VARIABLE VALVE TIMING CONTROL APPARATUS OF INTERNAL COMBUSTION ENGINE AND COOLING DEVICE FOR THE SAME

TECHNICAL FIELD

The present invention relates to a variable valve timing control apparatus of an internal combustion engine, configured to variably control engine valve timings (intake-valve open- and closure-timing or exhaust-valve open- and closure-timing) by means of an electromagnetic brake such as a hysteresis brake, and specifically to a cooling device configured to cool the inside of the valve timing control apparatus by way of cooling fluid such as cooling oil.

BACKGROUND ART

In recent years, there have been proposed and developed various electromagnetic-brake equipped variable valve timing control (VTC) systems. One such electromagnetic-brake equipped VTC system has been disclosed in Japanese Patent Provisional Publication No. 2004-239231 (hereinafter is referred to as “JP2004-239231”). In the electromagnetic-brake equipped VTC system disclosed in JP2004-239231, a phase change mechanism is interposed between a drive ring of the crankshaft side and a driven shaft member of the camshaft side, for changing a relative angular phase between the drive ring and the driven shaft member. The phase change mechanism is driven by means of a spiral spring and a hysteresis brake, which is used as an electromagnetic brake. Lubricating oil is used as cooling oil, for removing heat from the hysteresis brake. Lubricating oil is supplied into internal spaces defined between a hysteresis ring and each of inner-stator and outer-stator polar teeth of the hysteresis brake. In the VTC system disclosed in JP2004-239231, a temperature-sensitive valve, which is arranged in an oil supply passage and constructed by a bimetal element, is used as cooling-oil supply flow rate restriction means. The bimetal-type temperature-sensitive valve is configured to restrict a supply flow rate of lubricating oil (i.e., cooling oil) by decreasing an opening area of an opening part of the oil supply passage according to an oil temperature drop. More concretely, when the oil temperature becomes less than or equal to a predetermined temperature value (see a temperature value T₁in FIG. 10), the temperature-sensitive valve operates to shut off lubricating-oil supply to the hysteresis brake side by closing the opening part of the oil supply passage. Conversely when the oil temperature exceeds the predetermined temperature value, the temperature-sensitive valve is designed to open the opening part of the oil supply passage by way of a bending or distorting or deflecting movement of the bimetal element, so as to control a supply flow rate of cooling oil. Therefore, even when the viscosity of cooling oil becomes high due to a temperature drop in cooling oil, it is possible to prevent an undesirable change (i.e., an undesirable increase) in braking force of the hysteresis brake from occurring.

SUMMARY OF THE INVENTION

However, in the case of the VTC system disclosed in JP2004-239231, the opening part of the oil supply passage is opened or closed directly by way of a deflecting movement of the bimetal element of the bimetal-type temperature-sensitive valve, produced by a temperature change in cooling oil. That is, the bimetal-type temperature-sensitive valve is a so-called bimetal valve whose deflecting movement is reflected directly as a bimetal-valve opening area. Owing to repetitions of such deflecting movements of the bimetal element, produced by oil temperature changes, there is an increased tendency for a rate of change in the opening area of the opening part (i.e., the fluid-flow passage area of the oil supply passage) with respect to a deflecting movement (a displacement) of the bimetal element to be undesirably increased. In other words, owing to an aged deterioration in the bimetal element, a cooling-oil-temperature versus valve-opening-area characteristic of the bimetal-type temperature-sensitive valve would be changed undesiredly to a characteristic almost similar to an oil-temperature dependent ON-OFF switching valve, in which, when the oil temperature exceeds a specific narrow temperature range (see a temperature range from the temperature value T₁to a temperature value T₂in FIG. 10), the valve is kept fully open, and conversely when the oil temperature becomes less than the specific narrow temperature range, the valve is kept fully closed. For the reasons discussed above, when a deflecting movement of the bimetal element in a direction opening of the temperature-sensitive valve begins to occur due to a temperature rise in cooling oil, there is a possibility of an instantaneous increase in the amount of cooling oil passing through the opening part of the oil supply passage. This means that an excessive amount of cooling oil having a comparatively high viscosity is undesirably introduced into the hysteresis brake side of the phase change mechanism. Thus, a braking force of the hysteresis brake would be greatly affected by the excessive cooling oil. Even when lubricating oil of a comparatively low viscosity is used as cooling oil, a drag torque tends to increase for a brief moment, due to the instantaneous inflow of a large amount of oil into the hysteresis brake side. In such a case, valve timing compensation, executed by a controller incorporated in the VTC system, is not enough. Thus, there is a possibility of undesirable engine valve timing fluctuations.

It is, therefore, in view of the previously-described disadvantages of the prior art, an object of the invention to provide a variable valve timing control (VTC) apparatus of an internal combustion engine and a cooling device for the VTC apparatus, which is configured to more accurately control a flow rate of oil (cooling fluid) supplied inside of the VTC apparatus for cooling and lubricating the inside of the VTC apparatus (in particular, a phase control mechanism), depending on an oil temperature change.

In order to accomplish the aforementioned and other objects of the present invention, a variable valve timing control apparatus of an internal combustion engine comprises a driving rotational member adapted to be driven by a crankshaft, a driven rotational member fixedly connected to a camshaft, a phase change mechanism configured to control engine valve timing by changing a relative angular phase between the driving rotational member and the driven rotational member, an oil supply passage configured to supply oil from the camshaft into an inside of the phase change mechanism, and a flow control valve comprising a valve bore communicating the oil supply passage and a valve element, the flow control valve configured to control a flow rate of the oil supplied from the oil supply passage into the inside of the phase change mechanism, by changing an opening area of the valve bore by advancing or retreating the valve element in the valve bore responsively to a temperature of the oil.

According to another aspect of the invention, a variable valve timing control apparatus of an internal combustion engine comprises a driving rotational member adapted to be driven by a crankshaft, a driven rotational member fixedly connected to a camshaft, a phase change mechanism configured to control engine valve timing by changing a relative angular phase between the driving rotational member and the driven rotational member, an oil supply passage configured to supply oil from the camshaft into an inside of the phase change mechanism, and a flow control valve comprising a valve bore communicating the oil supply passage, a valve element configured to control a flow rate of the oil supplied from the oil supply passage into the inside of the phase change mechanism by changing an opening area of the valve bore by way of an advancing/retreating motion of the valve element in the valve bore, and a temperature-sensitive member configured to create the advancing/retreating motion of the valve element responsively to a temperature of the oil.

According to a further aspect of the invention, a variable valve timing control apparatus of an internal combustion engine comprises a driving rotational member adapted to be driven by a crankshaft, a driven rotational member fixedly connected to a camshaft, a link mechanism through which the driving rotational member and the driven rotational member are mechanically linked to each other, an oil supply passage configured to supply oil from the camshaft through an inside of the driven rotational member into the link mechanism, and a flow control valve comprising a valve bore communicating the oil supply passage, a valve element configured to control a flow rate of the oil supplied from the oil supply passage through the inside of the driven rotational member into the link mechanism by changing an opening area of the valve bore by way of an advancing/retreating motion of the valve element in the valve bore, and a temperature-sensitive member configured to create the advancing/retreating motion of the valve element responsively to a temperature of the oil.

