Internal Combustion Engine Control Device and Internal Combustion Engine Control System

Abstract

In an internal combustion engine control device, when the engine stops, a variable valve actuation device is controlled to change an operating mode of each of intake valves to a specific state where an all-cylinder valve closed period, during which the intake valves of all engine cylinders are all kept in their non-lifted states, occurs, and a crank position change mechanism is controlled to change a crankshaft-rotation stopped position to a crankangle included within the all-cylinder valve closed period. When restarting the engine, the variable valve actuation device is controlled to bring the operating mode of each of the intake valves closer to an engine-start desired lift characteristic suited to an engine condition, prior to cranking action.

Description

TECHNICAL FIELD

The present invention relates to a control device and a control system, capable of enhancing a restartability of a multi-cylinder internal combustion engine.

BACKGROUND ART

In recent years, there have been proposed and developed various multi-cylinder engine control devices configured to enhance a restartability. One such multi-cylinder internal combustion engine control device has been disclosed in Japanese Patent Provisional Publication No. 2003-56316 (hereinafter is referred to as “JP2003-056316”). Briefly speaking, as a variable valve actuation device, the engine control device disclosed in JP2003-056316 has a continuous variable valve event and lift control (VEL) mechanism capable of continuously varying a valve lift amount and a working angle of an intake valve and a variable phase control mechanism (a variable valve timing control (VTC) mechanism) capable of phase-shifting a central phase angle (e.g., a phase at a peak valve lift) of a valve lift characteristic curve of the intake valve. When an intake-valve drive shaft is rotated in synchronism with rotation of an engine crankshaft, rotary motion of the intake-valve drive shaft is converted into oscillating motion of a rockable cam via the VEL mechanism. Also provided is a hydraulically-operated valve-lash adjuster interleaved between the rockable cam and the intake-valve stem, so as to provide zero valve lash, utilizing hydraulic pressure. In an engine stopped state, valve lift characteristics of intake valves of all engine cylinders are set to realize a so-called all-cylinder zero-lift state where an intake-valve lift amount of each individual engine cylinder, to which the hydraulically-operated lash adjuster is applied, is almost zero.

That is to say, on the assumption that a stopped position of rotary motion of a crankshaft of a four-cylinder internal combustion engine is a substantially midpoint between a piston top dead center (TDC) position on compression stroke and a piston bottom dead center (BDC) position on compression stroke, this crankshaft-rotation stopped position is set, so that intake valves of all engine cylinders become kept at their zero-lift states. Hence, in the engine stopped state, it is possible to prevent the hydraulically-operated lash adjuster of each individual cylinder from undesirably contracting owing to working oil leakage, and whereby fluctuations in valve lift amounts between the engine cylinders can be minimized, thus improving an engine restartability.

SUMMARY OF THE INVENTION

However, in the case of the engine control device as disclosed in JP2003-056316, as previously discussed, the intake-valve lift characteristic (in particular, intake-valve open timing and intake-valve closure timing) of each individual engine cylinder is set to realize the previously-noted all-cylinder zero-lift state, on the assumption that a crankshaft-rotation stopped position is a substantially midpoint between a TDC position on compression stroke and a BDC position on compression stroke. Hence, in an engine stopped state, the peak-lift phase of the intake valve tends to become a phase near the TDC position (see FIG. 8 of JP2003-056316) or a phase near the BDC position (see FIG. 9 of JP2003-056316). Thus, when cranking the engine, the VTC mechanism must be operated greatly, but the VTC mechanism exhibits a bad operational responsiveness at very low engine speeds. Such a control device requires a comparatively long time duration until an initial explosion of the engine takes place. As a result of this, it is difficult to ensure a good restartability.

It is, therefore, in view of the previously-described disadvantages of the prior art, an object of the invention to provide a control device or a control system configured to control a variable valve actuation device in such a manner as to bring an actual lift characteristic of an intake valve of each individual cylinder of an internal combustion engine closer to a desired lift characteristic suited to a monitored engine condition, prior to cranking action, thereby ensuring a good engine restartability.

In order to accomplish the aforementioned and other objects of the present invention, an internal combustion engine control device comprises a crank position change mechanism configured to change a crankshaft-rotation stopped position of a crankshaft of an internal combustion engine, and a variable valve actuation device configured to change at least a working angle of each of intake valves of a plurality of engine cylinders by changing a position of a control shaft, wherein, when the engine stops, the variable valve actuation device is controlled to change an operating mode of each of the intake valves to a specific state where an all-cylinder valve closed period, during which the intake valves of the cylinders are all kept in their non-lifted states, occurs, and the crank position change mechanism is controlled to change the crankshaft-rotation stopped position to a crankangle included within the all-cylinder valve closed period, and wherein, when restarting the engine, the variable valve actuation device is controlled to bring the operating mode of each of the intake valves closer to an engine-start desired lift characteristic suited to an engine condition, prior to cranking action.

According to another aspect of the invention, an internal combustion engine control device comprises a crank position change mechanism configured to change a crankshaft-rotation stopped position of a crankshaft of an internal combustion engine, and a variable valve actuation device configured to change at least a working angle of each of intake valves of a plurality of engine cylinders by changing a position of a control shaft, wherein, when the engine stops, the crank position change mechanism, together with the variable valve actuation device, is controlled to execute crank position control as well as intake-valve operating characteristic control in such a manner as to realize a specific state where there is a less valve-spring reaction force acting on the control shaft, and wherein, when restarting the engine, the variable valve actuation device is controlled to bring the position of the control shaft closer to a desired position suited to start-up of the engine, prior to cranking action.

According to a further aspect of the invention, an internal combustion engine control system comprises a crank position change mechanism configured to change a crankshaft-rotation stopped position of a crankshaft of an internal combustion engine, a variable valve actuation device configured to change a valve lift as well as a working angle of each of intake valves of a plurality of engine cylinders, and a controller configured to control operations of the crank position change mechanism and the variable valve actuation device, wherein, when the engine stops, the controller controls the variable valve actuation device to change an operating characteristic of each of the intake valves to a specific state where an all-cylinder valve closed period, during which the intake valves of the cylinders are all kept in their non-lifted states, occurs, and the controller controls the crank position change mechanism to change the crankshaft-rotation stopped position to a crankangle included within the all-cylinder valve closed period, and wherein, when restarting the engine, the controller controls the variable valve actuation device to bring an engine-start desired lift characteristic of each of the intake valves, suited to an engine condition, prior to cranking action.

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 schematic system diagram illustrating an internal combustion engine to which a variable valve actuation device of an embodiment can be applied.

FIG. 2 is a perspective view illustrating the variable intake-valve actuation device of the embodiment, which includes a continuously variable valve event and lift control (VEL) mechanism and a variable valve timing control (VTC) mechanism.

FIGS. 3A-3B are axial views showing the operation of the intake-valve VEL mechanism during a small-lift control mode.

FIGS. 4A-4B are axial views showing the operation of the intake-valve VEL mechanism during a large-lift control mode.

FIG. 5 is a variable intake-valve lift and event (working angle) and phase characteristic diagram, obtained by both of the intake-valve VEL and VTC mechanisms of the variable valve actuation device of the embodiment.

FIG. 6 is a cross-sectional view showing the intake-valve VTC mechanism included in the variable valve actuation device of the embodiment.

FIG. 7 is a lateral cross-section taken along the line A-A of FIG. 6, and showing the maximum phase-retard state of the intake-valve VTC mechanism.

FIG. 8 is a lateral cross-section taken along the line A-A of FIG. 6, and showing the maximum phase-advance state of the intake-valve VTC mechanism.

FIGS. 9A-9D are characteristic diagrams showing the relationship among crankangle, intake-valve open timing, and intake-valve closure timing, obtained by the variable valve actuation device of the first embodiment, at each of #1, #3, #4, and #2 engine cylinders.

FIG. 10 is a flowchart showing a control routine executed within a controller incorporated in the engine control system of the present embodiment.

FIGS. 11A-11B are characteristic diagrams showing the relationship among crankangle, intake-valve open timing, and intake-valve closure timing, obtained by the variable valve actuation device of the second embodiment, at each of #1, and #2 engine cylinders.

FIG. 12 is an axial view showing the operation of the intake-valve VEL mechanism included in the variable valve actuation device of the second embodiment.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

First Embodiment

Referring now to the drawings, particularly to FIGS. 1-2, the engine control device of the first embodiment is exemplified in a four-cycle four-cylinder internal combustion gasoline engine having four valves per cylinder, namely two intake valves 4, 4 (see FIGS. 1-2) and two exhaust valves 08, 08 (see FIG. 1).

The construction of the multiple-cylinder internal combustion engine, to which the engine control device of the embodiment can be applied, is hereunder described in detail in reference to the system diagram of FIG. 1. Also, the engine control device of the embodiment can be applied to a hybrid vehicle (HV) employing an automatic engine stop-restart system capable of temporarily automatically stopping an internal combustion engine during idling without depending on a driver's will, in addition to an internal-combustion-engine equipped vehicle with a so-called idling-stop system that enables an idling-stop action. The engine of FIG. 1 is constructed by a cylinder block SB having a cylinder bore, a reciprocating piston 01 movable or slidable through a stroke in the cylinder bore, a cylinder head SH on the cylinder block SB, an intake port IP and an exhaust port EP formed in cylinder head SH, two intake valves 4, 4 each slidably installed on cylinder head SH for opening and closing the opening end of intake port IP, and two exhaust valves 08, 08 each slidably installed on cylinder head SH for opening and closing the opening end of exhaust port EP.

Piston 01 is connected to an engine crankshaft 02 via a connecting rod 03. A combustion chamber 04 is defined between the piston crown of piston 01 and the underside of cylinder head SH.

An electronically-controlled throttle valve unit SV is provided upstream of intake port IP and located in an interior space of an intake manifold Ia of an intake pipe I connected to intake port IP, for controlling a quantity of intake air. A fuel injector or a fuel injecting valve (not shown) is provided downstream of throttle valve unit SV. A spark plug 05 is located substantially in a middle of cylinder head SH.

A flywheel ring gear 09 of a flywheel attached to engine crankshaft 02 is in meshed-engagement with a pinion gear mechanism 06. Cranking action of crankshaft 02 can be initiated by driving the pinion gear mechanism 06 by means of an electric motor (or a cranking motor) 07, and simultaneously a rotational position of crankshaft 02 can be controlled. That is, electric motor 07 and pinion gear mechanism 06 construct a part of a crank position change mechanism.

Each of intake valves 4, 4 is forced by the associated valve spring 5 in a direction that the opening end of intake port IP is closed.

As clearly shown in FIGS. 1-2, particularly, in FIG. 2, the variable valve actuation device incorporated in the engine control system of the embodiment is comprised of an intake-valve variable valve event and lift control (VEL) mechanism 1 and an intake-valve variable valve timing control (VTC) mechanism (or a variable phase control mechanism) 2. Intake-valve VEL mechanism 1 is able to simultaneously control or adjust or change both of a valve lift and a lifted-period (a working angle, in other words, a valve open period) of each of intake valves 4, 4. Intake-valve VTC mechanism 2 is able to advance or retard only a phase of each of intake valves 4, 4, while keeping a valve lift and working angle characteristic of each intake valve 4 constant. In the shown embodiment, there is no exhaust-valve VTC mechanism in the exhaust valve side, and thus exhaust-valve open timing (EVO) and exhaust valve closure timing (EVC) are both fixed.

