussed, the axial position of the valve spool of electromagnetic directional control valve 41 is controlled by means of controller 35 depending on latest up-to-date information about the detected engine operating condition and the detected pump discharge pressure. Hence, with the engine at an idle rpm, at which the discharge pressure of working fluid discharged from oil pump 40 is unstable, the engaged states (locked states) of first, second, and third lock pins 27-29 are maintained.

After this, immediately before the engine operating condition shifts from the idling condition to a low-speed low-load operating range or a high-speed high-load operating range, a control current is outputted from controller 35 to the electromagnetic coil of electromagnetic directional control valve 41. Thus, the valve spool is slightly displaced against the spring force of the valve spring. The axial position of the valve spool, slightly displaced from the “first position” (the spring-offset position) is referred to as “sixth position”. With the valve spool held at the “sixth position”, fluid communication between the discharge passage 40a and the lock passage 20 becomes established. On the other hand, both of the phase-retard passage 18 and the phase-advance passage 19 remain kept in a fluid-communication relationship with the discharge passage 40a.

Therefore, working fluid can be supplied via the fluid-passage portion 20a of lock passage 20 to each of first, second, and third unlocking pressure-receiving chambers 32-34. Hence, movement of the tip 27a of first lock pin 27 out of engagement with the first lock hole 24 against the spring force of first spring 36, movement of the tip 28a of second lock pin 28 out of engagement with the second lock hole 25 against the spring force of second spring 37, and movement of the tip 29a of third lock pin 29 out of engagement with the third lock hole 26 against the spring force of third spring 38 simultaneously occur. Thus, free rotation of vane rotor 9 relative to sprocket 1 in the normal-rotational direction or in the reverse-rotational direction can be permitted. At the same time, working fluid is supplied to both of the phase-retard chamber 11 and the phase-advance chamber 12.

Hereupon, assume that working-fluid pressure is merely delivered to either one of phase-retard chamber 11 and phase-advance chamber 12. In such a case, a rotary motion of vane rotor 9 relative to sprocket 1 in either one of the phase-retard direction and the phase-advance direction occurs, and hence the first lock pin 27 has to receive a shearing force caused by a circumferential displacement of the first lock-pin hole 31a of rotor 15 relative to the first lock hole 24. In a similar manner, the second lock pin 28 has to receive a shearing force caused by a circumferential displacement of the second lock-pin hole 31b of rotor 15 relative to the second lock hole 25. In a similar manner, the third lock pin 29 has to receive a shearing force caused by a circumferential displacement of the second lock-pin hole 31c of rotor 15 relative to the second lock hole 26. As a result of this, the first lock pin 27 is brought into a so-called jammed (bitten) condition between the first lock-pin hole 31a and the first lock hole 24 displaced relatively. The second lock pin 28 is also brought into a so-called jammed (bitten) condition between the second lock-pin hole 31b and the second lock hole 25 displaced relatively. The third lock pin 29 is also brought into a so-called jammed (bitten) condition between the third lock-pin hole 31c and the third lock hole 26 displaced relatively. Hence, there is a possibility that the locked (engaged) state of lock pins 27-29 with respective lock holes 24-26 cannot be easily released.

Also, assume that there is no hydraulic-pressure supply to both of the phase-retard chamber 11 and the phase-advance chamber 12. In such a case, owing to alternating torque transmitted from the camshaft 2, vane rotor 9 tends to flutter, and thus vane rotor 9 (especially, the first vane 16a) is brought into collision-contact with the shoe 10a of housing body 10, and whereby there is an increased tendency for hammering noise to occur.

In contrast to the above, according to the valve timing control apparatus of the embodiment, working-fluid pressure (hydraulic pressure) can be simultaneously supplied to both of the phase-retard chamber 11 and the phase-advance chamber 12. Thus, it is possible to adequately suppress vane rotor 9 from fluttering and also to adequately suppress the jammed (bitten) condition of the first lock pin 27 between the first lock-pin hole 31a and the first lock hole 24, the jammed (bitten) condition of the second lock pin 28 between the second lock-pin hole 31b and the second lock hole 25, and the jammed (bitten) condition of the third lock pin 29 between the third lock-pin hole 31c and the third lock hole 26.

Thereafter, when the engine operating condition has been shifted to a low-speed low-load operating range, the valve spool is further displaced against the spring force of the valve spring by energizing the solenoid with a further increase in electric current flowing through the electromagnetic coil of electromagnetic directional control valve 41, and thus positioned at the “third position”. Both of the lock passage 20 and the phase-retard passage 18 remain kept in a fluid-communication relationship with the discharge passage 40a. Fluid-communication between the phase-advance passage 19 and the drain passage 43 becomes established.