According to a still further aspect of the invention, a cooling device for cooling a variable valve timing control apparatus of an internal combustion engine employing a driving rotational member driven by a crankshaft, a driven rotational member fixedly connected to a camshaft, and a phase change mechanism for controlling engine valve timing by changing a relative angular phase between the driving rotational member and the driven rotational member, the cooling device comprising an oil pump configured to be driven by either one of the engine and an electric motor, for discharging working oil, an oil supply passage configured to supply at least cooling oil from the pump through the camshaft and the driven rotational member into the phase change mechanism, and a flow control valve comprising a valve bore communicating the oil supply passage, a valve element configured to control a flow rate of the cooling oil supplied from the oil supply passage through the camshaft and the driven rotational member into the phase change mechanism by changing an opening area of the valve bore by way of an advancing/retreating motion of the valve element in the valve bore, and a temperature-sensitive member configured to create the advancing/retreating motion of the valve element responsively to a temperature of the cooling oil.

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 longitudinal cross-sectional view illustrating an embodiment of a variable valve timing control (VTC) apparatus.

FIG. 2 is an exploded perspective view illustrating the VTC apparatus of the embodiment, as viewed in a certain direction.

FIG. 3 is an exploded perspective view illustrating the VTC apparatus of the embodiment, as viewed in the other direction.

FIG. 4 is a front elevation illustrating a flow-rate control valve (a flow control valve) installed on a driven shaft member of the VTC apparatus of the embodiment.

FIG. 5 is an enlarged perspective view illustrating the detailed configuration of the flow control valve installed on the driven shaft member.

FIG. 6 is a side view illustrating a valve element of the flow control valve in the VTC apparatus of the embodiment.

FIG. 7 is a cross-sectional view taken along the line A-A in FIG. 6.

FIG. 8 is a perspective view illustrating a temperature-sensitive member linked to one axial end of the valve element of the flow control valve shown in FIGS. 6-7.

FIGS. 9A-9C are explanatory views illustrating the operation of the flow control valve, at three different oil temperatures, namely, (a) in an operating state at a very low oil temperature less than or equal to the predetermined temperature value T₁, (b) in an operating state at an oil temperature risen up to the predetermined temperature value T₁, and (c) in an operating state at an oil temperature exceeding a predetermined high temperature value T₃.

FIG. 10 is a comparative characteristic diagram showing both (i) a temperature-sensitive member temperature versus valve-opening-area characteristic obtained by the VTC apparatus of the embodiment, and (ii) a cooling-oil-temperature versus valve-opening-area characteristic obtained by a VTC system of a comparative example using a bimetal temperature-sensitive valve (simply, a bimetal valve) whose deflecting movement directly adjusts the valve opening area.

FIG. 11 is a longitudinal cross-sectional view illustrating the essential part of a first modified flow control valve system.

FIG. 12 is a longitudinal cross-sectional view illustrating the essential part of a second modified flow control valve system.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

Referring now to the drawings, particularly to FIGS. 1 to 3, the variable valve timing control (VTC) apparatus and its cooling device of the embodiment are exemplified in an intake-valve actuating mechanism of an internal combustion engine.

As clearly shown in FIGS. 1-3, the VTC apparatus of the embodiment is provided with a camshaft 1 rotatably supported on a cylinder head (not shown) of the internal combustion engine, a timing sprocket 2 (serving as a driving rotational member) installed on the front end of camshaft 1 such that relative rotation of timing sprocket 2 to camshaft 1 is permitted, and a phase change mechanism 3 arranged in the inner periphery of timing sprocket 2 and interposed between camshaft 1 and timing sprocket 2 for changing a relative angular phase between camshaft 1 and timing sprocket 2.

As best seen in FIGS. 2-3, camshaft 1 has two cams 1a, 1a per cylinder. Cams 1a, 1a are integrally formed on the outer periphery of camshaft 1, for operating the associated intake valves (not shown). As best seen in FIG. 1, a driven shaft member 4 (serving as a driven rotational member) is fixedly connected to one axial end of camshaft 1 by tightening a cam bolt 5 in the axial direction. A sleeve 6 is press-fitted onto one axial end (a cylindrical hollow portion 4a described later) of driven shaft member 4.

As clearly shown in FIGS. 1 and 3, driven shaft member 4 has an axially-extending cylindrical hollow portion 4a and a large-diameter flange portion 4b formed integral with each other. Cylindrical hollow portion 4a has a through hole, serving as a bolt insertion hole, through which cam bolt 5 is inserted into driven shaft member 4 and then screwed into the axial end of camshaft 1. Large-diameter flange portion 4b is formed integral with the opposite axial end of cylindrical hollow portion 4a, facing the one axial end of camshaft 1. Sleeve 6 is press-fitted onto the outer periphery of cylindrical hollow portion 4a of driven shaft member 4.

As can be seen from the cross-sectional view of FIG. 1, timing sprocket 2 has a ring-shaped external toothed portion 2a formed on its outer periphery and a substantially disk-shaped plate portion 2b extending radially inwards from the inner periphery of ring-shaped external toothed portion 2a. Ring-shaped external toothed portion 2a is mechanically linked to an engine crankshaft (not shown) through a timing chain (not shown) to transmit mechanical power from the crankshaft to the sprocket. Plate portion 2b is formed integral with a cylindrical central hub in which a central through hole 2c is formed. The inner peripheral wall surface of central through hole 2c of the hub of plate portion 2b is rotatably supported on the outer periphery of axially-extending cylindrical hollow portion 4a of driven shaft member 4.

As best seen in FIG. 2, plate portion 2b has a pair of circumferentially equidistant spaced, radially-extending elongated opposite slots (or elongated opposite windows) 7, 7 formed therein. Each of elongated slots (radially-extending elliptic through holes) 7, 7 has a pair of opposed parallel sidewalls which serve as a radial guide (described later). Also formed in plate portion 2b are a pair of circumferentially-extending circular-arc-shaped guide slots 2d, 2d. Circular-arc-shaped guide slots 2d, 2d are laid out between radially-extending elongated opposite slots 7, 7, in a manner so as to be circumferentially offset from the respective elongated slots 7, 7 and also circumferentially equidistant spaced from each other. As described later in reference to FIG. 3, a basal end 8a (a first end of both ends) of each of a pair of link members 8, 8 is engaged with and slidably held into the associated guide slot 2d.

Each of guide slots 2d, 2d is formed into a circular-arc shape and circumferentially elongated along the circumference of central through hole 2c of plate portion 2b. The circumferential length of guide slot 2d is dimensioned to permit a specified displacement, that is, a designed maximum displacement of the basal end 8a of link member 8 (in other words, a designed maximum phase angle of the camshaft relative to the timing sprocket).

Each of link members 8, 8 is slightly curved and formed as a substantially boomerang-shaped or circular-arc-shaped member. The basal end 8a (the first end) of link member 8 is formed integral with a cylindrical portion extending parallel to the axis of camshaft 1. In a similar manner, the top end 8b (the second end) of link member 8 is formed integral with a cylindrical portion extending parallel to the axis of camshaft 1. In more detail, the cylindrical portion of basal end 8a and the cylindrical portion of top end 8b of link member 8 are both protruded towards plate portion 2b of timing sprocket 2. As can be seen from the disassembled view of FIG. 2, large-diameter flange portion 4b of driven shaft member 4 is also formed at its inner peripheral portion with an axially backward protruded boss 4c extending towards the camshaft side and having a pair of radially-protruded pin-retention portions 4d, 4d, which will be hereinafter referred to as “lever protrusions”. Pin-retention through holes 4e, 4e are formed in the respective lever protrusions 4d, 4d of axially backward protruded boss 4c of large-diameter flange portion 4b, for retaining back ends of a pair of straight pins 9, 9 in place. In more detail, the back ends of pins 9, 9 are press-fitted into the respective pin-retention through holes 4e, 4e of lever protrusions 4d, 4d. The cylindrical portions of basal ends 8a, 8a of link members 8, 8 are rotatably linked onto the front ends of pins 9, 9.