As the VEL mechanism 1, the variable valve actuation system of the embodiment uses a continuously variable valve event and lift control mechanism as disclosed in Japanese Patent Provisional Publication No. 2003-172112. Briefly speaking, as shown in FIG. 2, VEL mechanism 1 is comprised of a cylindrical hollow drive shaft 6, a ring-shaped drive cam 7, two rockable cams 9, 9, and a multinodular-link motion transmitting mechanism (or a motion converter) mechanically linked between drive cam 7 and the rockable-cam pair (9, 9) for transmitting a torque created by drive cam (eccentric cam) 7 as an oscillating force of each of rockable cams 9, 9. Cylindrical hollow drive shaft 6 is rotatably supported by bearings in the upper part of cylinder head SH. Drive cam 7 is formed as an eccentric cam that is press-fitted or integrally connected onto the outer periphery of drive shaft 6. Rockable cams 9, 9 are oscillatingly or rockably supported on the outer periphery of drive shaft 6 and in sliding-contact with respective upper contact surfaces of two valve lifters 8, 8, which are located at the valve stem ends of intake valves 4, 4. In other words, the motion transmitting mechanism (or the motion converter) is provided to convert a rotary motion (input torque) of drive cam 7 into an up-and-down motion (a valve opening force) of each intake valve 4 (i.e., an oscillating force creating an oscillating motion of each rockable cam 9).

Torque is transmitted from engine crankshaft 02 through a timing sprocket 33 fixedly connected to one axial end of drive shaft 6 via a timing chain (not shown) to the drive shaft 6. As indicated by the arrow in FIG. 2, the direction of rotation of drive shaft 6 is set in a clockwise direction.

Drive cam 7 has an axial bore that is displaced from the geometric center of the cylindrical drive cam 7. Drive cam 7 is fixedly connected to the outer periphery of drive shaft 6, so that the inner peripheral surface of the axial bore of drive cam 7 is press-fitted onto the outer periphery of drive shaft 6. Thus, the center of drive cam 7 is offset from the shaft center of drive shaft 6 in the radial direction by a predetermined eccentricity (or a predetermined offset distance).

As best seen from the axial views shown in FIGS. 2, 3A-3B and 4A-4B, each of rockable cams 9, 9 is formed as a substantially raindrop-shaped cam. Rockable cams 9, 9 have the same cam profile. Rockable cams 9, 9 are formed integral with respective axial ends of a cylindrical-hollow camshaft 10. Cylindrical-hollow camshaft 10 is rotatably supported on drive shaft 6. The outer peripheral contacting surface of rockable cam 9, in sliding-contact with the upper contact surface of valve lifter 8, includes a cam surface 9a. The base-circle portion of rockable cam 9 is integrally formed with or integrally connected to camshaft 10, to permit oscillating motion of rockable cam 9 on the axis of drive shaft 6. The outer peripheral surface (cam surface 9a) of rockable cam 9 is constructed by a base-circle surface, a circular-arc shaped ramp surface extending from the base-circle surface to a cam-nose portion, a top-circle surface (simply, a top surface) that provides a maximum valve lift (or a maximum lift amount), and a lift surface by which the ramp surface and the top surface are joined. The base-circle surface, the ramp surface, the lift surface, and the top surface abut predetermined positions of the upper surface of valve lifter 8, depending on the oscillatory position of rockable cam 9.

The motion transmitting mechanism (the motion converter) is comprised of a rocker arm 11 laid out above drive shaft 6, a link arm 12 mechanically linking one end (or a first end portion 11a) of rocker arm 11 to the drive cam 7, and a link rod 13 mechanically linking the other end (a second end portion 11b) of rocker arm 11 to the cam-nose portion of rockable cam 9.

Rocker arm 11 is formed with an axially-extending center bore (a through opening). The rocker-arm center bore of rocker arm 11 is rotatably fitted onto the outer periphery of a control cam 18 (described later), to cause a pivotal motion (or an oscillating motion) of rocker arm 11 on the axis of control cam 18. The first arm portion 11a of rocker arm 11 extends from the axial center bore portion in a first radial direction, whereas the second arm portion lib of rocker arm 11 extends from the axial center bore portion in a second radial direction substantially opposite to the first radial direction. The first arm portion 11a of rocker arm 11 is rotatably pin-connected to link arm 12 by means of a connecting pin 14, while the second arm portion 11b of rocker arm 11 is rotatably pin-connected to one end (a first end 13a) of link rod 13 by means of a connecting pin 15.

Link arm 12 is comprised of a comparatively large-diameter annular base portion 12a and a comparatively small-diameter protruding end portion 12b radially outwardly extending from a predetermined portion of the outer periphery of large-diameter annular base portion 12a. Large-diameter annular base portion 12a is formed with a drive-cam retaining bore, which is rotatably fitted onto the outer periphery of drive cam 7. On the other hand, small-diameter protruding end portion 12b of link arm 12 is pin-connected to the first arm portion 11a of rocker arm 11 by means of connecting pin 14.

Link rod 13 is pin-connected at the other end (a second end 13b) to the cam-nose portion of rockable cam 9 by means of a connecting pin 16.

Also provided is a motion-converter attitude control mechanism that changes an initial actuated position (a fulcrum of oscillating motion of rocker arm 11) of the motion transmitting mechanism (or the motion converter). As clearly shown in FIGS. 3A-3B and 4A-4B, the attitude control mechanism includes a control shaft 17 and control cam 18. Control shaft 17 is located above and arranged in parallel with drive shaft 6 in such a manner as to extend in the longitudinal direction of the engine, and rotatably supported on cylinder head SH by means of the same bearing members as drive shaft 6. Control cam 18 is attached to the outer periphery of control shaft 17 and slidably fitted into and oscillatingly supported in a control-cam retaining bore formed in rocker arm 11. Control cam 18 serves as a fulcrum of oscillating motion of rocker arm 11. Control cam 18 is integrally formed with control shaft 17, so that control cam 18 is fixed onto the outer periphery of control shaft 17. Control cam 18 is formed as an eccentric cam having a cylindrical cam profile. The axis (the geometric center) of control cam 18 is displaced a predetermined distance from the axis of control shaft 17.

As shown in FIG. 2, the attitude control mechanism also includes a drive mechanism 19. Drive mechanism 19 is comprised of a geared motor or an electric control-shaft actuator 20 fixed to one end of a housing (not shown) and a ball-screw motion-transmitting mechanism (simply, a ball-screw mechanism) 21 that transmits a motor torque created by electric motor (simply, motor) 20 to control shaft 17. In more detail, motor 20 is constructed by a proportional control type direct-current (DC) motor. Motor 20 is controlled in response to a control signal, which is generated from the output interface circuitry of an electronic control unit (simply, a controller) 22 and whose signal value is determined based on engine/vehicle operating conditions.

Ball-screw mechanism 21 is comprised of a ball-screw shaft (or a worm shaft) 23 coaxially aligned with and connected to the motor output shaft of motor 20, a substantially cylindrical, movable ball nut 24 threadably engaged with the outer periphery of ball-screw shaft 23, a link arm 25 fixedly connected to the rear end 17a of control shaft 17, a link member 26 mechanically linking link arm 25 to ball nut 24, and recirculating balls interposed between the worm teeth of ball-screw shaft 23 and guide grooves cut in the inner peripheral wall surface of ball nut 24. In a conventional manner, a rotary motion (input torque) of ball-screw shaft 23 is converted into a rectilinear motion of ball nut 24 through the recirculating balls. Ball nut 24 is axially forced toward motor 20 by the spring force of a return spring (a coil spring) 30, serving as a biasing device (biasing means). The direction of the spring force (spring bias) of return spring 30 corresponds to a direction that the VEL mechanism is biased to a minimum valve lift and working angle characteristic. Hence, during an engine stopping period, ball nut 24 of ball-screw mechanism 21 can be forced and displaced by the spring bias of coil spring 30 in the axial direction of ball-screw shaft 23 corresponding to the minimum valve lift and working angle characteristic of the intake-valve VEL mechanism.

As shown in FIGS. 1-2, controller 22 generally comprises a microcomputer. Controller 22 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 22 receives input information from various engine/vehicle switches and sensors, namely a crankangle sensor (or a crankshaft position sensor) 27, an engine speed sensor, an accelerator opening sensor, a vehicle speed sensor, a range gear position switch, a drive-shaft angular position sensor 28, a control-shaft angular position sensor 29, an airflow meter AFM, an engine temperature sensor (i.e., an engine coolant temperature sensor, and the like. Within ECU 22, 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. The processor of ECU 22 determines the current engine/vehicle operating condition, based on input information from the engine/vehicle switches and sensors. Crankangle sensor 27 is provided to detect an angular position (crankangle) of crankshaft 02. Drive-shaft angular position sensor 28 is provided for detecting an angular position of drive shaft 6. Also, based on both of the sensor signals from crankangle sensor 27 and drive-shaft angular position sensor 28, an angular phase of drive shaft 6 relative to timing sprocket 33 is detected. Control-shaft angular position sensor 29 is provided to detect an angular position of control shaft 17. Airflow meter AFM is provided for measuring or detecting a quantity of air flowing through intake pipe I, and consequently for detecting or estimating the magnitude of engine load. The CPU of ECU 22 is responsible for carrying the control program stored in memories and is capable of performing necessary arithmetic and logic operations, for example, electronic throttle opening control achieved through the throttle actuator of electronically-controlled throttle valve unit SV, electronic fuel injection control achieved by the electronic fuel-injection system, electronic spark control achieved by the electronic ignition system, valve lift and working angle control executed by VEL mechanism 1, and phase control executed by VTC mechanism 2. Computational results (arithmetic calculation results), that is, calculated output signals are relayed through the output interface circuitry of ECU 22 to output stages, namely the throttle actuator of electronically-controlled throttle valve unit SV, electronically-controlled fuel injectors of the fuel-injection system, electronically-controlled spark plugs 05 of the electric ignition system, motor 20 of VEL mechanism 1, the solenoid of a directional control valve 47 (described later) for VTC mechanism 2, and electric motor 07 used for cranking motor control.

Hereunder described briefly in reference to FIGS. 2, 3A-3B, 4A-4B, and 5 is the basic operation of intake-valve VEL mechanism 1. In an engine operating range such as in a low-speed low-load range, the control-shaft actuator (motor 20) of VEL mechanism 1 is driven in one rotational direction in response to a control signal generated from the output interface circuitry of ECU 22. Thus, ball-screw shaft 23 is rotated in a direction corresponding to the one rotational direction of motor 20 by input torque created by motor 20, thereby producing a maximum rectilinear motion of ball nut 24 in one ball-nut axial direction that ball nut 24 approaches close to motor 20. As a result, control shaft 17 rotates in one rotational direction via a linkage comprised of link member 26 and link arm 25.

As can be seen from the angular position of control cam 18 shown in FIGS. 3A-3B, by way of revolving motion of the center of control cam 18 around the center of control shaft 17, the radially thick-walled portion of control cam 18 shifts upwards apart from drive shaft 6 and is held at the upwardly shifted position, with the result that the pivot (the connected point by connecting pin 15) between the second arm portion 11b of rocker arm 11 and the first rod end 13a of link rod 13 also shifts upwards with respect to drive shaft 6. As a result, the cam-nose portion of each of rockable cams 9, 9 is forcibly pulled up via the second rod end 13b of link rod 13. As viewed from the front end of drive shaft 6, the angular position of each rockable cam 9 shown in FIGS. 3A-3B is relatively shifted to the clockwise direction from the angular position of each rockable cam 9 shown in FIGS. 4A-4B.