As a result of this, first, second, and third lock pins 27-29 become kept out of engagement with respective lock holes 24-26. Also, working fluid in phase-advance chamber 12 is drained through the drain passage 43 and thus hydraulic pressure in phase-advance chamber 12 becomes low, whereas working fluid is delivered via the discharge passage 40a to the phase-retard chamber 11 and thus hydraulic pressure in phase-retard chamber 11 becomes high.

Accordingly, vane rotor 9 rotates relative to the housing 7 (i.e., sprocket 1) toward the maximum phase-retard angular position.

Accordingly, a valve overlap of open periods of intake and exhaust valves becomes small and thus the amount of in-cylinder residual gas also reduces, thereby enhancing a combustion efficiency and consequently ensuring stable engine revolutions and improved fuel economy.

Thereafter, when the engine operating condition has been shifted to a high-speed high-load operating range, the valve spool is displaced by energizing the solenoid of electromagnetic directional control valve 41 with a small amount of control current flowing through the electromagnetic coil, and thus positioned at the “second position”. As a result, fluid-communication between the phase-retard passage 18 and the drain passage 43 becomes established. The lock passage 20 remains kept in a fluid-communication relationship with the discharge passage 40a. At the same time, fluid-communication between the phase-advance passage 19 and the discharge passage 40a becomes established.

Therefore, first, second, and third lock pins 27-29 are kept out of engagement with respective lock holes 24-26. Also, working fluid in phase-retard chamber 11 is drained through the drain passage 43 and thus hydraulic pressure in phase-retard chamber 11 becomes low, whereas working fluid is delivered via the discharge passage 40a to the phase-advance chamber 12 and thus hydraulic pressure in phase-advance chamber 12 becomes high. Accordingly, vane rotor 9 rotates relative to the housing 7 (i.e., sprocket 1) toward the maximum phase-advance angular position (see FIG. 7). Thus, the angular phase of camshaft 2 relative to sprocket 1 is converted into the maximum advanced relative-rotation phase.

Accordingly, a valve overlap of open periods of intake and exhaust valves becomes large and thus the intake-air charging efficiency is increased, thereby improving engine torque output.

Conversely when the engine operating condition shifts from the low-speed low-load operating range or the high-speed high-load operating range to the idling condition, a supply of control current from controller 35 to the electromagnetic coil of electromagnetic directional control valve 41 is stopped and thus the solenoid is de-energized. Thus, the valve spool is positioned at the “first position” (i.e., the spring-offset position) shown in FIG. 1 by the spring force of the valve spring. The lock passage 20 communicates with the drain passage 43, whereas the discharge passage 40a communicates with both of the phase-retard passage 18 and the phase-advance passage 19. Accordingly, hydraulic pressures having almost the same pressure value are applied to respective hydraulic chambers (phase-retard chamber 11 and phase-advance chamber 12).

For the reasons discussed above, even when vane rotor 9 has been positioned at a phase-retard angular position, rotary motion of vane rotor 9 relative to sprocket 1 in the phase-advance direction occurs owing to alternating torque acting on camshaft 2. Hence, by the spring force of first spring 36 and by virtue of the ratchet action of the first lock guide groove (bottom face 24a), first lock pin 27 is brought into engagement with the bottom face 24a of first lock hole 24, owing to rotary motion of vane rotor 9 in the phase-advance direction. In a similar manner, by the spring force of second spring 37 and by virtue of the ratchet action of the second lock guide stepped groove (bottom faces 25a-25b), second lock pin 28 is brought into engagement with the first and second bottom faces 25a-25b of second lock hole 25, one-by-one, owing to rotary motion of vane rotor 9 in the phase-advance direction. Also, by the spring force of third spring 38 and by virtue of the ratchet action of the third lock guide stepped groove (bottom faces 26a-26b), third lock pin 29 is brought into engagement with the first and second bottom faces 26a-26b of third lock hole 26, one-by-one, owing to rotary motion of vane rotor 9 in the phase-advance direction. Hence, the angular position of vane rotor 9 relative to sprocket 1 is held or locked at the intermediate-phase angular position (see FIG. 6) between the maximum phase-retard angular position and the maximum phase-advance angular position.