On the other hand, the cylindrical portions of top ends 8b, 8b (the second ends) of link members 8, 8 are engaged with and slidably held into the respective radially-extending elongated slots 7, 7. As best seen in FIG. 1, each of the cylindrical portions of top ends 8b, 8b of link members 8, 8 has an axial hole 10 formed therein and axially forwardly opening. A spring-loaded, bullet-shaped engage pin 11 is operably held in axial hole 10, such that a substantially semi-spherical axial end of bullet-shaped engage pin 11 engages through the associated elongated slot 7 with a spiral groove 15, which is formed in a spiral guide-groove disk (simply, a spiral disk 13 described later) In FIG. 1, a return spring, denoted by reference sign 12, permanently forces bullet-shaped engage pin 11 axially forwards, so that the semi-spherical end of pin 11 is kept in engagement with the spiral groove 15 of spiral disk 13.

As seen in FIG. 1, when assembled, the top ends 8b, 8b of link members 8, 8 are engaged with the respective radially-extending elongated slots 7, 7 of plate portion 2b, and additionally basal ends 8a, 8a of link members 8, 8 are pivotably linked onto the front ends of pins 9, 9 fixedly connected to the respective lever protrusions 4d, 4d of axially backward protruded boss 4c of large-diameter flange portion 4b of driven shaft member 4. Under the assembled state, when the top ends 8b, 8b of link members 8, 8 radially displace along the respective elongated slots (radial guides) 7, 7 by way of external force application, timing sprocket 2 and driven shaft member 4 relatively rotates by an angular phase determined based on a radial displacement of each of top ends 8b, 8b, with displacements (circumferential sliding motions) of basal ends 8a, 8a along the respective guide slots 2d, 2d.

As clearly seen in FIGS. 1-2, spiral-guide groove disk (spiral disk) 13 is arranged to be opposed to the front face of plate portion 2b of timing sprocket 2. Spiral disk 13 serves as an intermediate rotational member, which is arranged between camshaft 1 (driven shaft member 4) and timing sprocket 2, and rotatably supported on the outer peripheral wall surface of cylindrical hollow portion 4a of driven shaft member 4. As best seen in FIG. 1, spiral disk 13 is comprised of an inner peripheral hub portion 13a and a disk portion 13b radially extending from inner peripheral hub portion 13a. The inner peripheral wall surface of hub portion 13a is kept in sliding-contact with the outer peripheral wall surface of cylindrical hollow portion 4a of driven shaft member 4 to permit relative rotation of spiral disk 13 to cylindrical hollow portion 4a of driven shaft member 4. Disk portion 13b of spiral disk 13 has double spiral grooves 15, 15, each having a substantially semi-circular shape in cross section and serving as a spiral guide and formed in the rear face of disk portion 13b facing the one axial end of camshaft 1. The substantially semi-spherical axial end of bullet-shaped engage pin 11 of link member 8 is slidably engaged with and guided by an associated one of spiral grooves 15, 15.

The previously-noted spiral grooves 15, 15 are formed in the rear face of disk portion 13b of spiral disk 13 separately from each other. Briefly speaking, each spiral groove 15 is formed or configured, so that the spiral radius of spiral groove 15, defined as a distance between a given point on the centerline of the spiral of spiral groove 15 and the axis of camshaft 1, reduces gradually in a direction of rotation of timing sprocket 2. The outermost groove section 15a of spiral groove 15 is bent or inflected radially inwards at a given angle at an inflection point between the outermost groove section 15a and a regular spiral groove section 15b (except outermost groove section 15a of spiral groove 15). Regular spiral groove section 15b extends spirally inwards from the inflection point. Additionally, one half of outermost groove section 15a, ranging from a substantially intermediate position of outermost groove section 15a in the longitudinal direction along the centerline of spiral groove 15 to the semi-circular closed end of outermost groove section 15a, is further inflected or curved radially inwards at a very small angle (a very small rate of change in the spiral or a very small curvature).

That is, the geometry of regular spiral groove section 15b of radial groove 15 except outermost groove section 15a is formed or configured, so that a rate of change in the spiral of regular spiral groove section 15b is constant. Note that the rate of change in the spiral of spiral groove 15 is substantially equivalent to a rate of change in relative angular phase between the camshaft (the driven side) and the timing sprocket (the driving side). On the other hand, a rate of change in the spiral of outermost groove section 15a, ranging from the inflection point to the semi-circular closed end of outermost groove section 15a, is dimensioned and configured to be less than that of regular spiral groove section 15b, such that the spiral of outermost groove section 15a is formed into a substantially straight line (in other words, a very moderately curved line) along a tangential line of spiral disk 13. The longitudinal length of the spiral of outermost groove section 15a is specified or dimensioned to realize a designed phase change control characteristic. Actually, the longitudinal length of the spiral of outermost groove section 15a is set to a comparatively long length substantially corresponding to a cam-angle of 45 degrees. Additionally, as discussed above, one half of outermost groove section 15a, ranging from the substantially intermediate position of outermost groove section 15a to the semi-circular closed end of outermost groove section 15a, is further inflected or curved radially inwards at a very small angle.

Suppose that spiral disk 13 rotates in a phase-retard direction relative to timing sprocket 2 in a state where bullet-shaped engage pins 11, 11 are kept in engagement with the respective spiral grooves 15, 15. At this time, the top ends 8b, 8b of link members 8, 8 move or displace radially inwards (towards the phase-advance side) along the respective spiral shapes of spiral grooves 15, 15, while being guided by radially-extending elongated slots 7, 7 of plate portion 2b. Conversely when spiral disk 13 relatively rotates in a phase-advance direction from this state, the top ends 8b, 8b of link members 8, 8 move radially outwards (towards the phase-advance side) along the respective spiral shapes of spiral grooves 15, 15, and then bullet-shaped engage pins 11, 11 reach the respective inflection points of spiral grooves 15, 15. Under such a state where engage pins 11, 11 become kept at the respective inflection points, the relative angular phase between camshaft 1 and timing sprocket 2 is controlled or adjusted to the maximum phase-retard position.

Under these conditions, when each of engage pins 11, 11 further displaces from the inflection point to the semi-circular closed end of outermost groove section 15a and then reaches a certain position within the outermost groove section 15a, the relative angular phase between camshaft 1 and timing sprocket 2 is controlled or adjusted to a phase of the intake valve, slightly phase-advanced from the maximum phase-retard position and suited to engine start-up operation.

In the presence of application of an operating force (or an actuating turning force) to spiral disk 13, which turning force produces relative rotation of spiral disk 13 to camshaft 1, the operating force (turning force) is transmitted through each of spiral grooves 15, 15 and each of the substantially semi-spherical axial ends of bullet-shaped engage pins 11, 11 to each of the top ends 8b, 8b of link members 8, 8. As a result of this, the top ends 8b, 8b displace radially along the respective elongated slots (radial guides) 7, 7. At this time, by virtue of motion-converting action of link members 8, 8, torque, which creates relative rotation between timing sprocket 2 and driven shaft member 4 (camshaft 1), is transmitted.