With control cam 18 held at the angular position shown in FIGS. 3A-3B, when drive cam 7 is rotated, a rotary motion of drive cam 7 is converted through link arm 12, the first arm portion 11a of rocker arm 11, the second arm portion 11b of rocker arm 11, and link rod 13 into an oscillating motion of rockable cam 9, but almost the base-circle surface area of rockable cam 9 is brought into sliding-contact with the upper contact surface of valve lifter 8 (see FIGS. 3A-3B). Thus, the actual intake-valve lift becomes a small lift L1 and simultaneously the actual intake-valve working angle becomes a small working angle D1 (see the small intake-valve lift L1 and small working angle D1 characteristic shown in FIG. 5).

Actually, a valve clearance exists between rockable cam 9 and valve lifter 8. Thus, a valve lift amount of intake valve 4 becomes less than a cam lift amount by the valve clearance, and therefore a lifted period from valve open timing to valve closure timing, fully taking into account the valve clearance, can be regarded as a working angle.

When shifting from the previously-noted low-speed low-load range to another engine operating condition such as mid-speed and mid-load operation, motor 20 is rotated in a reverse-rotational direction responsively to a control signal, which is generated from the output interface circuitry of ECU 22. Thus, ball-screw shaft 23 is also rotated in the reverse-rotational direction by reverse-rotation of the motor output shaft of motor 20, thereby producing the opposite rectilinear motion of ball nut 24 against the bias of coil spring 30. As a result, control shaft 17 rotates in the opposite rotation direction via the linkage (25, 26).

By way of revolving motion of the center of control cam 18 around the center of control shaft 17, the radially thick-walled portion of control cam 18 slightly downwardly shifts toward drive shaft 6 and is held at the slightly downwardly shifted position. Thus, the attitude of rocker arm 11 slightly shifts anticlockwise from the angular position of rocker arm 11 shown in FIGS. 3A-3B, with the result that the pivot (the connected point by connecting pin 15) between the second arm portion lib of rocker arm 11 and the first rod end 13a of link rod 13 also shifts slightly downwards. As a result, the cam-nose portion of each of rockable cams 9, 9 is forcibly slightly pushed down via the second rod end 13b of link rod 13. As viewed from the front end of drive shaft 6, the angular position of each rockable cam 9 is relatively shifted to the anticlockwise direction from the angular position of each rockable cam 9 shown in FIGS. 3A-3B.

With control cam 18 shifted from the angular position shown in FIGS. 3A-3B to the intermediate angular position located in a substantially middle of the angular position shown in FIGS. 3A-3B and the angular position shown in FIGS. 4A-4B, during rotation of drive cam 7, a rotary motion of drive cam 7 is converted through link arm 12, the first arm portion 11a of rocker arm 11, the second arm portion 11b of rocker arm 11, and link rod 13 into an oscillating motion of rockable cam 9. At this time, a part of the base-circle surface area, the ramp surface area, the lift surface area, and the top surface area are brought into sliding-contact with the upper contact surface of valve lifter 8. Thus, when varying from the angular position of control cam 18 shown in FIGS. 3A-3B to the intermediate angular position, the actual intake-valve lift and working angle characteristic can be quickly varied from the small intake-valve lift L1 and small working angle D1 characteristic to a middle intake-valve lift L2 and middle working angle D2 characteristic (see FIG. 5). That is, intake-valve working angle as well as intake-valve lift can be simultaneously increased. Owing to a valve lift increase (L1→L2) and a working angle increase (D1→D2), intake valve closure timing IVC of each of intake valves 4, 4 is phase-retarded and controlled to a timing value near BDC. Thus, an effective compression ratio becomes high to ensure good combustion. Additionally, a charging efficiency of fresh air tends to become high, thus resulting in an increase in torque generated by combustion.

After this, when the engine/vehicle operating condition is shifting from the low or middle load range to a high-speed high-load range, motor 20 is further driven in the reverse-rotational direction responsively to a control signal, which is generated from the output interface circuitry of ECU 22 and determined based on the high engine load condition. Thus, ball-screw shaft 23 is further rotated in the reverse-rotational direction by reverse-rotation of the motor output shaft of motor 20, thereby producing the further opposite rectilinear motion of ball nut 24. As a result, control shaft 17 further rotates in the opposite rotation direction via the linkage (25, 26). By way of further revolving motion of the center of control cam 18 around the center of control shaft 17, the radially thick-walled portion of control cam 18 further shifts downwards and is held at the downwardly shifted position. Thus, the attitude of rocker arm 11 further shifts anticlockwise, with the result that the pivot (the connected point by connecting pin 15) between the second arm portion 11b of rocker arm 11 and the first rod end 13a of link rod 13 further shifts downwards. As a result, the cam-nose portion of each of rockable cams 9, 9 is further forcibly pushed down via the second rod end 13b of link rod 13. As viewed from the front end of drive shaft 6, the angular position of each rockable cam 9 is further shifted anticlockwise. With control cam 18 shifted to the angular position (suited to high load operation) shown in FIGS. 4A-4B, during rotation of drive cam 7, a rotary motion of drive cam 7 is converted through the motion transmitting mechanism (links 11, 12, and 13) into an oscillating motion of rockable cam 9. At this time, a part of the base-circle surface area, the ramp surface area, the lift surface area, and the top surface area are brought into sliding-contact with the upper contact surface of valve lifter 8. Thus, when switching from the intermediate angular position (suited to middle load operation) of control cam 18 to the angular position (suited to high load operation) shown in FIGS. 4A-4B, the actual intake-valve lift and working angle characteristic can be continuously varied from the middle intake-valve lift L2 and middle working angle D2 characteristic to a large intake-valve lift L3 and large working angle D3 characteristic (see FIG. 5). As a result, in a high-speed high-load range, a charging efficiency of fresh air tends to become higher, thus more greatly enhancing engine power output.

As can be appreciated from a plurality of intake-valve lift L and intake-valve working angle D characteristic curves (or a plurality of intake-valve lift L and lifted-period D characteristic curves) shown in FIG. 5, according to VEL mechanism 1 incorporated in the internal combustion engine control system of the embodiment, through all engine operating conditions from low-speed low-load operation to high-speed high-load operation, the intake-valve lift and working angle characteristic can be continuously controlled or adjusted from the small intake-valve lift L1 and working angle D1 characteristic via the middle intake-valve lift L2 and working angle D2 characteristic to the large intake-valve lift L3 and working angle D3 characteristic, or vice versa. That is to say, the intake-valve lift and working angle characteristic can be controlled or adjusted to an optimal characteristic suited to the latest up-to-date information concerning engine operating condition.

During an engine stopping period, as previously described, ball nut 24 of ball-screw mechanism 21 is forced and stably held by the spring bias of coil spring 30 in the axial direction of ball-screw shaft 23 corresponding to the small lift L1 and small working angle D1 range. Hence, it is possible to reduce a frictional loss in the valve operating system, thus ensuring a good startability (or a good restartability).

As discussed above, as best seen from the variable intake-valve lift and event (working angle) and phase characteristic diagram of FIG. 5, a slight valve clearance A exists between the base-circle surface of rockable cam 9 oscillating and the valve lifter 8. Hence, the effective valve lift amount L tends to be reduced from the cam lift amount of the rockable cam by the valve clearance A, and thus the effective working angle D also tends to be reduced slightly. The previously-discussed valve lift amounts L1 to L3 and working angles D1 to D3 represent the effective valve lifts and working angles, from which the valve clearance A is excepted.

Hereunder described briefly in reference to FIGS. 6, 7, and 8, is the construction of intake-valve VTC mechanism 2. As can be appreciated from the cross sections of FIGS. 6-8, VTC mechanism 2 comprises a so-called hydraulically-operated rotary vane type VTC mechanism. Intake-valve VTC mechanism 2 is comprised of timing sprocket 33 driven by crankshaft 02 and fixedly connected to drive shaft 6 for torque transmission, a four-blade vane member 32 fixedly connected or bolted to the shaft end of drive shaft 6 and rotatably accommodated in the internal space of timing sprocket 33, and a hydraulic circuit, which hydraulically operates vane member 32 in a manner so as to rotate vane member 32 in selected one of normal-rotational and reverse-rotational directions.

Timing sprocket 33 is comprised of a substantially cylindrical, phase-converter housing 34 rotatably accommodating therein vane member 32, a disk-shaped front cover 35 hermetically covering the front opening end of housing 34, and a disk-shaped rear cover 36 hermetically covering the rear opening end of housing 34. Housing 34 and front and rear covers 35-36 are axially connected integral with each other by tightening four bolts 37.

Housing 34 is substantially cylindrical in shape and opened at both axial ends. Housing 34 has four shoes 34a, 34a, 34a, 34a evenly spaced around its entire circumference and serving as four partition walls radially inwardly extending from the inner periphery of the housing.

Each of shoes 34a is trapezoidal in shape, and has an axially-extending bolt insertion hole 34b formed in its substantially central portion such that bolt 37 is inserted into the bolt insertion hole. As best seen in FIG. 7, each of shoes 34a has an axially-elongated seal groove formed in its apex. Four elongated oil seals 38, 38, 38, 38 each having a substantially C-shape in lateral cross section, are fitted into and retained in the respective seal grooves of shoes 34a. Although it is not clearly shown in FIG. 7, actually, four leaf springs are fitted into and retained in the respective seal grooves of shoes 34a in such a manner as to radially inwardly force the respective oil seals 38 against the outer peripheral wall surface of a vane rotor 32a (described later).

The previously-noted disk-shaped front cover 35 has a comparatively large-diameter center supporting bore 35a and circumferentially equidistant-spaced bolt holes (not numbered) bored to axially conform to the respective bolt insertion holes 34b of shoes 34a of housing 34.

The previously-noted disk-shaped rear cover 36 is integrally formed at its rear end with a toothed portion 36a, which is in meshed-engagement with the timing chain. Also, rear cover 36 has a substantially center bearing bore 36b having a comparatively large diameter.

Vane member 32 is comprised of a substantially annular ring-shaped vane rotor 32a formed with a center bolt insertion hole and radially-extending four vanes or blades 32b, 32b, 32b, 32b evenly spaced around the entire circumference of vane rotor 32a and integrally formed on the outer periphery of vane rotor 32a.

A small-diameter, cylindrical-hollow front end portion of vane rotor 32a is rotatably supported in the center bore 35a of front cover 35. A small-diameter, cylindrical-hollow rear end portion of vane rotor 32a is also rotatably supported in the bearing bore 36b of rear cover 36.

Vane rotor 32a of vane member 32 has an axially-extending central bore 14a into which a vane mounting bolt 39b is inserted for bolting vane member 32 to the front axial end of drive shaft 6 by axially tightening vane mounting bolt 39b.

One of four vane blades 32b, 32b, 32b, 32b is configured to have an inverted trapezoidal shape in lateral cross section, whereas the remaining three vane blades are configured to be substantially rectangular in lateral cross section. The remaining three blades have almost the same circumferential width and the same radial length. The circumferential width of the one blade having the inverted trapezoidal shape is dimensioned to be greater than that of each of the remaining three rectangular blades, taking account of total weight balance of vane member 32, in other words, reduced rotational unbalance of vane member 32 having four blades 32b.

Each of four blades 32b, 32b, 32b, 32b is disposed in an internal space defined between the associated two adjacent shoes 34a and 34a. As best seen in FIG. 7, four apex seals 40, 40, 40, and 40, each being substantially C-shaped in lateral cross section, are fitted into and retained in respective seal grooves formed in apexes of four blades 32b, so that each of blades 32b is slidable along the inner peripheral wall surface of phase-converter housing 34. Although it is not clearly shown in FIG. 7, actually, four leaf springs are fitted into and retained in the respective seal grooves of the apexes of blades 32b in such a manner as to radially inwardly force the respective apex seals 40 against the inner peripheral wall surface of housing 34. The backward sidewall surface of each blade 32b, opposing to the rotational direction of drive shaft 6, is formed with substantially circular, two concave grooves 32c and 32c, which serve as spring retaining holes for two rows of return springs 55-56. Return springs 55-56 are disposed between the spring-retaining-hole equipped backward sidewall surface of blade 32b and a spring-retaining sidewall surface of shoe 34a opposing to the backward sidewall surface of blade 32b.