Also, when manually stopping the engine, the ignition switch is turned OFF. As previously described, first, second, and third lock pins 27-29 are maintained in their locked states where the tip 27a of first lock pin 27 has been engaged with the bottom face 24a of first lock hole 24, the tip 28a of second lock pin 28 has been engaged with the second bottom face 25b of second lock hole 25, and the tip 29a of third lock pin 29 has been engaged with the second bottom face 26b of third lock hole 26.

Furthermore, assume that the engine is operating continuously in a given engine operating range, the electromagnetic coil of the solenoid of electromagnetic directional control valve 41 is energized with a given amount of control current, and thus the valve spool is positioned at a substantially intermediate axial position, that is, the “fourth position”. As a result, fluid communication between the phase-advance passage 19 and the discharge passage 40a is blocked and fluid communication between the phase-retard passage 18 and the drain passage 43 is blocked. On the other hand, fluid communication between the discharge passage 40a and the lock passage 20 is established. Hence, hydraulic pressure of working fluid in each of phase-retard chambers 11 and hydraulic pressure of working fluid in each of phase-advance chambers 12 are held constant. Also, by the hydraulic-pressure supply from the discharge passage 40a to the lock passage 20, first, second, and third lock pins 27-29 are kept out of engagement with respective lock holes 24-26, that is, held in their unlocked states.

Therefore, the angular position of vane rotor 9 relative to sprocket 1 is held at a desired angular position corresponding to the given amount of control current, and thus the angular phase of camshaft 2 relative to sprocket 1 (i.e., housing 7) is held at a desired relative-rotation phase. Accordingly, intake valve open timing (IVO) and intake valve closure timing (IVC) can be held at respective desired timing values.

In this manner, by energizing the solenoid of electromagnetic directional control valve 41 with a desired amount of control current or de-energizing the solenoid, by means of controller 35 depending on latest up-to-date information about an engine operating condition, and thus controlling axial movement of the valve spool, the axial position of the valve spool can be controlled to either one of the first, second, third, and fourth positions. As discussed above, the angular phase of camshaft 2 relative to sprocket 1 (i.e., housing 7) can be adjusted or controlled to a desired relative-rotation phase (an optimal relative-rotation phase) by controlling both of the phase-change mechanism 3 and the lock mechanism 4, thus more certainly enhancing the control accuracy of valve timing control.

[Operation During a Restarting Period after an Engine Stall, Occurred During Low-Temperature Engine Operating Condition]

For instance, assume that the engine has stalled with the vane rotor 9 positioned nearer the phase-retard side than the intermediate lock position (i.e., near the maximum phase-retard position) after cold-start operation and thus the engine has stopped abnormally. In such a case, the engine is cranked for restarting by turning the ignition switch ON. At this point of time, working fluid is supplied to both of the phase-retard chamber 11 and the phase-advance chamber 12, but a viscous resistance of working fluid in phase-retard hydraulic chambers and phase-advance hydraulic chambers tends to reduce a fluttering motion of vane rotor 9, occurring due to positive and negative alternating torque. The reduced fluttering motion results in an undesirable increase in recovery time of vane rotor 9 to the intermediate-phase angular position (i.e., the intermediate lock position) suited for starting.

In contrast, according to the VTC apparatus of the embodiment employing recessed-groove passages 50, as shown in FIGS. 4 and 5A, phase-retard chamber 11 and phase-advance chamber 12 are kept in a communicating state through the associated recessed-groove passage 50. When vane rotor 9 is momentarily rotated to the phase-advance side by negative alternating torque created at the beginning of cranking, replaced-flow of working fluid from phase-retard chamber 11 to phase-advance chamber 12 through the associated recessed-groove passage 50 takes place by virtue of the torque acting on the rotor. Also, as previously discussed, the radial length as well as the circumferential length L of recessed-groove passage 50, dimensioned to be greater than the depth D, contributes to an increase in communication-passage opening area for replaced-flow of working fluid between the adjacent hydraulic chambers (i.e., phase-retard chamber 11 and phase-advance chamber 12) through the recessed-groove passage 50. The increased opening area for replaced-flow also contributes to a reduced flow resistance to the replaced working-fluid flow, in other words, an increase in replaced-flow velocity, thus effectively increasing the fluttering motion of vane rotor 9.

Thus, as shown in FIG. 5B, vane rotor 9 can greatly rapidly rotate in the phase-advance direction due to the negative torque created at the beginning of cranking and by virtue of the fluttering motion (or the fluttering angle) of vane rotor 9, increased or amplified by fluid-communication between phase-retard chamber 11 and phase-advance chamber 12 by way of recessed-groove passage 50.