As shown in FIGS. 1-4, an operating force application mechanism, which is configured to apply the previously-noted operating force (actuating turning force) to spiral disk 13, is comprised of at least a return spring 16 and a hysteresis brake (an electromagnetic brake or an electromagnetic actuator) 17. In the shown embodiment, return spring 16 is constructed by a coiled torsion spring that forces spiral disk 13 in the rotation direction (i.e., towards the phase-advance side) of timing sprocket 2 via sleeve 6. Hysteresis brake 17 is provided to force spiral disk 13 in the direction opposite to the rotation direction of timing sprocket 2 by way of braking torque application (or braking force application). The magnitude of braking force produced by hysteresis brake 17 depending on an engine operating condition, such as engine load and engine speed, can be controlled by means of an electronic control unit (ECU), simply, a controller 50 (described later) incorporated in the VTC system. By appropriately controlling the braking force, produced by hysteresis brake 17, depending on an engine operating condition by means of controller 50, spiral disk 13 can be rotated relatively with respect to timing sprocket 2, or angular positions of spiral disk 13 and timing sprocket 2 can be retained unchanged.

Torsion spring 16 is installed on sleeve 6 in such a manner as to be wound on the outer periphery of sleeve 6. As shown in FIG. 1, one torsion-spring arm portion 16a of a comparatively short length is engaged with a radial spring-hanger hole formed in one axial end of sleeve 6, whereas the other torsion-spring arm portion 16b of a comparatively long length is engaged with an axial spring-hanger hole (an axial through hole) formed in inner peripheral hub portion 13a of spiral disk 13. Thus, torsion spring 16 serves to circumferentially force spiral disk 13 towards an angular phase position suited to engine start-up operation with hysteresis brake 17 deactivated after the engine has stopped.

As seen in FIGS. 1-3, hysteresis brake 17 includes an annular plate 14, a hysteresis ring 18, a substantially annular coil yoke 19, and an electromagnetic coil 20. Annular plate 14 is made of a non-magnetic material and fixedly connected to the perimeter of the front end face of disk portion 13b of spiral disk 13. Hysteresis ring 18 is fixedly connected to the perimeter of the front end face of annular plate 14. Coil yoke 19 is arranged in front of hysteresis ring 18 in such a manner as to circumferentially surround almost all of hysteresis ring 18. Electromagnetic coil 20 is circumferentially surrounded with coil yoke 19 to induce a magnetic flux in coil yoke 19.

Non-magnetic annular plate 14 is made of an austenitic stainless steel material and formed into an annular shape having a given width between inside and outside circles forming the annulus of annular plate 14. Annular plate 14 is fixedly connected to the perimeter of the front end face of disk portion 13b of spiral disk 13 by way of welding. The outside diameter of annular plate 14 is dimensioned to be greater than that of spiral disk 13.

As clearly shown in FIG. 1, hysteresis ring 18 is configured or formed into a thin-walled cylindrical-hollow shape, so that a radial width of hysteresis ring 18 (i.e., a radial dimension between inner and outer peripheral wall of hysteresis ring 18) is set to be considerably less than the given width of annular plate 14. Hysteresis ring 18 is fixedly connected to the perimeter of the front end face of annular plate 14 by way of welding. Hysteresis ring 18 is made of a hysteresis material (a semi-hard magnetic material) having a characteristic that a magnetic flux change occurs with a phase lag behind a change in external magnetic field.

Coil yoke 19 is comprised of an inner stator 22, an outer stator 23, and an annular yoke 24 bridging the front ends of inner and outer stators 22-23 in such a manner as to close one axial end of the annular space defined between inner and outer stators 22-23. Inner and outer stators 22-23 and annular yoke 24 are integrally formed with each other. As seen from the cross-sectional view in FIG. 1, coil yoke 19 is configured or formed into a substantially cylindrical shape that circumferentially surrounds the inner and outer peripheries and the front end face of electromagnetic coil 20 by the outer peripheral wall surface of inner stator 22, the inner peripheral wall surface of outer stator 23, and the right-hand sidewall of annular yoke 24.

An annular inner-stator component part 22a is integrally connected to the outer periphery of inner stator 22 by way of press fitting, in such a manner as to circumferentially surround the annular back face of electromagnetic coil 20. Inner stator 22 is integrally formed on its inner peripheral wall with a radially-inward protruded bearing-retention protrusion 22b. A radial ball bearing 25 is retained by bearing-retention protrusion 22b of inner stator 22. Thus, spiral disk 13 is rotatably supported on inner stator 22 via ball bearing 25.

Inner stator 22 (especially, press-fitted inner-stator component part 22a) has circumferentially equidistant spaced inner-stator polar teeth 26, 26, 26, 26, . . . , 26 (e.g., the number of teeth is “40”) formed on the outer periphery of press-fitted inner-stator component part 22a and serving as a south magnetic pole (a negative pole). In contrast, outer stator 23 has circumferentially equidistant spaced outer-stator polar teeth 27, 27, 27, 27, . . . , 27 (e.g., the number of teeth is “40”) formed on the inner periphery of outer stator 23 and serving as a north magnetic pole (a positive pole). As seen in FIGS. 1-3, inner-stator polar teeth 26 are formed as an external toothed portion. In contrast, as seen in FIG. 1, outer-stator polar teeth 27 are formed as an internal toothed portion. The external toothed portion (inner-stator polar teeth 26) and the internal toothed portion (outer-stator polar teeth 27) are concentrically arranged with respect to the axis of camshaft 1 and spaced apart from each other by a predetermined clearance space (related to a designed magnetic gap). In more detail, the number (e.g., forty teeth) of inner-stator polar teeth 26 are equal to the number (e.g., forty teeth) of outer-stator polar teeth 27, and additionally inner-stator polar teeth 26 are laid out to be circumferentially offset from the respective outer-stator polar teeth 27, such that one inner-stator tooth 26, located adjacent to two adjacent outer-stator teeth 27, 27, is opposed to the recess defined between the two adjacent outer-stator teeth 27, 27. That is to say, inner-stator and outer-stator polar teeth 26 and 27 are arranged to alternate with each other in their circumferential directions. In other words, inner-stator and outer-stator polar teeth 26 and 27 are arranged to be radially obliquely opposed to each other.

Owing to the previously-discussed radially obliquely opposed and circumferentially-alternated layout of inner-stator and outer-stator polar teeth 26 and 27, a magnetic field can be produced between the adjacent obliquely-opposed pair of inner-stator and outer-stator polar teeth 26 and 27, upon excitation of electromagnetic coil 20. Thus, a direction of the produced magnetic field, which passes through the inside of hysteresis ring 18, is inclined obliquely relative to the circumferential direction.

The top land of each of inner-stator polar teeth 26 and the inner peripheral wall surface of hysteresis ring 18 are kept out of contact with each other with a slight air gap, while being radially opposed to each other. Additionally, the top land of each of outer-stator polar teeth 27 and the outer peripheral wall surface of hysteresis ring 18 are kept out of contact with each other with a slight air gap, while being radially opposed to each other. In order to ensure a large magnetic force, these air gaps are set to infinitesimal clearance spaces.

Annular yoke 24 has a through hole 24a, which is formed in a predetermined angular position (in the circumferential direction of annular yoke 24), and through which a harness 20a of electromagnetic coil 20 is wired to controller 50 (see FIG. 1).

When an exciting current (a magnetizing current) is applied from the output interface of controller 50 through harness 20a to electromagnetic coil 20, a magnetic field (or a magnetic force) is produced via coil yoke 19. Thus, a braking torque is applied to hysteresis ring 18 by way of such an electromagnetic force. That is, when an exciting current is applied to electromagnetic coil 20 to induce a magnetic flux in coil yoke 19 and also hysteresis ring 18 displaces in a magnetic field produced between inner-stator and outer-stator polar teeth 26 and 27 radially obliquely opposed to each other, a braking force can be produced owing to a difference between a direction of the magnetic flux induced in hysteresis ring 18 and a direction of the produced magnetic field. At this time, the magnitude of braking force becomes substantially in proportion to a strength of the produced magnetic field (that is, a magnitude of the exciting current applied to electromagnetic coil 20), irrespective of a rotational speed of hysteresis ring 18, exactly, a relative velocity of hysteresis ring 18 to each of the external toothed portion (inner-stator polar teeth 26) and the internal toothed portion (outer-stator polar teeth 27). Assuming that the magnitude of the applied exciting current to exciting coil 20 is kept constant, the magnitude of braking force applied to hysteresis ring 18 also becomes constant.