Four blades 32b of vane member 32 and four shoes 34a of housing 34 cooperate with each other to define four variable-volume phase-advance chambers 41 and four variable-volume phase-retard chambers 42. In more detail, each of phase-advance chambers 41 is defined between the spring-retaining-hole equipped backward sidewall surface of blade 32b and the opposing spring-retaining sidewall surface of shoe 34a. Each of phase-retard chambers 42 is defined between the non-spring-retaining-hole equipped forward sidewall surface of blade 32b and the opposing non-spring-retaining sidewall surface of shoe 34a.

As clearly shown in FIG. 6, the previously-noted hydraulic circuit is comprised of a first hydraulic line 43 provided to supply and exhaust working fluid (hydraulic pressure) to and from each of phase-advance chambers 41, and a second hydraulic line 44 provided to supply and exhaust working fluid (hydraulic pressure) to and from each of phase-retard chambers 42. That is, the hydraulic circuit comprises a dual hydraulic line system (43, 44). Each of hydraulic lines 43 and 44 are connected through an electromagnetic solenoid-operated directional control valve 47 to a working-fluid supply passage 45 and a working-fluid drain passage 46. A one-way oil pump 49 is disposed in supply passage 45 for sucking working fluid in an oil pan 48 and for discharging the pressurized working fluid from its discharge port. The downstream end of drain passage 46 communicates oil pan 48.

First and second hydraulic lines 43 and 44 are formed in a substantially cylindrical flow-passage structure 39. One end (i.e., a first end) of flow-passage structure 39 is inserted through the left-hand axial opening end of the small-diameter, cylindrical-hollow front end portion of vane rotor 32a into a cylindrical bore 32d formed in vane rotor 32a. The other end (i.e., a second end) of flow-passage structure 39 is connected to electromagnetic solenoid-operated directional control valve 47. Three annular seals 39s, 39s, 39s are disposed between the outer periphery of the first end of flow-passage structure 39 and the inner periphery of cylindrical bore 32d of vane rotor 32a. In more detail, annular seals 39s are fitted into and retained in respective seal grooves formed in the outer periphery of the first end of flow-passage structure 39. These annular seals 39s act to partition between a phase-advance-chamber communication port of first hydraulic line 43 and a phase-retard-chamber communication port of second hydraulic line 44 in a fluid-tight fashion.

First hydraulic line 43 is further provided with a working-fluid chamber 43a and four branch passages 43b, 43b, 43b, 43b. First hydraulic line 43 penetrates through the first end face of flow-passage structure 39, and the axial passage of first hydraulic line 43 communicates working-fluid chamber 43a. Working-fluid chamber 43a is formed as the inner half of cylindrical bore 32d of vane rotor 32a, facing drive shaft 6. Four branch passages 43b are formed in vane rotor 32a in such a manner as to substantially radially extend from the inner periphery of cylindrical bore 32d. Four phase-advance chambers 41 are communicated with working-fluid chamber 43a via respective branch passages 43b.

On the other hand, the axial passage of second hydraulic line 44 extends near the first end face of flow-passage structure 39. Second hydraulic line 44 is further provided with an annular chamber 44a and a second working-fluid passage 44b. Annular chamber 44a is formed in the outer periphery of the cylindrical portion of the first end of flow-passages structure 39. Although it is not clearly shown in the drawing, second working-fluid passage 44b has a substantially L shape and formed in vane rotor 32a. Annular chamber 44a and each of phase-retard chambers 42 are communicated with each other via second working-fluid passage 44b.

In the shown embodiment, electromagnetic solenoid-operated directional control valve 47 is constructed by a four-port, three-position, spring-offset solenoid-actuated directional control valve. Directional control valve 47 uses a sliding valve spool to change the path of flow through the directional control valve. For a given position of the valve spool, a unique flow path configuration exists within the valve. Concretely, directional control valve 47 is designed to switch among three positions of the spool, namely a spring-offset position shown in FIG. 6, a block-off position (a center position created due to the balancing opposing forces, that is, the return spring force and the electromagnetic force produced by the solenoid), and a fully solenoid-actuated position. In the spring-offset position, fluid communication between first hydraulic line 43 and drain passage 46 is established, and fluid communication between second hydraulic line 44 and supply passage 45 is established. In the block-off position, fluid communication between each of first and second hydraulic lines 43-44 and each of supply passage 45 and drain passage 46 is blocked. In the fully solenoid-actuated position, fluid communication between first hydraulic line 43 and supply passage 45 is established, and fluid communication between second hydraulic line 44 and drain passage 46 is established. Switching operation among the three positions of the valve spool of directional control valve 47 is executed responsively to a control command signal generated from the output interface circuitry of ECU 22 to the solenoid.

The controller (ECU) 22 is common to both of VEL mechanism 1 and VTC mechanism 2.

Also provided is a lock mechanism (or an interlocking device or interlocking means) disposed between vane member 32 and housing 34, for disabling rotary motion of vane member 32 relative to housing 34 by locking and engaging vane member 32 with housing 34, and for enabling rotary motion of vane member 32 relative to housing 34 by unlocking (or disengaging) vane member 32 from housing 34. That is, by the interlocking means, intake valve closure timing IVC of each of intake valves 4, 4 can be locked or fixed to a predetermined timing value.

As can be seen from the longitudinal cross section of FIG. 6, the lock mechanism (interlocking means) is comprised of a lock-pin sliding-motion permitting bore (simply, a lock-pin bore) 50, a lock pin 51, an engaging-hole structural member 52 having a substantially C shape in lateral cross section and press-fitted into a through hole formed in rear cover 36, an engaging hole 52a defined in the C-shaped engaging-hole structural member 52, a spring retainer 53, and a return spring (a coiled compression spring) 54. Lock-pin bore 50 is formed in the inverted trapezoidal blade 32b of the relatively greater circumferential width (the maximum circumferential width) and formed in rear cover 36, such that lock-pin bore 50 extends in the axial direction of drive shaft 6. Lock pin 51 is slidably accommodated in lock-pin bore 50 and has a cylindrical bore closed at one end. A tapered head portion 51a of lock pin 51 is engaged with or disengaged from engaging hole 52a. Spring retainer 53 is fitted into a space defined by the inner peripheral wall surface of front cover 35 and lock-pin bore 51. Return spring 54 is provided to permanently force lock pin 51 toward the internal space of engaging hole 52a. Although it is not clearly shown in FIG. 6, the phase-converter housing structure, constructed by front and rear covers 35-36 and cylindrical housing 34, is also designed to supply working oil (hydraulic pressure) in phase-retard chamber 42 and/or working oil (hydraulic pressure) discharged from oil pump 49 via an oil hole formed in the phase-converter housing structure into engaging hole 52a.

Lock pin 51 operates to disable relative rotation between timing sprocket 33 and drive shaft 6 by locking and engaging tapered head portion 51a of lock pin 51 with engaging hole 52a in a predetermined position where vane member 32 reaches its maximum phase-retard position, by way of the spring force of return spring 54. Relative rotation between timing sprocket 33 and drive shaft 6 is enabled by unlocking (or disengaging) tapered head portion 51a of lock pin 51 from engaging hole 52a by way of the hydraulic pressure delivered from phase-retard chamber 42 and/or oil pump 49 into engaging hole 52a. That is, tapered head portion 51a of lock pin 51 is forced out of engaging hole 52a under hydraulic pressure fed into the engaging hole from phase-retard chamber 42 and/or oil pump 49.

As previously described with reference to FIG. 7, two rows of return springs 55-56, each of which serves as a biasing device or biasing means, are disposed between the spring-retaining-hole equipped backward sidewall surface of blade 32b and the spring-retaining sidewall surface of shoe 34a, for permanently biasing the associated blade 32b (vane member 32) toward the phase-retard side. In the shown embodiment, return springs 55-56 are constructed by coil springs having the same size and the same spring stiffness.

Although, in FIGS. 7-8, two return springs 55-56 are illustrated to be overlapped with each other, actually, two return springs 55-56 are disposed in parallel with each other. As can be seen from the lateral cross section of FIG. 7, the axial length of each of springs 55-56 is dimensioned to be greater than the circumferential distance between the spring-retaining-hole equipped backward sidewall surface of blade 32b and the spring-retaining sidewall surface of shoe 34a with the blade 32b held at the maximum phase-retard position. Return springs (coil springs) 55-56 have the same free height.

The distance between the axes of two parallel coil springs 55-56 is preset to a predetermined distance that the outer peripheries of coil springs 55-56 are not brought into contact with each other under a condition of maximum compressive deformation of each of coil springs 55-56 (see FIG. 8). One end of each of coil springs 55-56, facing the associated blade 32b, is retained in a thin-plate spring retainer (not shown) fitted to concave groove (spring retaining hole) 32c.

Hereinafter described in detail is the basic operation of intake-valve VTC mechanism 2. First, during an engine stopped period, the output of control current (exciting current) from ECU 22 to the solenoid of directional control valve 47 is stopped. Thus, by means of springs 55-56, the valve spool of directional control valve 47 is mechanically shifted to its spring-offset position (a default position) at which fluid communication between first hydraulic line 43 and drain passage 46 is established, and simultaneously fluid communication between second hydraulic line 44 and supply passage 45 is established. Also, when the engine is in its stopped state, the discharge pressure from oil pump 49 becomes zero and thus there is no hydraulic pressure supplied to the VTC mechanism 2.

Under these conditions, as shown in FIG. 7, vane member 32 is forced toward the maximum phase-retard side by the spring forces of return springs 55-56, such that the inverted trapezoidal vane blade 32b of the maximum circumferential width is brought into abutted-engagement with the sidewall of shoe 34a facing phase-advance chamber 41. At the same time, tapered head portion 51a of lock pin 51 of the lock mechanism is engaged with the engaging hole 52a, and thus vane member 32 can be stably held at the maximum phase-retard position. That is, with directional control valve 47 kept at its default position (the spring-offset position), intake-valve VTC mechanism 2 can be stably held at the maximum phase-retard position mechanically by means of the lock mechanism (50, 51, 52a, 54) and return springs 55-56.

Hereunder described briefly is the operation of intake-valve VTC mechanism 2 under several engine operating conditions. First, during an engine starting period, with the ignition switch turned ON, electric motor 07 is driven to initiate cranking action for crankshaft 02. At the early stage of cranking, the output interface circuitry of ECU 22 begins to generate a control command signal to the solenoid of directional control valve 47, but the hydraulic pressure of working fluid discharged from oil pump 49 does not yet rise adequately just after the engine has been started. Thus, vane member 32 is still held at the maximum phase-retard position by means of the lock mechanism (50, 51, 52a, 54) and return springs 55-56.

At this time, the solenoid of directional control valve 47 is held at its default position (the spring-offset position) responsively to a control signal from ECU 22, such that fluid communication between second hydraulic line 44 and supply passage 45 is established and fluid communication between first hydraulic line 43 and drain passage 46 is established. Under these conditions, on the one hand, owing to a gradual hydraulic pressure rise, hydraulic pressure produced by oil pump 49 is supplied through supply passage 45 and second hydraulic line 44 into each of phase-retard chambers 42. On the other hand, there is no supply of hydraulic pressure to each of phase-advance chambers 41 in the same manner as the engine stopped state. That is, hydraulic pressure is relieved from each of phase-advance chambers 41 through first hydraulic line 43 and drain passage 46 into oil pan 48 and thus the hydraulic pressure in each of phase-advance chambers 41 is kept low.