When vane rotary motion of rotor 9 relative to housing 7 in the phase-advance direction reaches a given amount, as shown in FIG. 5B, the clockwise end 50b of recessed-groove passage 50 is closed by the front end face (the upper end face, viewing FIG. 5B) of the first vane 16a, facing the inside end face of front plate 13. At this point of time, replaced-flow of working fluid from phase-retard chamber 11 to phase-advance chamber 12 through the associated recessed-groove passage 50 is blocked.

After this, vane rotor 9 rotates to the intermediate-phase angular position by virtue of the previously-described ratchet action. Hence, it is possible to shorten the recovery time of vane rotor 9 to the initial position (i.e., the intermediate lock position) during the cranking period, thus enhancing the startability of the engine.

Also, during the previously-discussed engine-stalling condition, a supply of control current to the electromagnetic coil of electromagnetic directional control valve 41 is cut off. The term “cut-off” of the electric-current supply includes the other factors, for example, breaking of the electromagnetic coil, or a disabling state of switching among the ports, that is, a disabling state of a change in the path of flow through the electromagnetic directional control valve 41 due to the spool stuck due to contamination, dirt or debris (e.g., a very small piece of metal) contained in working fluid during sliding movement of the spool and jammed between the edge of each of land portions of the spool and the edge of each of the ports. In the presence of working-fluid supply to both of the phase-retard chamber 11 and the phase-advance chamber 12 under the “cut-off” state of the electric-current supply due to the other factors as previously discussed, replaced-flow of working fluid from phase-retard chamber 11 to phase-advance chamber 12 through the associated recessed-groove passage 50 occurs during a restarting period of the engine even when vane rotor 9 is positioned at the maximum phase-retard position. The replaced working-fluid flow ensures a smooth, rapid rotary motion of vane rotor 9 in the phase-advance direction.

[Automatic-Engine-Stop]

When the engine is automatically stopped by means of an idling-stop system, in a similar manner to the previously-discussed manual-engine-stop operation, during idling before the engine automatically stops, electromagnetic directional control valve 41 is still energized by the controller 35, so that the valve spool of electromagnetic directional control valve 41 is positioned at the “third position”. Fluid-communication between the phase-retard passage 18 and the discharge passage 40a is established, while fluid-communication between the phase-advance passage 19 and the drain passage 43 is established. At the same time, fluid-communication between the lock passage 20 and the discharge passage 40a is established. Therefore, first, second, and third lock pins 27-29 are kept at their retracted positions under hydraulic pressure. Working fluid is delivered via the discharge passage 40a to the phase-retard chamber 11 and thus hydraulic pressure in phase-retard chamber 11 becomes high, whereas working fluid in phase-advance chamber 12 is drained through the drain passage 43 and thus hydraulic pressure in phase-advance chamber 12 becomes low. Hence, vane rotor 9 becomes placed into the maximum phase-retard angular position shown in FIG. 3.

Immediately when vane rotor 9 reaches the maximum phase-retard angular position shown in FIG. 3, a pulse current output from controller 35 to the electromagnetic coil of electromagnetic directional control valve 41 becomes stopped, and thus the valve spool of electromagnetic directional control valve 41 becomes positioned at the “first position” (i.e., the spring-offset position) shown in FIG. 1, so that the lock passage 20 communicates with the drain passage 43. At this time, there is no supply of working fluid from oil pump 40 to each of unlocking pressure-receiving chambers 32-34, and thus first, second, and third lock pins 27-29 are forced in their extending directions by the biasing forces of first, second, and third springs 36-38. As a result, as seen in FIG. 8, the second and third lock pins 28-29 are kept out of engagement with the respective lock holes 25-26 but kept in abutted-engagement with the inner face 1c of sprocket 1 under preload (by the biasing forces of second and third springs 37-38). On the other hand, the first lock pin 27 is kept in engagement with the second lock hole 25 by the biasing force of first spring 36.

Hence, vane rotor 9 can be stably surely held or locked at the maximum phase-retard angular position (see FIG. 3). Thereafter, when automatically restarting the engine, that is, at the beginning of cranking, the engine can be restarted at intake-valve timing corresponding to the maximum retarded phase. This contributes to the appropriately reduced effective compression ratio, thereby adequately suppressing noise and vibrations of the engine, while ensuring a good startability.

By the way, after the engine has been automatically restarted, in the same manner as previously discussed, electromagnetic directional control valve 41 becomes energized. Depending on the axial position of the sliding spool, fluid-communication between the discharge passage 40a and the lock passage 20 becomes established. Thus, movement of the tip 27a of first lock pin 27 out of engagement with the second lock hole 25 against the spring force of first spring 36 occurs. Thus, free rotation of vane rotor 9 relative to sprocket 1 in the normal-rotational direction or in the reverse-rotational direction can be permitted.