Controller 50 shown in FIG. 1 generally comprises a microcomputer. Controller 50 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 50 receives input information from various engine/vehicle sensors, namely a crankangle sensor (a crankshaft position sensor), an airflow meter, a throttle position sensor, an engine temperature sensor (e.g., an engine coolant temperature sensor), and the like. The crankangle sensor is provided for detecting revolutions of an engine crankshaft, i.e., “engine speed”. The airflow meter is provided for detecting an intake-air flow rate, regarded as “engine load”. The throttle position sensor is provided for detecting a throttle opening. The engine temperature sensor is provided for detecting the actual operating temperature (e.g., coolant temperature) of the engine. Within controller 50, the central processing unit (CPU) allows the access by the I/O interface of input informational data signals from the previously-discussed engine/vehicle sensors. The CPU of controller 50 is responsible for carrying the hysteresis brake control program stored in memories and is capable of performing necessary arithmetic and logic operations containing the hysteresis-brake electromagnetic-coil control current control processing. A computational result (an arithmetic calculation result), that is, a calculated electromagnetic-coil control current output signal is relayed through the output interface circuitry of the controller to an output stage, namely electromagnetic coil 20 of hysteresis brake 17.

As appreciated from the above, phase change mechanism 3 is comprised of radially-extending elongated opposite slots 7, 7, formed in timing-sprocket plate portion 2b, link members 8, 8, engage pins 11, 11, lever protrusions 4d, 4d of boss 4c of large-diameter flange portion 4b of driven shaft member 4, spiral disk 13, spiral groove 15, and hysteresis brake 17.

A cooling device (serving as cooling/lubricating oil supply means) is also provided in the VTC apparatus of the shown embodiment, for supplying cooling oil to phase change mechanism 3.

As shown in FIG. 1, the cooling device is constructed by an annular passage 28, an oblique oil supply passage 29, and a flow-rate control valve (simply, a flow control valve) 30. Annular passage 28 is defined between the inner peripheral wall of camshaft 1, which is cylindrical hollow in shape, and the outer peripheral wall (containing the external screw-threaded portion) of cam bolt 5 screwed into the front axial end of camshaft 1. Oblique oil supply passage 29 is formed in driven shaft member 4 in such a manner as to obliquely extend from the inner periphery of cylindrical hollow portion 4a (nearby the inner periphery of axially backward protruded boss 4c) into large-diameter flange portion 4b at an inclination angle with respect to the radial direction. The upstream end of oblique oil supply passage 29 communicates annular passage 28. Flow control valve 30 communicates the downstream end 29a of oblique oil supply passage 29, for controlling or adjusting a flow rate of cooling oil flowing through oil supply passage 29, depending on an oil temperature.

Although it is not clearly shown in the drawings, annular passage 28 is communicated with a main oil gallery (not shown) via which working oil (cooling/lubricating oil) discharged from an oil pump (not shown) is supplied to moving engine parts. That is, part of working oil discharged from the oil pump is used as cooling oil to be delivered or supplied to phase change mechanism 3 of the VTC apparatus. Generally, a driving source of the oil pump is an internal combustion engine. In lieu thereof, an electric motor may be used as a driving source of the oil pump.

The previously-noted oblique oil supply passage 29 is communicated at its downstream end 29a with the inside of phase change mechanism 3 through a valve bore 31 of flow control valve 30 (hereunder described in detail).

As best seen in FIGS. 1, 4, and 5, flow control valve 30 is mainly constructed by the valve bore 31, which bore is communicated with the downstream end 29a of oblique oil supply passage 29 and formed as an axial through hole in the valve housing (i.e., large-diameter flange portion 4b of driven shaft member 4), a valve element 32 slidably provided in valve bore 31 to permit an axial sliding motion (axial advancing/retreating motion) of valve element 32, and a temperature-sensitive member 33 whose deflection occurs owing to a change in atmospheric temperature including a temperature change in the supplied cooling oil. In the cooling device of the shown embodiment, temperature-sensitive member 33 is mechanically linked to one axial end of valve element 32, so as to create the axial sliding motion of valve element 32 in valve bore 31 by way of such a deflecting movement of temperature-sensitive member 33.

Valve bore 31 is formed in the inner peripheral portion of large-diameter flange portion 4b as a right-circular cylinder whose inside diameter is approximately uniform. One axial opening end (i.e., the inside opening end) 31a of valve bore 31 faces towards a clearance space C defined between the front face of large-diameter flange portion 4b of driven shaft member 4 and the back face of plate portion 2b of timing sprocket 2. The downstream end 29a of oblique oil supply passage 29 is formed to open into the substantially intermediate position of valve bore 31 in the axial direction of valve bore 31.

As shown in FIGS. 6-7, valve element 32 is a stepped cylindrical member having a substantially center small-diameter shaft section 34, a cylindrical large-diameter land section 35, a substantially cylindrical large-diameter valve section 36, and an annularly-grooved engaged section 37. Land section 35 is machined to axially slide in a very close-fitting bore (that is, axial valve bore 31) of the valve housing (large-diameter flange portion 4b). Land section 35 is formed integral with the rear end of small-diameter shaft section 34. Valve section 36 is also machined to axially slide in a very close-fitting bore (that is, axial valve bore 31) of the valve housing. Valve section 36 is formed integral with the front end of small-diameter shaft section 34. On the other hand, annularly-grooved engaged section 37 is formed integral with the front end of valve section 36.

As shown in FIGS. 9A-9C, the annular groove, defined on the outer periphery of small-diameter shaft portion 34 between large-diameter land section 35 and large-diameter valve section 36, provides an annular oil introduction chamber 38. The downstream end 29a of oblique oil supply passage 29 is permanently communicated with oil introduction chamber 38, regardless of the axial position of valve element 32. Additionally, an oblique oil exhaust passage 39 is formed in large-diameter flange portion 4b of driven shaft member 4 such that the downstream end 29a of oblique oil supply passage 29 and the upstream end 39a of oblique oil exhaust passage 39 are radially opposed to each other, sandwiching axial valve bore 31 between them. Oil exhaust passage 39 is provided for exhausting a surplus of oil introduced into oil introduction chamber 38 into the exterior space. The lateral cross-sectional area (fluid-flow passage area) of oil exhaust passage 39 is dimensioned to be adequately less than that of oil supply passage 29. Additionally, the size and dimensions of oil exhaust passage 39 are configured or designed to allow oil-temperature transmission (thermal conduction) through at least large-diameter flanged portion 4b to temperature-sensitive member 33 even during low-oil-temperature operating conditions (e.g., when starting a cold engine).

Two opposing inside ends 35a and 36a of land section 35 and valve section 36 are configured to serve as pressure-receiving surfaces for oil introduced into oil introduction chamber 38. The areas of the above-mentioned two opposing pressure-receiving surfaces are the same in projected area. The outside diameter of each of land section 35 and valve section 36 is dimensioned to be slightly less than the inside diameter of valve bore 31. Actually, in order to ensure properly low-friction, smooth axial sliding motion of land section 35 and valve section 36 in valve bore 31, a very small radial clearance is defined between the inner peripheral wall surface of valve bore 31 and the outer peripheral wall surface of each of land section 35 and valve section 36. The very small radial clearance is dimensioned to such size as to permit a film of oil to be interposed between the inner peripheral wall surface of valve bore 31 and the outer peripheral wall surface of each of land section 35 and valve section 36.