After the hydraulic pressure produced by oil pump 49 has risen adequately, the variable phase control system (intake-valve VTC mechanism 2) enables rapid and accurate vane position control via directional control valve 47. In more detail, working fluid, supplied into phase-retard chamber 42, is also delivered from phase-retard chamber 42 into engaging hole 52a, and thus owing to a hydraulic pressure rise of phase-retard chamber 42 a hydraulic pressure in engaging hole 52a also rises. As a result, lock pin 51 moves backwards against the spring bias of return spring 54 and then tapered head portion 51a of lock pin 51 is forced out of engaging hole 52a, so as to permit relative rotation between phase-converter housing 34 and vane member 32, and thus to enable rapid and accurate vane position control.

For instance, in an engine idling state after the internal combustion engine has been warmed up, the solenoid of directional control valve 47 is also held at its default position responsively to a control signal from ECU 22, such that fluid communication between second hydraulic line 44 and supply passage 45 is established and fluid communication between first hydraulic line 43 and drain passage 46 is established. Therefore, owing to a rise in hydraulic pressure supplied to each of phase-retard chambers 42, vane member 32 is maintained at the relative position shown in FIG. 7 by the aid of the spring forces of return springs 55-56. Thus, the angular phase of drive shaft 6 relative to timing sprocket 33 is also maintained at the phase-retard side.

Thereafter, when the engine operating condition is shifted to a low-speed middle-load range, the solenoid of directional control valve 47 is shifted to its fully solenoid-actuated position responsively to a control signal from ECU 22, such that fluid communication between first hydraulic line 43 and supply passage 45 is established and fluid communication between second hydraulic line 44 and drain passage 46 is established.

Hence, hydraulic pressure in each of phase-retard chambers 42 is returned through second hydraulic line 44 and drain passage 46 into oil pan 48, and thus the hydraulic pressure in each of phase-retard chambers 42 becomes low, whereas the hydraulic pressure in each of phase-advance chambers 41 becomes high.

Therefore, owing to a rise in hydraulic pressure supplied to each of phase-advance chambers 41, vane member 32 is rotated clockwise to the relative position shown in FIG. 8 against the spring forces of return springs 55-56. Thus, the angular phase of drive shaft 6 relative to timing sprocket 33 is converted to the phase-advance side. By switching the axial position of the spool of directional control valve 47 to the block-off position (the center position created due to the balancing opposing forces, that is, the return spring force and the electromagnetic force produced by the solenoid) during rotary motion of drive shaft 6 relative to timing sprocket 33, it is possible to hold the angular phase of drive shaft 6 relative to timing sprocket 33 at an arbitrary relative phase angle.

Furthermore, when the engine operating condition is shifted from a low-speed range to a normal middle-speed range, and further shifted to a high-speed range, in the same manner as the engine idling state after engine warm-up, the solenoid of directional control valve 47 is controlled to the default position. As a result, a fall in hydraulic pressure in each of phase-advance chambers 41 occurs and simultaneously a rise in hydraulic pressure supplied to each of phase-retard chambers 42 occurs. By a combined force of the supplied hydraulic pressure and the spring forces of return springs 55-56, the angular phase of drive shaft 6 relative to timing sprocket 33 is converted to the phase-retard side (see FIG. 7).

Prior to detailed explanation of control actions executed by ECU 22, the relationship among crankangle of crankshaft 02, intake-valve open timing IVO, and intake-valve closure timing IVC, obtained by the variable valve actuation device of the first embodiment, at each of #1, #3, #4, and #2 engine cylinders, are hereunder described in detail in reference to the characteristic diagrams of FIGS. 9A-9D. Hereupon, in the shown embodiment, the firing order of the four engine cylinders of the internal combustion engine is #1→#3→#4→#2. The characteristic diagrams of FIGS. 9A-9D are based on the assumption that the internal combustion engine is in a stopped state, and thus a central phase angle (e.g., a phase at a peak valve lift) of a valve lift characteristic curve of each of intake valves 4, 4 is stably kept at the maximum phase-retard side (at the default valve timing) by means of intake-valve VTC mechanism 2.

Now, suppose that control shaft 17 of intake-valve VEL mechanism 1 is rotated to control the lift and working angle characteristic of each of intake valves 4, 4 to the large intake-valve lift L3 and large working angle D3 characteristic and also the angular position of crankshaft 02 reaches a crankangle near the TDC position on compression stroke at the #1 cylinder. At this time, as a matter of course, each of intake valves 4, 4 of the #1 cylinder does not open (see FIG. 9A), but each of intake valves 4, 4 (controlled to the large intake-valve lift L3 and large working angle D3 characteristic) of the #3 cylinder of the next cycle opens (see FIG. 9B) and also each of intake valves 4, 4 of the #4 cylinder of the next but one cycle opens (see FIG. 9C). That is, intake valves 4, 4 of the two different engine cylinders, namely, the #3 cylinder and the #4 cylinder, simultaneously open.

Next, check for a specific state where intake valves 4, 4 of all engine cylinders simultaneously close at different degrees of crankshaft rotation. As can be seen from the characteristic diagrams of FIGS. 9A-9D, such a specific state does not exist. In other words, a crankangle area, in which (i) a lift and working angle characteristic curve of each of intake valves 4, 4 of a first cylinder selected from the four engine cylinders and (ii) a lift and working angle characteristic curve of each of intake valves 4, 4 of a second cylinder selected from the four engine cylinders can be partially overlapped with each other, exists.

Even as for all crankangles, each of intake valves 4, 4 of at least one engine cylinder is kept open, and thus a spring reaction force of valve springs 5, 5 acts on control cam 18. Under these conditions, when actuating control shaft 17 by motor 20 (the control-shaft actuator) through ball-screw mechanism 21 with the internal combustion engine kept in its stopped state, control shaft 17 cannot be smoothly rotated due to the spring reaction force acting on control cam 18 and a large static friction coefficient of the sliding-contact portion between control cam 18 and rocker arm 11.

Next, suppose that control shaft 17 of intake-valve VEL mechanism 1 is rotated to control the lift and working angle characteristic of each of intake valves 4, 4 to the middle intake-valve lift L2 and middle working angle D2 characteristic. In a similar manner to a case that the intake-valve lift and working angle characteristic is controlled to the large intake-valve lift L3 and large working angle D3 characteristic, a crankangle area, in which (i) an intake-valve lift and working angle characteristic curve of a first cylinder selected from the four engine cylinders and (ii) an intake-valve lift and working angle characteristic curve of a second cylinder selected from the four engine cylinders can be partially overlapped with each other, exists (see the fine solid lines of the characteristic curves illustrated in FIGS. 9A-9D). Even as for all crankangles, each of intake valves 4, 4 of at least one engine cylinder is kept open, and thus a spring reaction force of valve springs 5, 5 acts on control cam 18. Under these conditions, when actuating control shaft 17 by motor 20 (the control-shaft actuator) through ball-screw mechanism 21 with the internal combustion engine kept in its stopped state, control shaft 17 cannot be smoothly rotated due to the spring reaction force acting on control cam 18 and a large static friction coefficient of the sliding-contact portion between control cam 18 and rocker arm 11, even in the case of the middle working angle D2 characteristic, in a similar manner to the large working angle D3 characteristic.

Next, suppose that control shaft 17 of intake-valve VEL mechanism 1 is rotated to control the lift and working angle characteristic of each of intake valves 4, 4 to the small intake-valve lift L1 and small working angle D1 characteristic. Assuming that the angular position of crankshaft 02 reaches a crankangle corresponding to a point “A” indicated by the asterisk in FIGS. 9A-9D, each of intake valves 4, 4 of the #1, #4, and #2 cylinders does not open (see FIGS. 9A, 9C, and 9D), but each of intake valves 4, 4 (controlled to the small intake-valve lift L1 and small working angle D1 characteristic) of the #3 cylinder opens (see FIG. 9B). Thus, a spring reaction force of valve springs 5, 5 acts on control cam 18. Under these conditions, when actuating control shaft 17 by motor 20 through ball-screw mechanism 21 with the internal combustion engine kept in its stopped state, control shaft 17 cannot be smoothly rotated due to the spring reaction force and a large static friction coefficient of the sliding-contact portion between control cam 18 and rocker arm 11, even in the case that the intake-valve lift and working angle characteristic is controlled to the small intake-valve lift L1 and small working angle D1 characteristic.

Notice that, in the case that the intake-valve lift and working angle characteristic has been controlled to the small intake-valve lift L1 and small working angle D1 characteristic, a specific crankangle area (i.e., a crankangle area α1, a crankangle area α2, a crankangle area α3, and a crankangle area α4), in which (i) a lift and working angle characteristic curve of each of intake valves 4, 4 of one cylinder selected from the four engine cylinders and (ii) a lift and working angle characteristic curve of each of intake valves 4, 4 of the other cylinder selected from the four engine cylinders cannot be overlapped with each other, exists. Within the specific crankangle area (α1, α2, α3, α4), intake valves 4, 4 of all engine cylinders simultaneously close. Therefore, assuming that the angular position of crankshaft 02 reaches a specific crankangle corresponding to a point “B” indicated by the asterisk in FIGS. 9A-9D and included within the specific crankangle area α1, there is a less valve-spring reaction force acting on control cam 18 (or control shaft 17), thus ensuring a smooth rotary motion of control cam 18. As soon as control cam 18 begins to smoothly rotate, a friction coefficient of the sliding-contact portion between control cam 18 and rocker arm 11 changes from a large static friction coefficient to a small kinetic friction coefficient. Hence, in the case that the angular position of crankshaft 02 has been controlled to a crankangle included within the specific crankangle area (α1, α2, α3, α4), it is possible to smoothly change or convert the intake-valve lift and event (working angle) characteristic from the small lift L1 and working angle D1 characteristic to the middle lift L2 and working angle D2 characteristic, and further to smoothly change or convert to the large lift L3 and working angle D3 characteristic. Actually, even when the angular position of crankshaft 02 has been adjusted to a specific crankangle corresponding to the point “B” indicated by the asterisk in FIGS. 9A-9D and included within the specific crankangle area α1, according to the progress of a change (or a conversion) in the intake-valve lift and working angle characteristic to either the middle lift L2 and working angle D2 characteristic or the large lift L3 and working angle D3 characteristic, of all four engine cylinders, intake valves 4, 4 of some cylinders (two cylinders in the shown embodiment) begin to open. At this time, by virtue of a transition of the friction coefficient of the sliding-contact portion of control cam 18 and rocker arm 11 to a small kinetic friction coefficient and rotational inertia of control shaft 17, which begins to already rotate, it is possible to maintain and ensure a good valve lift and event conversion responsiveness of intake-valve VEL mechanism 1.

That is, the engine control system of the first embodiment is configured, so that, during an internal combustion engine stopping period, a crankshaft-rotation stopped position of crankshaft 02 is adjusted to a crankangle included within the specific crankangle area (α1, α2, α3, α4) by means of electric motor 07. At this time, the intake-valve lift and event (working angle) characteristic is controlled to the small lift L1 and working angle D1 characteristic.