As discussed above, in the valve timing control apparatus of the embodiment, working fluid in phase-retard chamber 11 flows rapidly through the associated recessed-groove passage 50 into the phase-advance chamber 12 during a restarting period after an engine stall during a low-temperature engine starting operation. Therefore, vane rotor 9, which is positioned at the maximum phase-retard position, can be rapidly rotated to the intermediate-phase angular position (i.e., the intermediate lock position) suited for starting, thus ensuring a good restartability.

Also, in the valve timing control apparatus of the embodiment, first, second, and third lock pins 27-29 are installed in the rotor 15 of vane rotor 9 via respective lock-pin holes 31a-31c, without installing in the vanes 16a-16d of vane rotor 9. Thus, it is possible to adequately reduce a circumferential thickness of each of vanes 16a-16d, thereby adequately enlarging a relative-rotation angle of vane rotor 9 relative to housing 7. Also, this contributes to a more compact VTC apparatus.

Hitherto, in order to retain or hold lock pins, the rotor diameter of a vane rotor (a vane member) in itself had to be expanded. In contrast, in the apparatus of the embodiment, the rotor 15 of vane rotor 9 has partly-expanded, circumferentially-spaced large-diameter portions 15e-15f without expanding the entire circumference of rotor 15, and three lock pins 27-29 are installed in the partly-expanded large-diameter portions 15e-15f of rotor 15. By virtue of the different-diameter deformed outer peripheral surface of rotor 15, the total volumetric capacity of hydraulic chambers 11a and 12a, located in the area corresponding to the small-diameter portion (each of first and second small-diameter portions 15c-15d) of rotor 15, is set to be greater than the total volumetric capacity of hydraulic chambers 11b and 12b, located in the area corresponding to the large-diameter portion (each of first and second large-diameter portions 15e-15f).

Thus, the pressure-receiving surface area of each of side faces 16e-16h of vanes 16a-16d, facing hydraulic chambers 11a and 12a located in the area corresponding to the small-diameter portion (each of first and second small-diameter portions 15c-15d), is set to be adequately greater than that of each of side faces of vanes 16a-16d, facing hydraulic chambers 11b and 12b located in the area corresponding to the large-diameter portion (each of first and second large-diameter portions 15e-15f). Hence, during valve timing control, a relative-rotation speed of vane rotor 9 to housing 7 can be increased, thereby adequately enhancing a conversion responsiveness of the relative-rotation phase of camshaft 2 to housing 7 (the crankshaft) and satisfactorily improving a responsiveness of intake-valve timing control.

Furthermore, two small-diameter portions 15c-15d are arranged at angular positions circumferentially spaced apart from each other and diametrically opposed to each other (concretely, by approximately 180 degrees), whereas two large-diameter portions 15e-15f are arranged at angular positions circumferentially spaced apart from each other and diametrically opposed to each other (concretely, by approximately 180 degrees). As a whole, the weight of vane rotor 9 can be circumferentially balanced and uniformed, thereby avoiding rotational unbalance of vane rotor 9. This ensures a smooth rotary motion of vane rotor 9 relative to housing 7.

Additionally, in the embodiment, when the engine is automatically stopped, vane rotor 9 can be locked or held at the maximum phase-retard angular position mechanically by means of the lock mechanism 4, rather than hydraulically. This eliminates the necessity of a separate hydraulic pressure source for holding the vane rotor at the maximum phase-retard angular position. This contributes to more simplified VTC apparatus and reduced system costs.

Additionally, in the embodiment, a function of hydraulic-pressure control for each of the hydraulic pressure chambers (phase-retard chamber 11 and phase-advance chamber 12) and a function of hydraulic-pressure control for each of first, second, and third unlocking pressure-receiving chambers 32-34 are both achieved by means of the single electromagnetic directional control valve 41. Thus, it is possible to enhance the flexibility of layout of the VTC system on the engine body, thus ensuring lower system installation time and costs.

Furthermore, it is possible to enhance the ability to hold the angular position of vane rotor 9 relative to sprocket 1 at the intermediate-phase angular position by means of the lock mechanism 4, when the engine is manually stopped. Additionally, by virtue of the second lock guide groove (the two-stage stepped lock guide groove with two bottom faces 25a-25b, serving as a one-way clutch, in other words, a ratchet) and the third lock guide groove (the two-stage stepped lock guide groove with two bottom faces 26a-26b, serving as a one-way clutch, in other words, a ratchet), movement of second lock pin 28 only into engagement with the second lock hole 25 and movement of third lock pin 29 only into engagement with the third lock hole 26 are permitted, thus assuring more safe and certain guiding action for movement of lock pins 28-29 into engagement.