The lateral cross-sectional area of oil introduction chamber 38 is dimensioned to be greater than a summed value of the lateral cross-sectional area of oil supply passage 29, the lateral cross-sectional area of oil exhaust passage 39, and the summed cross-sectional area of lateral cross-sectional areas of a pair of control-flow-passage grooves 40, 40 (described later).

The outer periphery of valve section 36 is partly recessed to form a control-flow-passage recess. The control-flow-passage recess is comprised of two control-flow-passage grooves 40, 40 circumferentially spaced apart from each other. In the shown embodiment, the control-flow-passage grooves 40, 40 are circumferentially equidistant-spaced apart from each other approximately 180 degrees. Each of control-flow-passage grooves 40, 40 is formed as a stepped groove upwardly sloped stepwise from the pressure-receiving surface 36a. Each of control-flow-passage grooves 40, 40 (especially, in longitudinal cross section of the bottom face of each control-flow-passage groove 40) is comprised of a flat surface section 40a deeply recessed in close proximity to the pressure-receiving surface 36a (or the shaft section 34), a moderately-sloped surface section (or an intermediate sloped surface section) 40b up-sloped forwards from the deeply-recessed flat surface section 40a in such a manner as to gradually shallow forwards from the deeply-recessed flat surface section 40a, and a slightly-sloped end section 40c formed continuously with the upper end of intermediate sloped surface section 40b and slightly up-sloped from intermediate sloped surface section 40b in such a manner as to further shallow as compared to intermediate sloped surface section 40b.

The previously-noted annularly-grooved engaged section 37 is axially protruded from the center of the front end of valve section 36. Annularly-grooved engaged section 37 has an annular engaged groove 37a formed near the front end of valve section 36, and thus the top end of engaged section 37 is formed as an annular protrusion 37b. When assembling, annular engaged groove 37a is kept in engagement with the top end of temperature-sensitive member 33.

As shown in FIGS. 5, 8, and 9A-9C, temperature-sensitive member 33 is formed of or constructed by a substantially rectangular, four-ply temperature-sensitive plate, such as four bimetallic strips, each composed of two thin dissimilar metals with different coefficients of thermal expansion, bonded together or welded together. In the shown embodiment, four bonded bimetallic strips, each composed of two thin dissimilar metals with different temperature coefficients of expansion, bonded together, are used as temperature-sensitive member 33. A bolt insertion hole 33a is formed in the substantially circular-arc shaped basal end of temperature-sensitive member 33. Additionally, a U-shaped notch (or an engage notch) 33b is formed by cutting out a U shape from the substantially rectangular top end of temperature-sensitive member 33. As seen in FIGS. 5 and 8, a forked end (33c, 33c) of temperature-sensitive member 33, which forked end is constructed by a pair of parallel engage sections 33c, 33c and defines the aforesaid U-shaped notch 33b, is engaged with or fitted in annular engaged groove 37a, such that the annular engaged groove 37a is interposed between the two parallel engage sections 33c, 33c. The basal end of temperature-sensitive member 33 is fixedly connected to large-diameter flanged portion 4b of driven shaft member 4 by screwing a bolt 41, together with a circular washer 42, through bolt insertion hole 33a and the center hole of washer 42 into a female screw-threaded hole 4c formed in large-diameter flange portion 4b. As discussed above, the two engage sections 33c, 33c are engaged with or fitted in annular engaged groove 37a through U-shaped notch 33b, in a manner so as to enable a deflecting movement of temperature-sensitive member 33 to be motion-converted into an axial sliding motion of valve element 32 via the two opposing inside edges of U-shaped notch 33b (or the forked end 33c, 33c) of temperature-sensitive member 33.

Annular protrusion 37b of valve element 32 is configured or dimensioned to satisfactorily prevent the forked end 33c, 33c of temperature-sensitive member 33 from being undesirably disengaged from annular engaged groove 37a during operation of phase change mechanism 3 of the VTC apparatus.

As clearly shown in FIG. 1, oil supplied into clearance space C, which is defined between large-diameter flange portion 4b and timing-sprocket plate portion 2b, is fed via the fluid-flow passage around the basal end 8a of link member 8 into an aperture defined between timing-sprocket plate portion 2b and disk portion 13b of spiral disk 13, and then fed through an oil hole (a through hole) 43 formed or bored in disk portion 13b to ball bearing 25, and further delivered into a clearance space defined between the inner periphery of hysteresis ring 18 and the external toothed portion (inner-stator polar teeth 26) and a clearance space defined between the outer periphery of hysteresis ring 18 and the internal toothed portion (outer-stator polar teeth 27).

With the previously-discussed arrangement, the VTC apparatus of the embodiment operates as follows. Under the engine stopped state, there is no exciting current application from the output interface of controller 50 to electromagnetic coil 20. Thus, spiral disk 13 is fully rotated relative to timing sprocket 2 in the rotation direction of the engine by way of the spring bias of torsion spring 16. At this time, the semi-spherical axial end of bullet-shaped engage pin 11 is shifted towards and kept abutted-engagement with the semi-circular closed end of outermost groove section 15a of spiral groove 15. Therefore, the relative angular phase of camshaft 1 to the crankshaft (that is, engine valve timing, more concretely, intake-valve open- and closure-timing in the shown embodiment) is shifted to the engine start-up phase, which is slightly phase-advanced from the maximum phase-retard position, and then kept at this phase.

After having started up the engine, when the engine shifts to a low-speed operating range such as idling speed, electromagnetic coil 20 of the hysteresis brake is energized by an exciting current output from controller 50. With coil 20 energized, a braking torque is applied to hysteresis ring 18, and thus a braking force, resulting from the braking torque application, is also applied to spiral disk 13 against the spring force of torsion spring 16.

Therefore, the semi-spherical axial end of bullet-shaped engage pin 11 rapidly retreats from the semi-circular closed end of outermost groove section 15a of spiral groove 15, and then moves towards the inflection point between outermost groove section 15a and regular spiral groove section 15b. Hence, spiral disk 13 rotates slightly in the direction opposite to the direction of rotation of timing sprocket 2. Thus, top end 8b of link member 8 slightly displaces radially outwards along the associated elongated slot (radial guide) 7 with an oscillating motion of link member 8, while bullet-shaped engage pin 11, which is operably held in axial hole 10 of top end 8b, is guided by spiral groove 15. Owing to the oscillating motion of link member 8, the relative angular phase between camshaft 1 (driven shaft member 4) and timing sprocket 2 (the engine crankshaft) is changed towards the maximum phase-retard position. As a result of this, the relative angular phase of camshaft 1 to the crankshaft can be changed to an appropriate phase suited to an engine operating condition. For instance, it is a phase-retard position or the maximum phase-retard position suited to low-speed operation. This improves the fuel economy as well as the stability of rotation of the engine during idling.

Under these conditions, when the engine operating condition is changed to a normal operating condition such as high-temperature high-speed operation, controller 50 generates a control command signal (a further large exciting current) to electromagnetic coil 20 in order to change the relative angular phase to the maximum phase-advance position. A large braking force is applied to spiral disk 13 via hysteresis ring 18. Spiral disk 13 further rotates relative to timing sprocket 2 in the direction opposite to the direction of rotation of timing sprocket 2 against the spring force of torsion spring 16. Thus, top end 8b of link member 8 further displaces radially inwards along the associated elongated slot (radial guide) 7 with an oscillating motion of link member 8, while bullet-shaped engage pin 11 is guided by spiral groove 15. Owing to the oscillating motion of link member 8, the relative angular phase between camshaft 1 (driven shaft member 4) and timing sprocket 2 is changed towards the maximum phase-advance position. In other words, the relative angular phase of camshaft 1 to the crankshaft can be changed towards the maximum phase-advance position. This enables a high-powered engine (the largest effective output of engine power).