On the other hand, during an engine restarting period, a control signal is outputted to motor 20 of intake-valve VEL mechanism 1 for the purpose of converting to a desired working angle by means of the attitude control mechanism (including control shaft 17 and control cam 18) before initiating cranking action of crankshaft 02. Hereby, a valve lift and event conversion to the desired working angle can be initiated before cranking the engine. Thus, it is possible to improve a valve lift and event conversion responsiveness of intake-valve VEL mechanism 1 (by virtue of a transition of the friction coefficient of the sliding-contact portion between control cam 18 and rocker arm 11 from a large static friction coefficient to a small kinetic friction coefficient), and also to shorten a converting time (a response time) required for conversion to the desired working angle (by virtue of rotational inertia of control shaft 17 as well as a transition of the friction coefficient of the sliding-contact portion between control cam 18 and rocker arm 11 to a small kinetic friction coefficient).

By the way, a desired working angle, required for each of intake valves 4, 4 during an engine starting period, is different depending on a condition on the engine or vehicle, such as an engine temperature. For instance, when the engine temperature is very low, intake-valve closure timing IVC has to approach closer to a timing value near a piston bottom dead center (BDC) position in order to ensure better combustion. In such a case, a middle working angle D2 is selected as a desired working angle.

In contrast to the above, when the engine temperature is high, in order to suppress preignition and starting-period vibrations, a large working angle (a maximum working angle) D3 is selected as a desired working angle. When the large working angle D3 is selected, intake-valve closure timing IVC is greatly phase-retarded with respect to the BDC position, and thus the temporarily drawn-in fresh air is discharged from the combustion chamber to the intake port, thereby reducing an effective compression ratio. Such a decompressing action contributes to a suppression in preignition and/or a reduction in noise and vibrations when starting the engine.

During a general engine restarting period that the engine temperature is neither very low nor high, a small working angle (a minimum working angle) D1 is selected as a desired working angle. When the minimum working angle D1 is selected, the actual intake-valve lift becomes a small lift L1 and simultaneously the actual intake-valve working angle becomes a small working angle D1, and thus it is possible to reduce a frictional loss in the valve operating system and thus to ensure a smooth engine speed rise, thereby realizing a good restartability (a smooth and rapid engine restart).

The minimum working angle D1 contributes to a reduced effective compression ratio (i.e., a decompressing effect), but leads to a drawback that preignition can be somewhat promoted owing to an intake-air mixing effect, caused by a phase-retarded intake-valve open timing IVO. For the reasons discussed above, in the case of high engine temperatures (high engine oil temperatures), the maximum working angle D3 is superior to the minimum working angle D1.

As discussed above, during an engine stopping period, a crankangle is preset or pre-adjusted within an all-cylinder valve closed period, in other words, an all-cylinder intake-valve-closed crankangle area (α1, α2, α3, α4), in which intake valves 4, 4 of all cylinders are kept in their non-lifted states, by means of electric motor 07 of the crank position change mechanism. During an engine restarting period, prior to cranking action of crankshaft 02, a control signal is outputted to intake-valve VEL mechanism 1 so as to achieve a desired working angle (an engine-start desired working angle Dt described later), determined based on an engine condition (or an engine/vehicle condition) such as engine temperature (concretely, engine oil temperature or engine coolant temperature). Therefore, it is possible to shorten a converting time required for conversion to the desired working angle.

Details of the control flow executed by ECU 22 are hereunder described in reference to the flowchart of FIG. 10.

First, at step S1, a check is made to determine, based on the current engine/vehicle conditions, whether an engine-stop condition is satisfied. In other words, a check is made to determine whether a necessary condition that the ignition switch should be turned OFF is satisfied. In the case of an automatic-engine-stop-restart system equipped hybrid vehicle, a check is made to determine whether a necessary condition that the engine should be automatically stopped is satisfied. When the answer to step S1 is in the negative (NO), one execution cycle terminates without any control action. Conversely when the answer to step S1 is in the affirmative (YES), that is, when the engine-stop condition has been satisfied, step S2 occurs.

At step S2, a control signal is outputted from ECU 22 to the control-shaft actuator (motor 20) of intake-valve VEL mechanism 1 in such a manner as to switch the operating mode of each of intake valves to the minimum working angle D1 operating mode (i.e., the small intake-valve lift L1 and working angle D1 characteristic).

At step S3, a control signal is outputted from ECU 22 to electric motor 07 included in the crank position change mechanism in such a manner as to control or adjust the angular position of crankshaft 02 to a crankangle included within the previously-discussed all-cylinder valve closed period (i.e., the all-cylinder intake-valve-closed crankangle area (α1, α2, α3, α4), for example the crankangle area α1).

At step S4, a check is made to determine whether the actual working angle of each of intake valves 4, 4 has been controlled to the minimum working angle D1 by means of intake-valve VEL mechanism 1 and the actual crankangle of crankshaft 02 has been controlled to a crankangle included within the previously-discussed all-cylinder intake-valve-closed crankangle area (e.g., α1). When the answer to step S4 is negative (NO), the routine returns from step S4 to step S2. Conversely when the answer to step S4 is affirmative (YES), step S5 occurs.

At step S5, an engine-stop signal is generated from ECU 22.

At step S6 subsequently to step S5, rotation of the engine is actually stopped.

The internal combustion engine is held in its engine stopped state, until a subsequent engine-restart action is initiated. In the engine stopped state, intake-valve VEL mechanism 1 is held at a default position where the operating mode of each of intake valves 4, 4 is stably kept at its minimum working angle D1 operating mode by the spring force of coil spring 30, and whereby the small intake-valve lift L1 and small working angle D1 characteristic can be maintained. In the engine stopped state, intake-valve VTC mechanism 2 is also held at a default position where valve timing (intake-valve open timing IVO and intake-valve closure timing IVC) of each of intake valves 4, 4 is stably kept at the maximum phase-retard position by the spring forces of return springs 55-56, and whereby the maximum phase-retard position of intake-valve VTC mechanism 2 can be maintained. On the other hand, crankshaft 02 is maintained at an angular position corresponding to a crankangle included within the previously-discussed all-cylinder intake-valve-closed crankangle area (e.g., α1), for example a specific crankangle corresponding to the point “B” indicated by the asterisk in FIGS. 9A-9D and included within the specific crankangle area α1.

At step S7, a check is made to determine whether an engine-restart condition is satisfied. For instance, in the presence of a requirement of re-acceleration on automatic-engine-stop-restart system equipped hybrid vehicles, ECU 22 determines that such an engine-restart condition is satisfied. When the answer to step S7 is negative (NO), one execution cycle terminates. Conversely when the answer to step S7 is affirmative (YES), step S8 occurs.

At step S8, the current engine/vehicle conditions, for example, the current value of engine temperature T, detected by the engine temperature sensor, are read. Thereafter, step S9 occurs.

At step S9, a check is made to determine whether the engine temperature T (read through step S8) is higher than a predetermined first temperature value T1. When the answer to step S9 is negative (NO), that is, when T≦T1, in other words, when ECU 22 determines that the engine is cold, the routine proceeds to step S10.

At step S10, the middle working angle D2 is set to an engine-start desired working angle Dt for intake-valve VEL mechanism 1. Thereafter, the routine proceeds to step S14. At the point of time, corresponding to step S10, valve timing of each of intake valves 4, 4 is stably kept at the maximum phase-retard position by means of intake-valve IVC mechanism 2, and also intake-valve closure timing IVC is controlled to a timing value near the BDC position. For instance, intake-valve closure timing IVC of the #1 cylinder exists near the BDC position of the #1 cylinder (in other words, near the TDC position of the #2 cylinder on compression stroke). Hence, it is possible to set the effective compression ratio to a high value, thus improving combustion during cold-engine operation.

Returning to step S9, when the answer to step S9 is affirmative (YES), that is, when T>T1, the routine proceeds from step S9 to step S11.

At step S11, a check is made to determine whether the engine temperature T is higher than or equal to a predetermined second temperature value T2. When the answer to step S11 is affirmative (YES), that is, when T≧T2, in other words, when ECU 22 determines that the engine temperature is high, the routine proceeds to step S12.

At step S12, the large working angle (the maximum working angle) D3 is set to engine-start desired working angle Dt for intake-valve VEL mechanism 1. At the point of time, corresponding to step S12, intake-valve VTC mechanism 2 is still stably kept at its default position (that is, at the maximum phase-retard position). Thus, intake-valve closure timing IVC of the #1 cylinder is greatly phase-retarded from the BDC position of the #1 cylinder (in other words, from the TDC position of the #2 cylinder on compression stroke). In a similar manner, as for the other cylinders, intake-valve closure timings IVC tend to be greatly phase-retarded from the respective BDC positions. Hence, it is possible to set the effective compression ratio to a low value, thus effectively suppressing the occurrence of preignition by virtue of such a decompressing action. Owing to high engine oil temperature (in other words, low-viscosity engine oil) there is a possibility of increased starting-period noise and vibrations, but by virtue of the decompressing action it is possible to suppress such starting-period noise and vibrations.

Conversely when the answer to step S11 is negative (NO), that is, when T1<T<T2, in other words, when ECU 22 determines that the engine condition is a general engine-start condition, the routine proceeds from step S11 to step S13.

At step S13, the small working angle (the minimum working angle) D1 is set to engine-start desired working angle Dt for intake-valve VEL mechanism 1.

After engine-start desired working angle Dt has been determined as previously (see steps S10, S12, S13 in FIG. 10), the routine proceeds to step S14.

At step S14, a control signal is outputted from ECU 22 to the control-shaft actuator (motor 20) of intake-valve VEL mechanism 1 in such a manner as to switch or convert the operating mode of each of intake valves to engine-start desired working angle Dt operating mode prior to cranking action of crankshaft 02.

Herein, a specific crankangle indicated by the above-mentioned point “B” in FIGS. 9A-9D and included within the specific crankangle area α1, corresponds to a crankangle at which intake valves 4, 4 of all engine cylinders simultaneously close, and therefore there is a less valve-spring reaction force acting on control cam 18 (or control shaft 17), thus ensuring a smooth rotary motion of control cam 18. By virtue of the smooth rotary motion of control cam 18, a smooth conversion to the desired intake-valve lift and event (working angle) characteristic can be initiated. Furthermore, the sliding-contact portion between control cam 18 and rocker arm 11 changes from a static friction area (of a large friction coefficient) to a kinetic friction area (of a small friction coefficient). Such a multiplied and combined effect of the smooth rotary motion of control cam 18 (the smooth conversion to the desired intake-valve lift and event (working angle) characteristic) and the friction-coefficient change of the sliding-contact portion from a large static friction coefficient to a small kinetic friction coefficient, ensures a more smooth lift-and-event (working angle) characteristic conversion of intake-valve VEL mechanism 1. Even during a conversion from the minimum lift L1 and working angle D1 characteristic to the middle lift L2 and working angle D2 characteristic during which intake valves 4, 4 of some cylinders (two cylinders in the shown embodiment) begin to open or even during a further conversion to the maximum lift L3 and working angle D3 characteristic, the previously-discussed smooth lift-and-event (working angle) characteristic conversion action of intake-valve VEL mechanism 1 can be continuously executed. Moreover, by virtue of rotational inertia of control shaft 17, which begins to already rotate, the smooth conversion action can be continuously executed.

At step S15, cranking action of crankshaft 02 is initiated by means of electric motor 07. The timing of initiation of cranking action may be set to a point of time when engine-start desired working angle Dt for intake-valve VEL mechanism 1 has been reached. Alternatively, the cranking action may be initiated during a time period from the start time of conversion to engine-start desired working angle Dt to the finish time of conversion to engine-start desired working angle Dt, in other words, under an unconfirmed state of completion of conversion to engine-start desired working angle Dt.