Even when vane rotor 9 tends to rotate relative to sprocket 1 in the phase-retard direction owing to the positive torque, it is possible to safely certainly guide the vane rotor 9 toward the intermediate-phase angular position by virtue of a long four-stage ratchet action, created by two bottom faces 25a-25b of second lock hole 25, and two bottom faces 26a-26b of third lock hole 26.

Hydraulic pressure in each of phase-retard chamber 11 and phase-advance chamber 12 is not used as hydraulic pressure acting on each of first, second, and third unlocking pressure-receiving chambers 32-34. In comparison with a system that hydraulic pressure in each of phase-retard chamber 11 and phase-advance chamber 12 is also used as hydraulic pressure acting on each of unlocking pressure-receiving chambers, a responsiveness of the hydraulic system of the embodiment to hydraulic pressure supply to each of unlocking pressure-receiving chambers 32-34 can be greatly improved. Thus, it is possible to improve a responsiveness of each of lock pins 27-29 to backward movement for unlocking (disengaging). Also, the hydraulic system of the embodiment, in which hydraulic pressure can be supplied to each of unlocking pressure-receiving chambers 32-34 without using hydraulic pressure in each of phase-retard chamber 11 and phase-advance chamber 12, more concretely, the single electromagnetic directional control valve 41 eliminates the need for a fluid-tight sealing device between each of phase-retard chamber 11 and phase-advance chamber 12 and each of unlocking pressure-receiving chambers 32-34.

In addition to the above, in the shown embodiment, lock mechanism 4 is comprised of three separate lock devices, that is, (i) the first lock pin 27 and the first lock guide groove with bottom face 24a (ii) the second lock pin 28 and the second lock guide groove (the two-stage stepped groove) with first and second bottom faces 25a-25b, and (iii) the third lock pin 29 and the third lock guide groove (the two-stage stepped groove) with first and second bottom faces 26a-26b. Hence, it is possible to reduce the wall thickness of sprocket 1 in which each of lock holes 24-26 is formed. In more detail, for instance, assume that the lock mechanism is constructed by a single lock pin and a single lock guide groove (a single multi-stage stepped groove). In such a case, five bottom faces have to be formed in the sprocket in a manner so as to continuously lower stepwise from the phase-retard side to the phase-advance side. As a matter of course, to provide the five-stage stepped groove, the wall thickness of the sprocket also has to be increased. In contrast, the embodiment adopts three separate lock devices (27, 24a; 28, 25a-25b; 29, 26a-26b) as the lock mechanism, and hence it is possible to reduce the thickness of sprocket 1, thereby shortening the axial length of the VTC apparatus and consequently enhancing the flexibility of layout of the VTC system on the engine body.

Additionally, in the shown embodiment, in order to more certainly enhance the restartability during low-temperature engine operating condition, the circumferential length of the stepped bottom face of second lock hole 25 is set to be less than or equal to an angle of fluttering motion of vane rotor 9, oscillating by positive and negative alternating torque acting on camshaft 2 due to spring forces of valve springs. Also, controller 35 is configured to set a phase-angle range of vane rotor 9 relative to housing 7, to be controlled by electromagnetic directional control valve 41 after the engine has started (restarted), to a phase-angle range corresponding to the non-communicating state where fluid-communication between phase-retard hydraulic chamber 11 and phase-advance hydraulic chamber 12 by way of the recessed-groove passage 50 is blocked. Furthermore, electromagnetic directional control valve 41 is held at its initial valve position (i.e., the spring-loaded position), at which working fluid is supplied to both the phase-advance chamber 12 and the phase-retard chamber 11, in a non-controlled state where electromagnetic directional control valve 41 is not controlled by the controller 35.

Second Embodiment

Referring now to FIGS. 14A-14B, there is shown the partial cross section of the VTC apparatus of the second embodiment. The VTC apparatus of the second embodiment shown in FIGS. 14A-14B differs from the first embodiment shown in FIGS. 1-13, in that the position of formation of four recessed-groove passages 51 (serving as fluid-communication passages), through which phase-retard chamber 11 and phase-advance chamber 12 are communicated with each other at the maximum phase-retard position of vane rotor 9, is somewhat modified. Actually, in the second embodiment, recessed-groove passages 51 are formed in the inner face 1c of sprocket 1 rather than the inside end face of front plate 13.