In addition to the above, in the case of the VTC apparatus of the shown embodiment employing the cooling device for cooling the inside of phase change mechanism 3, when a temperature of oil (lubricating oil used as cooling oil for phase change mechanism 3), which is introduced from oil supply passage 29 into oil introduction chamber 38, is a very low temperature (see a temperature value Ta in FIG. 10), for example temperatures below 10° C., there is no deflection of temperature-sensitive member 33 and thus the shape of temperature-sensitive member 33 is maintained substantially straight in longitudinal cross section (see FIG. 9A). Under these conditions, valve element 32 is kept at its valve closed state where valve section 36 fully closes oil introduction chamber 38. Thus, oil (cooling/lubricating fluid), introduced from annular passage 28 via oil supply passage 29 into oil introduction chamber 38, is prevented from further flowing into clearance space C. That is to say, fluid communication between oil introduction chamber 38 and clearance space C is blocked by means of valve element 32 fully closed. Oil in oil introduction chamber 38 is exhausted through oil exhaust passage 39 into the upper portion of the cylinder head. At this time, however, oil pressures tend to uniformly act on the opposing inside ends 35a and 36a (the pressure-receiving surfaces of land section 35 and valve section 36), since the lateral cross-sectional area (fluid-flow passage area) of oil exhaust passage 39 is dimensioned to be adequately less than that of oil supply passage 29. As a result, two forces (the above oil pressures acting on the opposing inside ends 35a and 36a) have the same magnitude and the same line of action but different sense, and thus any axial sliding motion (any advancing/retreating motion) of valve element 32 does not occur due to the two opposite axial forces balanced to each other, and therefore valve element 32 is kept at its valve closed position.

Under these conditions, when the temperature rise of oil, introduced into oil introduction chamber 38, further develops and the oil temperature becomes greater than a certain temperature value, heat is transferred from large-diameter flange portion 4b to temperature-sensitive member 33 effectively through bolt 41 having a good thermal conductivity. Thus, under such a somewhat warm oil-temperature state, the temperature of temperature-sensitive member 33 itself becomes held at almost the same temperature value as the oil temperature. By virtue of the heat-transfer path (from flange portion 4b through bolt 41 to temperature-sensitive member 33) as discussed previously, it is possible to effectively suppress undesired fluctuations in the valve actuating temperature (the valve-actuation starting point) of flow control valve 30, at which flow control valve 31 starts to open.

Thereafter, when the oil temperature rises up to a temperature (see a temperature value Tb in FIG. 10) greater than or equal to a predetermined temperature value (see the temperature value T₁in FIG. 10), as shown in FIG. 9B, the top end of temperature-sensitive member 33 displaces away from the front face of large-diameter flange portion 4b and slightly displaces towards the back face of timing-sprocket plate portion 2b owing to a positive deflecting movement of temperature-sensitive member 33. At the same time, valve element 32 axially advances towards the back face of timing-sprocket plate portion 2b through annularly-grooved engaged section 37, which is engaged with the forked end (33c, 33c) of temperature-sensitive member 33. As a result, slightly-sloped end section 40c and the upper end of intermediate sloped surface section 40b of control-flow-passage groove 40 enter or advance into clearance space C, and thus a small valve opening area (a small fluid-flow passage area) becomes defined between the outer periphery of the large-diameter valve section 36 and the inner periphery of valve bore 31. The valve opening area is hereinafter referred to as “an opening area of valve bore 31”. Accordingly, part of the oil in oil introduction chamber 38 is exhausted from oil exhaust passage 39, whereas the remainder of oil flows via control-flow-passage groove 40 into clearance space C.

Thereafter, owing to a further oil temperature rise, the positive deflection of temperature-sensitive member 33 further develops and thus the advancing motion of valve element 32 towards the back face of timing-sprocket plate portion 2b also develops. At this time, the opening area of valve bore 31 gradually increases owing to a properly-tuned gradual increase in the fluid-flow passage area defined by control-flow-passage groove 40 opening into clearance space C, arising from a slight increase in the advancing motion of valve element 32. The specified relationship between (i) the axial advancing motion (axial displacement) of valve element 32 and (ii) the fluid-flow passage area defined by control-flow-passage groove 40 opening into clearance space C, enables a gradual increase in the amount of oil flowing through flow control valve 30 into clearance space C. That is, even from the initial stage of actuation of valve element 32 (i.e., even from the valve-actuation starting point of flow control valve 30), the flow control valve 30 included in the cooling device of the embodiment permits a gradual increase in the amount of oil flowing through flow control valve 30 into clearance space C. Hence, even from the initial stage of actuation of flow control valve 30, the cooling system of the embodiment can achieve a gradual, moderate increase in the amount of oil (cooling/lubricating fluid) into phase change mechanism 3 (especially, hysteresis brake 17) of the VTC apparatus, more concretely, ball bearing 25, and a clearance space defined between the inner periphery of hysteresis ring 18 and the external toothed portion (inner-stator polar teeth 26) and a clearance space defined between the outer periphery of hysteresis ring 18 and the internal toothed portion (outer-stator polar teeth 27). Therefore, by virtue of the previously-discussed properly-controlled cooling oil supply to the hysteresis brake, in other words, the fairly-tuned cooling-oil-temperature versus valve-opening-area characteristic, even when the engine temperature is still low or even when the engine is warming up or still cold, it is possible to satisfactorily prevent or avoid an undesirably wasteful braking force from being produced by a drag torque, occurring owing to the viscosity of oil (cooling/lubricating fluid) delivered to hysteresis ring 18.

Subsequently to the above, when the oil temperature further rises and exceeds a predetermined high temperature value (see the temperature value T₃in FIG. 10), due to a great positive deflection of temperature-sensitive member 33, the degree of advancing motion of valve element 32 reaches its maximum positive axial displacement. Under these conditions, as shown in FIG. 9C, annular protrusion 37b of valve element 32 is brought into abutted-engagement with the back face of timing-sprocket plate portion 2b, and thus a further displacement of valve element 32 is limited or restricted. At this time, that is, when the oil temperature (or the temperature of temperature-sensitive member 33 itself) becomes greater than or equal to the predetermined high temperature value (see the temperature value T₃in FIG. 10), the fluid-flow passage area, which is defined by control-flow-passage groove 40 opening into clearance space C, in other words, the opening area of valve bore 31 becomes maximum. Thus, a fairly large amount of oil can be supplied via clearance space C into the inside of phase change mechanism 3. Therefore, it is possible to effectively adequately cool and lubricate main component parts, constructing the electromagnetic-brake type VTC apparatus, such as hysteresis ring 18, electromagnetic coil 20, and the motion-converter comprised of a link mechanism including at least link members 8, 8.

Conversely when the oil temperature drops from the predetermined high temperature value (see the temperature value T₃in FIG. 10) , due to a negative deflection of temperature-sensitive member 33, the retreating motion (the negative axial displacement) of valve element 32 into valve bore 31 occurs. In this manner, the valve opening area can be fairly controlled responsively to an oil temperature change (or a temperature-sensitive member temperature change).