In the case of the former, that is, in the confirmed state where completion of conversion to engine-start desired working angle Dt has been confirmed, as a matter of course, at the early stage of cranking action the lift and event (working angle) characteristic has already been controlled or adjusted to the engine-start desired working angle Dt. Thus, it is possible to provide a desirable high startability (restartability) of the internal combustion engine. Additionally, at the point of time when cranking action is initiated by electric motor 07, it has been already past the time of the peak current of the control-shaft actuator (motor 20) of VEL mechanism 1. Thus, the adequate battery voltage can be supplied to electric motor 07, thereby enabling satisfactory cranking action with much electric energy.

Conversely in the case of the latter, that is, in the unconfirmed state of completion of conversion to engine-start desired working angle Dt, cranking action can be quickly initiated before completion of conversion to engine-start desired working angle Dt. That is, immediately when step S7 of FIG. 10 determines that the engine-restart condition is satisfied, cranking action can be quickly initiated, thus enabling a rapid transition to a burning state of the engine. This means a merit of a rapid acceleration rate of the vehicle, for instance, when accelerating from vehicle standstill. The flowchart illustrated in FIG. 10, is mainly based on the assumption that cranking action can be quickly initiated under the unconfirmed state of completion of conversion to engine-start desired working angle Dt. Hence, through step S15, cranking action is quickly initiated in the unconfirmed state of completion of conversion to engine-start desired working angle Dt.

When taking account of the initial compression of the early stage of cranking action, the initial compression in the #1 cylinder takes place at a specific crankangle indicated by the point “B” in FIGS. 9A-9D (that is, slightly before the TDC position on compression stroke). Therefore, atmospheric pressure has already been flown via the piston-to-cylinder clearance space into the engine cylinder after the engine has been stopped, and hence the air-fuel mixture is compressed under atmospheric pressure as an initial condition, and then the compression increases during a period from the specific crankangle indicated by the point “B” in FIGS. 9A-9D to the TDC position on compression stroke.

However, in the first place, the specific crankangle indicated by the point “B” is near the TDC position on compression stroke (exactly, slightly before the TDC position on compression stroke), and thus a piston stroke to the TDC position is very short. This means a slight work of compression, in other words, a smooth cranking speed increase. From this viewpoint, it is possible to more greatly enhance the engine startability (the engine restartability). Suppose that the specific crankangle indicated by the point “B” is just after the TDC position on compression stroke. The initial compression in the #1 cylinder itself does not take place. In such a case, it is possible to more smoothly increase the cranking speed.

Hence, as a matter of course, preignition and starting-period noise/vibrations can be effectively suppressed, and also speedy starting action can be realized.

At step S16 subsequently to step S15, a check is made to determine whether engine-start desired working angle Dt for intake-valve VEL mechanism 1 has been reached. When the answer to step S16 is affirmative (YES), the routine proceeds to step S17.

At step S17, complete explosion control (fuel injection and ignition) is executed, and thus certain and speedy starting action has been completed.

Returning to step S16, when the answer to step S16 is negative (NO), that is, when engine-start desired working angle Dt is not yet reached, the routine returns from step S16 to step S14, to generate a control signal corresponding to engine-start desired working angle Dt once again through step S14, and also to continually execute cranking action through step S15. Such a flow returning from step S16 back to step S14 is repeatedly executed until engine-start desired working angle Dt has been reached.

By the way, in the first embodiment, as an engine condition (or an engine/vehicle condition), engine temperature (concretely, engine oil temperature and/or engine coolant temperature) is used. In addition to the engine temperature, a vehicle speed may be added to the engine/vehicle condition. By virtue of the use of more informational data about the engine/vehicle condition, containing vehicle speed as well as engine temperature, it is possible to more accurately set or determine a desired working angle (i.e., engine-start desired working angle Dt).

Second Embodiment

Referring now to FIGS. 11A-11B and 12, there is shown the engine control device of the second embodiment wherein the inventive concept is applied to an in-line two-cylinder internal combustion engine. The basic configurations of intake-valve VEL mechanism 1 and intake-valve VTC mechanism 2, both constructing the variable valve actuation device incorporated in the engine control system of the second embodiment, are the same as those of the first embodiment.

As can be seen from the characteristic diagrams of FIGS. 11A-11B, in the engine stopped state, intake-valve VEL mechanism 1 is held at a default position where the operating mode of each of intake valves 4, 4 is kept at its minimum working angle D1′ operating mode.

As clearly seen in FIGS. 11A-11B, there is a specific crankangle area α1′, in which intake valves 4, 4 of the two engine cylinders (#1 and #2 cylinders) are kept simultaneously in their closed states, near the TDC position of the #1 cylinder on compression stroke. The specific crankangle area α1′ corresponds to an interval ranging from a crankangle at which intake valves 4, 4 of the #1 cylinder close (i.e., intake-valve closure timing IVC of the #1 cylinder) to a crankangle at which intake valves 4, 4 of the #2 cylinder open (i.e., intake-valve open timing IVO of the #2 cylinder). The length of the specific crankangle area α1′ (see FIGS. 11A-11B) of the second embodiment, wherein the inventive concept is applied to a two-cylinder engine, is sufficiently enlarged as compared to the length of the specific crankangle area α1 (see FIGS. 9A-9D) of the first embodiment, wherein the inventive concept is applied to a four-cylinder engine. This is because the interval between two engine cylinders in a four-cylinder engine is 180 degrees of crankangle, whereas the interval between two engine cylinders in a two-cylinder engine is 360 degrees of crankangle.

Therefore, in the second embodiment, the control area of a desired crankangle to be achieved by the crank position change mechanism can be enlarged to the specific crankangle area α1′. Even when the crank position change mechanism has a somewhat low control accuracy, the control system of the second embodiment can easily achieve the desired crankangle, because of the enlarged specific crankangle area α1′.

Additionally, in the second embodiment, the minimum working angle D1′ (that is, a valve open period from intake-valve open timing IVO to intake-valve closure timing IVC at the minimum working angle operating mode) is assumed to be shorter than the minimum working angle D1 of the first embodiment, that is, D1′<D1.

The minimum working angle D1 of the first embodiment, obtained by intake-valve VEL mechanism 1, means an effective working angle (an effective lifted period) from an actual valve-open point to an actual valve-closed point, fully taking account of the valve clearance Δ (see FIG. 5). On the other hand, as shown in FIG. 12, the minimum working angle D1′ of the second embodiment, obtained by intake-valve VEL mechanism 1, means a really effective lift section ranging from the point immediately after the leading edge of the positive valve opening acceleration (except a moderate valve-opening ramp section (i.e., a very small lift ΔL), which permits moderate valve movement in the first stage (the initial stage) of opening motion of intake valve 4) to the point immediately before the trailing edge of the positive valve closing acceleration (except a moderate valve-closing ramp section (i.e., a very small lift ΔL), which permits moderate valve movement in the last stage of closing motion of intake valve 4).

As can be appreciated from the relationship defined by the inequality of D1′<D1, the length of the specific crankangle area α1′ (see FIGS. 11A-11B) of the second embodiment can be further enlarged in comparison with the length of the specific crankangle area α1 (see FIGS. 9A-9D) of the first embodiment, by the difference (D1−D1′) between the minimum working angle D1 and the minimum working angle D1′. As a result of this, the level of the required control accuracy of the crank position change mechanism can be relaxed or reduced. In other words, owing to the enlarged control area of a desired crankshaft-rotation stopped position, the controllability of the crank position change mechanism can be greatly improved.

FIG. 12 shows the linkage attitude of the multinodular-link motion transmitting mechanism (the motion converter) of intake-valve VEL mechanism 1, during a ramp-lift period. As seen from the axial view of FIG. 12, during the ramp-lift period, the offset distance of the point of action of force (a load Fs) from the center X of oscillating motion of rockable cam 9 becomes a sufficiently small offset distance ΔT. Even when load Fs (the spring force of valve spring 5) acts on rockable cam 9, a moment acting on rockable cam 9 also becomes a sufficiently small moment ΔM, because of the sufficiently small offset distance ΔT. Hence, the magnitude of load transmitted via link rod 13 and acting on control cam 18 also becomes sufficiently small. This enables a smooth rotary motion of control shaft 17 even during the engine stopping period. That is, it is possible to realize a smooth operating mode shift to a desired intake-valve lift and event (working angle) characteristic under a substantially zero-lift state.

For instance, in the case of a small lift L1 exceeding the ramp lift period, as shown in FIG. 3B, the offset distance of the point of action of load Fs from the center X of oscillating motion of rockable cam 9 becomes a large offset distance T. A moment acting on rockable cam 9 also becomes a large moment M, because of the large offset distance T. Hence, the magnitude of load transmitted via link rod 13 and acting on control cam 18 becomes remarkably large, and thus it is difficult to smoothly rotate control shaft 17 under an engine stopped state.

Furthermore, suppose that a midpoint “B′” (indicated by the asterisk in FIGS. 11A-11B) of the control area of a desired crankangle to be achieved by the crank position change mechanism (that is, the specific crankangle area α1′) is set as a target point (a desired crankshaft-rotation stopped position of crankshaft 02), and then crankshaft rotation position control (simply, crank position control) is performed by electric motor 07 of the crank position change mechanism. In this case, regardless of each piston's individual operating characteristics, for example a difference in frictional resistances between reciprocating pistons 01, it is possible to accurately control or adjust the angular position of crankshaft 02 to a crankangle included within the desired-crankangle control area (i.e., the specific crankangle area α1′).

As will be appreciated from the above, the inventive concept of the invention is applied to a four-cycle four-cylinder internal combustion engine (the first embodiment) and also to a two-cylinder internal combustion engine (the second embodiment). 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. For instance, the number of engine cylinders is not limited to “4” or “2”. As appreciated, the all-cylinder intake-valve-closed crankangle area (e.g., the specific crankangle area α1 (or α1′) in which intake valves 4, 4 of all cylinders are kept in their valve-closed states) tends to be enlarged, as the number of engine cylinders reduces (for example, 6-cylinder engine→5-cylinder engine→4-cylinder engine→3-cylinder engine→2-cylinder engine). Thus, the fewer the number of engine cylinders, the easier the crank position control is achieved.

In the shown embodiments, the crank position change mechanism is constructed mainly by electric motor 07 and pinion gear mechanism 06. In lieu thereof, an electric motor may be directly connected to the rear end of crankshaft 02.

As a vehicle to which the engine control device (or the engine control system) of the embodiment is applied, an idling-stop system and multi-cylinder internal-combustion-engine equipped vehicle is exemplified. In this case, the vehicle also stops simultaneously with an automatic stop of the engine. The inventive concept can be applied to an automatic engine stop-restart system equipped hybrid vehicle, which can be propelled by means of a motor/generator or an electric motor (serving as a propelling power source) in an engine stopped state.

In the shown embodiments, the initial attitude of rockable cam 9 is changed via the multinodular-link motion transmitting mechanism (the motion converter) of intake-valve VEL mechanism 1, constructing part of the variable valve actuation device, by rotary motion of control shaft 17. Instead of using such a rotary motion of control shaft 17, the motion converter may be configured such that the initial attitude of rockable cam 9 is changed by displacing an axial position of control shaft 17.

OPERATIONS AND EFFECTS

(1) According to the internal combustion engine control device of the shown embodiments, when the engine stops, the variable valve actuation device is controlled to change an operating mode of each of the intake valves to a specific state where an all-cylinder valve closed period, during which the intake valves of the cylinders are all kept in their non-lifted states, occurs, and the crank position change mechanism is controlled to change the crankshaft-rotation stopped position to a crankangle included within the all-cylinder valve closed period. When restarting the engine, the variable valve actuation device is controlled to bring the operating mode of each of the intake valves closer to an engine-start desired lift characteristic Dt suited to an engine condition (e.g., engine temperature T), prior to cranking action. This ensures a good engine restartability.