The circumferential length L, the depth D, the position of formation (in the circumferential direction and in the radial direction) of each of recessed-groove passages 51 of the VTC apparatus of the second embodiment of FIGS. 14A-14B is the same as that described for the first embodiment.

Thus, the VTC apparatus of the second embodiment can provide the same operation and effects as the first embodiment. That is, when vane rotor 9 has to be momentarily rotated to the phase-advance side by negative alternating torque during a restarting period immediately after an engine stall occurred during a low-temperature engine starting operation, replaced-flow of working fluid from phase-retard chamber 11 to phase-advance chamber 12 through the associated recessed-groove passage 51 takes place. Hence, it is possible to increase a speed of recovery of vane rotor 9 to the initial position (i.e., the intermediate lock position), thus enhancing the startability of the engine.

Third Embodiment

Referring now to FIG. 15, there is shown the front elevation, viewed from the front-plate side of the VTC apparatus of third embodiment. In the VTC apparatus of the third embodiment shown in FIG. 15, in a similar manner to the first embodiment, four recessed-groove passages 50 are configured to be substantially conformable to the maximum phase-retard positions of four vanes 16a-16d. In addition to the four recessed-groove passages 50, four recessed-groove passages 52 are formed in the inside end face of front plate 13 and configured to be substantially conformable to the maximum phase-advance positions of four vanes 16a-16d. Each of recessed-groove passages 50, configured to be substantially conformable to the maximum phase-retard positions of vanes 16a-16d, is hereinafter referred to as “first recessed-groove passage”, whereas each of recessed-groove passages 52, configured to be substantially conformable to the maximum phase-advance positions of vanes 16a-16d, is hereinafter referred to as “second recessed-groove passage”.

The circumferential length L of second recessed-groove passage 52 is dimensioned to be identical to that of the first recessed-groove passage 50, and also dimensioned to be slightly greater than the circumferential width W of each of vanes 16a-16d. At the maximum phase-advance position of vane rotor 9, one circumferential end (a clockwise end 52a) of each of recessed-groove passages 52 is formed in a position that the one circumferential end 52a faces the phase-retard chamber 11 and overlaps with the third shoe 10c. The other circumferential end (an anticlockwise end 52b) of each of recessed-groove passages 52 is formed in a position that the other circumferential end 52b faces the phase-advance chamber 12. At this point of time when vane rotor 9 reaches the maximum phase-advance position, phase-retard chamber 11 and phase-advance chamber 12 are communicated with each other through the associated recessed-groove passage 52.

Hence, when the engine has stopped rotating owing to engine stalling and additionally vane rotor 9 has been held at its maximum phase-advance position, a rotary motion of vane rotor 9 relative to housing 7 in the phase-retard direction (anticlockwise, viewing FIG. 15) occurs owing to the positive alternating torque created at the beginning of cranking operation for restarting. At this time, replaced-flow of working fluid from phase-advance chamber 12 to phase-retard chamber 11 through the associated recessed-groove passage 52 occurs. By virtue of the fluttering motion (the fluttering angle) of vane rotor 9, increased or amplified by fluid-communication between phase-retard chamber 11 and phase-advance chamber 12 by way of recessed-groove passage 52, vane rotor 9, which is positioned at the maximum phase-advance position, can be rapidly rotated to the intermediate-phase angular position (i.e., the intermediate lock position) suited for starting, thus ensuring a good restartability. By the way, the VTC apparatus of the third embodiment has the first recessed-groove passage 50 as well as the second recessed-groove passage 52. In a similar manner to the first embodiment, when the engine has stopped rotating owing to engine stalling and additionally vane rotor 9 has been held at its maximum phase-retard position, a rotary motion of vane rotor 9 relative to housing 7 in the phase-advance direction (clockwise, viewing FIG. 15) occurs owing to the negative alternating torque created at the beginning of cranking operation for restarting. Thus, replaced-flow of working fluid from phase-retard chamber 11 to phase-advance chamber 12 through the associated recessed-groove passage 50 occurs, thus increasing or amplifying the fluttering motion (the fluttering angle) of vane rotor 9. As a result, the VTC apparatus of the third embodiment can provide the same operation and effects as the first embodiment

Fourth Embodiment

Referring now to FIGS. 16A-16B, there is shown the partial cross section of the VTC apparatus of the fourth embodiment. In the VTC apparatus of the fourth embodiment shown in FIGS. 16A-16B, in a similar manner to the first embodiment, four recessed-groove passages 50 (hereinafter are referred to as “first recessed-groove passages”) are formed in the inside end face of front plate 13 and configured to be substantially conformable to the maximum phase-retard positions of four vanes 16a-16d. In addition to the four recessed-groove passages 50, four recessed-groove passages 53 (hereinafter are referred to as “second recessed-groove passages”) are formed in the inside end face (inner face 1c) of sprocket 1 and configured to be opposed to respective first recessed-groove passages 50.