Referring now to FIG. 10, there is shown the comparative characteristic diagram showing both (i) the relationship (indicated by the broken line) between a temperature of a temperature-sensitive member (that is, a deflection in temperature-sensitive member 33, more concretely, a deflection in the bimetallic plate of the temperature-sensitive bimetallic-plate actuated flow control valve 30) and a valve opening area (that is, the previously-discussed opening area of valve bore 31 defined by control-flow-passage groove 40 opening into clearance space C), obtained by the VTC apparatus of the embodiment, and (ii) the relationship (indicated by the thick solid line) between a cooling oil temperature (in other words, a bimetal-valve temperature) and a bimetal-valve opening area, obtained by a VTC system of a comparative example using a bimetal temperature-sensitive valve (simply, a bimetal valve) whose deflecting movement directly adjusts the valve opening area. In FIG. 10, the one-dotted line “a” indicates an operating state at a very low temperature “Ta”, which is less than or equal to the predetermined temperature value T₁. The one-dotted line “b” indicates an operating state at an intermediate temperature “Tb”, where the temperature-sensitive member (or the cooling/lubricating fluid) rises up to a temperature value greater than the predetermined temperature value T₁. On the other hand, the one-dotted line “c” indicates an operating state at a high temperature value “Tc” exceeding the predetermined high temperature value T₃.

As can be seen from the cooling-oil-temperature versus valve-opening-area characteristic obtained by the VTC system of the comparative example using the bimetal valve, indicated by the broken line in FIG. 10, as soon as the oil temperature (the temperature of the bimetal valve) exceeds the predetermined temperature value T₁, the opening part of the bimetal valve is rapidly opened by way of a deflection of the bimetal valve and thus due to such an instantaneous increase in the bimetal valve opening, an undesirably large amount of oil would be supplied into the phase change mechanism.

In contrast to the above, according to the cooling device employed in the VTC apparatus of the embodiment, as can be seen from the temperature-sensitive member temperature versus valve-opening-area characteristic, indicated by the thick solid line in FIG. 10, as the oil temperature (the temperature of temperature-sensitive member 33) further rises from the predetermined temperature value T₁, a gradual advancing motion of valve element 32 occurs owing to a deflecting movement of temperature-sensitive member 33. Thus, it is possible to gradually increase the opening area of valve bore 31 along the moderate characteristic curve (as indicated between the predetermined temperature values T₁and T₃in FIG. 10) via control-flow-passage grooves 40, 40. Therefore, the flow control valve system incorporated in the VTC apparatus of the embodiment, enables high-precision oil-flow-rate control for oil (cooling/lubricating fluid) supplied inside of phase change mechanism 3, responsively to an oil temperature change (or a temperature change in temperature-sensitive member 33). As a result of this, even during low engine temperature operation, it is possible to satisfactorily prevent an undesirably wasteful braking force from being produced by a drag torque, occurring owing to the viscosity of oil (cooling/lubricating fluid) delivered to hysteresis ring 18 of hysteresis brake 17. Thus, even when the engine temperature is low during engine warm-up, valve timing compensation, executed by controller 50 incorporated in the VTC apparatus of the embodiment, can be fairly appropriately achieved without any wasteful braking torque.

Additionally, according to the flow control valve system incorporated in the VTC apparatus of the embodiment, when starting a cold engine (at a very low temperature below 10° C.), two forces (oil pressures acting on the opposing inside ends 35a and 36a) have the same magnitude and the same line of action but different sense, and thus any axial sliding motion (any advancing/retreating motion) of valve element 32 does not occur due to the two opposite axial forces balanced to each other, and therefore valve element 32 is kept at its valve closed position. In other words, at the very low temperatures, the axial position (the valve closed position) of valve element 32 can be held by way of the two opposite axial forces balanced to each other (the two opposite pressures balanced to each other). It is possible to efficiently and adequately reduce the magnitude of force needed to actuate (advance/retreat) valve element 32 of flow control valve 30 by means of temperature-sensitive member 33, which is deflected depending on a temperature change. Therefore, it is possible to downsize the flow-control-valve actuating mechanism (i.e., temperature-sensitive member 33).

Furthermore, in the flow control valve system incorporated in the VTC apparatus of the embodiment, a proper, very small clearance is defined between annular engaged groove 37a of valve element 32 and U-shaped notch (engage notch) 33b of temperature-sensitive member 33. This clearance enables or permits valve element 32 to constantly smoothly slide axially in valve bore 31, depending on a deflection (a deflecting movement) of temperature-sensitive member 33. Thus, it is possible to more accurately control the flow rate of oil (cooling/lubricating fluid) flown into the phase change mechanism responsively to an oil temperature change (or a temperature change in temperature-sensitive member 33).

Referring now to FIG. 11, there is shown the longitudinal cross section of the essential part of the first modified flow control valve system. The first modified flow control valve system of FIG. 11, eliminates an oblique oil exhaust passage (denoted by reference sign 39), as formed in large-diameter flange portion 4b in the flow control valve system of the VTC apparatus of the embodiment as clearly shown in FIGS. 1, and 9A-9C. In lieu thereof, an axial notch (an axial groove) is formed by axially cutting out a V shape or a U shape from the outer periphery of land section 35 of valve element 32. Thus, an oil exhaust passage 44 is defined between the recessed wall surface of the axial notch (the axial groove) formed in valve land 35 and the inner peripheral wall surface of valve bore 31. According to the first modified flow control valve system of FIG. 11, oil exhaust passage 44 can be easily formed by cutting work of the outer peripheral wall surface of valve land 35. This ensures lower machining time and costs.

Referring now to FIG. 12, there is shown the longitudinal cross section of the essential part of the second modified flow control valve system. The second modified flow control valve system of FIG. 12, also eliminates an oblique oil exhaust passage (denoted by reference sign 39), as formed in large-diameter flange portion 4b in the flow control valve system of the VTC apparatus of the embodiment as clearly shown in FIGS. 1, and 9A-9C. In lieu thereof, an oil exhaust passage 45 is constructed by (i) a through hole (exactly, a radial through hole), which is formed in small-diameter shaft section 34 of valve element 32 and whose both ends open into the oil introduction chamber 38, and (ii) an axial bore 45b formed in valve element 32 in such a manner as to range from the rear end of small-diameter shaft section 34 to the rearmost end of large-diameter land section 35. One axial end of axial bore 45b communicates with the intermediate portion of radial through hole 45a, whereas the other axial end of axial bore 45b communicates the exterior space. According to the second modified flow control valve system of FIG. 12, oil exhaust passage 45 can be easily formed by drilling work of the inside of valve element 32. This ensures lower machining time and costs.

In the shown embodiment, the variable valve timing control apparatus and its cooling device is exemplified in an intake-valve actuating mechanism of an internal combustion engine. It will be appreciated that the VTC apparatus of the embodiment can be applied to an exhaust-valve actuating mechanism.

Furthermore, in the shown embodiment, a bimetallic member, consisting of a plurality of (two or more) bimetallic strips bonded together, is used as temperature-sensitive member 33. Instead of using such a bimetallic member, a shape memory alloy or a wax pellet, which expands with increasing temperature and opens a flow control valve, may be used.

Moreover, in the shown embodiment, the control-flow-passage recess of flow control valve 30 is exemplified in a pair of control-flow-passage grooves (40, 40) circumferentially spaced apart from each other. In lieu thereof, only one control-flow-passage groove, or three or more control-flow-passage grooves may be formed as the control-flow-passage recess.

The entire contents of Japanese Patent Application No. 2007-228428 (filed Sep. 4, 2007) 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.

VARIABLE VALVE TIMING CONTROL APPARATUS OF INTERNAL COMBUSTION ENGINE AND COOLING DEVICE FOR THE SAME

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