(2) Concretely, in the shown embodiments, when the engine stops, the variable valve actuation device (in particular, intake-valve VEL mechanism 1) is controlled to change the working angle of the intake valve to a small working angle side within a controllable working angle range.

(3) Also, the variable valve actuation device (in particular, intake-valve VEL mechanism 1) is configured so that a lower limit of a controllable working angle range of the intake valve can be controlled to a minimum working angle D1 that does not reach a zero working angle.

Thus, conversion into a desired working angle (i.e., engine-start desired working angle Dt) starts from an initial working angle (e.g., the minimum working angle D1), which does not yet reach a zero working angle. Owing to such a narrow width of conversion, it is possible to shorten a response time of conversion into the desired working angle, and also to enables a more simplified configuration of the variable valve actuation device.

(4) As may be appreciated from the flowchart of FIG. 10 and the characteristic diagrams of FIGS. 9A-9D, when engine temperature T, detected when restarting the engine, is lower than or equal to a predetermined first temperature value T1, valve closure timing IVC of the intake valve is controlled to a timing value near a piston bottom dead center position on intake stroke. Thus, in the case that engine temperature T, detected when restarting the engine, is very low (T≦T1), it is possible to improve combustion.

(5) Additionally, when the engine temperature T, detected when restarting the engine, is higher than or equal to a predetermined second temperature value T2 exceeding the first temperature value T1, the valve closure timing IVC of the intake valve is controlled to a timing value phase-retarded from the piston bottom dead center position on intake stroke. Thus, in the case that engine temperature T, detected when restarting the engine, is high (T2≦T), it is possible to suppress preignition and starting-period noise/vibrations.

(6) Furthermore, an intake-valve working angle (e.g., D2), suited to the restarting-period engine temperature lower than or equal to the first temperature value T1, is controlled to be greater than an intake-valve working angle (e.g., D1), suited to the restarting-period engine temperature higher than the first temperature value T1 and lower than the second temperature value T2, and also controlled to be less than an intake-valve working angle (e.g., D3), suited to the restarting-period engine temperature higher than the second temperature value T2.

(7) When the engine stops, the crankshaft-rotation stopped position is changed to a crankangle included within the all-cylinder valve closed period by the crank position change mechanism, after the variable valve actuation device has been controlled to change the operating mode of each of the intake valves to the specific state where the all-cylinder valve closed period, during which the intake valves of the cylinders are all kept in their non-lifted states, occurs.

(8) The internal combustion engine may be mounted on an automotive vehicle having an idling-stop function.

(9) The crank position change mechanism may be configured to change the crankshaft-rotation stopped position by controlling an electric motor used as a propelling power source of the vehicle.

(10) The internal combustion engine may be mounted on a hybrid vehicle that can be propelled by only an electric motor, under a stopped state of the engine.

(11) The crank position change mechanism may be configured to change the crankshaft-rotation stopped position by controlling an alternator that converts mechanical energy into electrical energy for charging a battery.

(12) When restarting the engine, the variable valve actuation device may be controlled to bring the operating mode of each of the intake valves closer to the engine-start desired lift characteristic (i.e., engine-start desired working angle Dt) suited to the engine condition (e.g., engine temperature T), immediately after an electric power source has been switched ON with an ignition switch turned ON.

(13) When initiating cranking action forcibly by means of the ignition switch before controlling the variable valve actuation device to bring the operating mode of each of the intake valves closer to the engine-start desired lift characteristic (i.e., engine-start desired working angle Dt) suited to the engine condition (e.g., engine temperature T), the variable valve actuation device may be controlled to bring the engine-start desired lift characteristic of the intake valve during the forcible cranking action.

(14) Alternatively, The cranking action may be initiated, after the engine-start desired lift characteristic (i.e., engine-start desired working angle Dt) of the intake valve has been reached via the variable valve actuation device.

(15) The control shaft 17 may be configured to be driven directly by a driving force produced by an electric motor.

(16) A specific crankangle area, including a specific crankangle (e.g., point “B”) corresponding to the crankshaft-rotation stopped position to be reached when the engine stops, may be set to a given crankangle area (e.g., α1) including a piston top dead center position on compression stroke with respect to a given engine cylinder (e.g., #1 cylinder).

(17) In lieu thereof, a specific crankangle area, including a specific crankangle corresponding to the crankshaft-rotation stopped position to be reached when the engine stops, may be set to a given crankangle area (e.g., α1) including a crankangle just before a piston top dead center position on compression stroke with respect to a given engine cylinder (e.g., #1 cylinder).

In the case of the engine control device (the engine control system) as described in the above items (16)-(17), during a period from the time when the engine stops to the time when the engine is restarted, an in-cylinder pressure in the given engine cylinder (e.g., #1 cylinder) in close proximity to or just before a piston top dead center position on compression stroke, tends to lower to atmospheric pressure, but a decompressing effect can be realized owing to a less piston stroke or a very short piston stroke during a first compressing action. Therefore, it is possible to more remarkably suppress or reduce preignition and starting-period noise and vibrations.

The entire contents of Japanese Patent Application No. 2009-238810 (filed Oct. 16, 2009) 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.

Claims

1. An internal combustion engine control device comprising:

a crank position change mechanism configured to change a crankshaft-rotation stopped position of a crankshaft of an internal combustion engine; and

a variable valve actuation device configured to change at least a working angle of each of intake valves of a plurality of engine cylinders by changing a position of a control shaft,

wherein, when the engine stops, the variable valve actuation device is controlled to change an operating mode of each of the intake valves to a specific state where an all-cylinder valve closed period, during which the intake valves of the cylinders are all kept in their non-lifted states, occurs, and the crank position change mechanism is controlled to change the crankshaft-rotation stopped position to a crankangle included within the all-cylinder valve closed period, and

wherein, when restarting the engine, the variable valve actuation device is controlled to bring the operating mode of each of the intake valves closer to an engine-start desired lift characteristic suited to an engine condition, prior to cranking action.

2. The internal combustion engine control device as claimed in claim 1, wherein:

when the engine stops, the variable valve actuation device is controlled to change the working angle of the intake valve to a small working angle side within a controllable working angle range.

3. The internal combustion engine control device as claimed in claim 1, wherein:

when an engine temperature, detected when restarting the engine, is lower than or equal to a predetermined first temperature value, valve closure timing of the intake valve is controlled to a timing value near a piston bottom dead center position on intake stroke.

4. The internal combustion engine control device as claimed in claim 3, wherein:

when the engine temperature, detected when restarting the engine, is higher than or equal to a predetermined second temperature value exceeding the first temperature value, the valve closure timing of the intake valve is controlled to a timing value phase-retarded from the piston bottom dead center position on intake stroke.

5. The internal combustion engine control device as claimed in claim 4, wherein:

an intake-valve working angle, suited to the restarting-period engine temperature lower than or equal to the first temperature value, is controlled to be greater than an intake-valve working angle, suited to the restarting-period engine temperature higher than the first temperature value and lower than the second temperature value, and also controlled to be less than an intake-valve working angle, suited to the restarting-period engine temperature higher than the second temperature value.

6. The internal combustion engine control device as claimed in claim 1, wherein:

when the engine stops, the crankshaft-rotation stopped position is changed to a crankangle included within the all-cylinder valve closed period by the crank position change mechanism, after the variable valve actuation device has been controlled to change the operating mode of each of the intake valves to the specific state where the all-cylinder valve closed period, during which the intake valves of the cylinders are all kept in their non-lifted states, occurs.

7. The internal combustion engine control device as claimed in claim 1, wherein:

the internal combustion engine is mounted on an automotive vehicle having an idling-stop function.

8. The internal combustion engine control device as claimed in claim 7, wherein:

the crank position change mechanism is configured to change the crankshaft-rotation stopped position by controlling an electric motor used as a propelling power source of the vehicle.

9. The internal combustion engine control device as claimed in claim 1, wherein:

the crank position change mechanism is configured to change the crankshaft-rotation stopped position by controlling an electric motor.

10. The internal combustion engine control device as claimed in claim 1, wherein:

the internal combustion engine is mounted on a hybrid vehicle that can be propelled by only an electric motor, under a stopped state of the engine.

11. The internal combustion engine control device as claimed in claim 1, wherein:

the crank position change mechanism is configured to change the crankshaft-rotation stopped position by controlling an alternator that converts mechanical energy into electrical energy for charging a battery.

12. The internal combustion engine control device as claimed in claim 1, wherein:

when restarting the engine, the variable valve actuation device is controlled to bring the operating mode of each of the intake valves closer to the engine-start desired lift characteristic suited to the engine condition, immediately after an electric power source has been switched ON with an ignition switch turned ON.

13. The internal combustion engine control device as claimed in claim 12, wherein:

when initiating cranking action forcibly by means of the ignition switch before controlling the variable valve actuation device to bring the operating mode of each of the intake valves closer to the engine-start desired lift characteristic suited to the engine condition, the variable valve actuation device is controlled to bring the engine-start desired lift characteristic of the intake valve during the forcible cranking action.

14. The internal combustion engine control device as claimed in claim 12, wherein:

the cranking action is initiated, after the engine-start desired lift characteristic of the intake valve has been reached via the variable valve actuation device.

15. The internal combustion engine control device as claimed in claim 14, wherein:

the control shaft is driven directly by a driving force produced by an electric motor.

16. The internal combustion engine control device as claimed in claim 1, wherein:

a specific crankangle area, including a specific crankangle corresponding to the crankshaft-rotation stopped position to be reached when the engine stops, is set to a given crankangle area including a piston top dead center position on compression stroke with respect to a given engine cylinder.

17. The internal combustion engine control device as claimed in claim 1, wherein:

a specific crankangle area, including a specific crankangle corresponding to the crankshaft-rotation stopped position to be reached when the engine stops, is set to a given crankangle area including a crankangle just before a piston top dead center position on compression stroke with respect to a given engine cylinder.

18. An internal combustion engine control device comprising:

a crank position change mechanism configured to change a crankshaft-rotation stopped position of a crankshaft of an internal combustion engine; and

a variable valve actuation device configured to change at least a working angle of each of intake valves of a plurality of engine cylinders by changing a position of a control shaft,

wherein, when the engine stops, the crank position change mechanism, together with the variable valve actuation device, is controlled to execute crank position control as well as intake-valve operating characteristic control in such a manner as to realize a specific state where there is a less valve-spring reaction force acting on the control shaft, and

wherein, when restarting the engine, the variable valve actuation device is controlled to bring the position of the control shaft closer to a desired position suited to start-up of the engine, prior to cranking action.

19. The internal combustion engine control device as claimed in claim 18, wherein:

the variable valve actuation device is configured so that a lower limit of a controllable working angle range of the intake valve can be controlled to a minimum working angle that does not reach a zero working angle.

20. An internal combustion engine control system comprising:

a crank position change mechanism configured to change a crankshaft-rotation stopped position of a crankshaft of an internal combustion engine;

a variable valve actuation device configured to change a valve lift as well as a working angle of each of intake valves of a plurality of engine cylinders; and

a controller configured to control operations of the crank position change mechanism and the variable valve actuation device,

wherein, when the engine stops, the controller controls the variable valve actuation device to change an operating characteristic of each of the intake valves to a specific state where an all-cylinder valve closed period, during which the intake valves of the cylinders are all kept in their non-lifted states, occurs, and the controller controls the crank position change mechanism to change the crankshaft-rotation stopped position to a crankangle included within the all-cylinder valve closed period, and

wherein, when restarting the engine, the controller controls the variable valve actuation device to bring an engine-start desired lift characteristic of each of the intake valves, suited to an engine condition, prior to cranking action.