Therefore, according to the VTC apparatus of the fourth embodiment, by the provision of second recessed-groove passages 53 as well as first recessed-groove passages 50, the total fluid-flow passage cross-sectional area of the fluid-communication passages, through which phase-retard chamber 11 and phase-advance chamber 12 are communicated with each other at the maximum phase-retard position of vane rotor 9, can be enlarged. This contributes to the reduced flow resistance to working-fluid flow from phase-retard chamber 11 to phase-advance chamber 12. This means a further rise in replaced-flow velocity of working fluid from phase-retard chamber 11 to phase-advance chamber 12. By virtue of the further risen replaced-flow velocity, vane rotor 9 can more rapidly rotate relative to the housing 7 toward the side of phase-advance chamber 12 (i.e., in the phase-advance direction) by alternating torque (in particular, negative alternating torque), thus ensuring a better restartability of the engine.

Fifth Embodiment

Referring now to FIG. 17, there is shown the partial cross section of the VTC apparatus of the fifth embodiment. In the fifth embodiment, as clearly seen from the partial cross section of FIG. 17, each of recessed-groove passages 50 has a circular-arc shape in cross section taken along the circumferential line C-C as shown in FIG. 4. That is, the bottom face of each of recessed-groove passages 50 is configured as a curved bottom face 50c whose depth is dimensioned to gradually shallow from the central deepest portion to the one circumferential end 50a and also dimensioned to gradually shallow from the central deepest portion to the other circumferential end 50b.

The cross-section of recessed-groove passage 50 is formed into a circular-arc shape, and hence working-fluid flow can be smoothly guided from phase-retard chamber 11 into the recessed-groove passage 50, and then smoothly flow into the phase-advance chamber 12. That is, it is possible to reduce a flow resistance to replaced flow of working fluid from phase-retard chamber 11 to phase-advance chamber 12 by way of the associated recessed-groove passage 50 of the circular-arc shaped bottom face 51c. The reduced flow resistance contributes to a further rise in replaced-flow velocity of working fluid between phase-retard chamber 11 and phase-advance chamber 12. By virtue of the further risen replaced-flow velocity, a rotational speed of vane rotor 9 toward the intermediate-phase angular position (i.e., the intermediate lock position) can be effectively increased, thus ensuring a further enhanced restartability.

In the fifth embodiment, recessed-groove passage 50 has a circular-arc cross-sectional form throughout its circumferential length L. In lieu thereof, at least the two circumferential ends 50a-50b of recessed-groove passage 50 may be partially formed into a circular-arc shape, for the purpose of ensuring a rise in replaced-flow velocity of working fluid.

By the way, the circular-arc cross-sectional form of recessed-groove passage 50, described for the fifth embodiment (see FIG. 17), can be applied to each of recessed-groove passage 51 of the second embodiment (see FIGS. 14A-14B), recessed-groove passage 52 of the third embodiment (see FIG. 15), and recessed-groove passage 53 of the fourth embodiment (see FIGS. 16A-16B).

It will be appreciated that the invention is not limited to the particular embodiments shown and described herein, but that various changes and modifications may be made. For instance, the cross-sectional form, the depth D, and the circumferential length L of each of recessed-groove passages 50-53 of the shown embodiments may be arbitrarily changed depending on the size/specification of the VTC apparatus.

The valve timing control (VTC) apparatus of the shown embodiments is exemplified in the phase control apparatus applied to an intake-valve side of an internal combustion engine. In lieu thereof, the VTC apparatus may be used for a phase control apparatus installed on an exhaust-valve side. The previously-discussed fundamental inventive concept may be applied to all types of hydraulically-operated vane rotor equipped variable valve timing control (VTC) device.

The entire contents of Japanese Patent Application No. 2011-269495 (filed Dec. 9, 2011) are incorporated herein by reference.

While the foregoing is a description of the preferred embodiments carried out the invention, it will be understood that the invention is not limited to the particular embodiments shown and described herein, but that various changes and modifications may be made without departing from the scope or spirit of this invention as defined by the following claims.

VALVE TIMING CONTROL APPARATUS OF INTERNAL COMBUSTION ENGINE

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