US20090159027A1

US20090159027A1 - Variable valve actuating apparatus for internal combustion engine, and controller for variable valve actuating apparatus

Info

Publication number: US20090159027A1
Application number: US12/341,668
Authority: US
Inventors: Makoto Nakamura
Original assignee: Hitachi Ltd
Current assignee: Hitachi Ltd
Priority date: 2007-12-25
Filing date: 2008-12-22
Publication date: 2009-06-25
Also published as: JP2009156029A

Abstract

A variable valve actuating apparatus for an internal combustion engine includes an exhaust valve operating angle varying mechanism, and an exhaust valve timing varying mechanism. When the internal combustion engine is at rest, the exhaust: valve operating angle varying mechanism sets an exhaust valve operating angle relatively small, and the exhaust valve timing varying mechanism sets an exhaust valve opening timing and an exhaust valve closing timing relatively retarded. Accordingly, when the internal combustion engine is at rest, the variable valve actuating apparatus sets the exhaust valve closing timing at or close to top dead center so as to enable startup of the internal combustion engine.

Description

BACKGROUND OF THE INVENTION

The present invention relates to variable valve actuating apparatuses or systems for controlling opening and closing timings of intake valves and/or exhaust valves of internal combustion engines, and controllers for variable valve actuating systems.
U.S. Pat. No. 6,502,535 corresponding to Japanese Patent Application Publication No. 2001-355469 discloses a variable valve actuating system for an internal combustion engine. The variable valve actuating system includes a first valve actuating mechanism for varying an exhaust valve operating angle of the internal combustion engine, and a second valve actuating mechanism for advancing or retarding an exhaust valve lift phase of the internal combustion engine, which are controlled according to engine operating state for improving engine performance. At cold start of the internal combustion engine, the first valve actuating mechanism sets the exhaust valve operating angle minimized, and the second valve actuating mechanism sets the exhaust valve lift phase most advanced, so as to advance the exhaust valve closing timing, for increasing the quantity of residual burned gas in a cylinder, and thereby quickly warming up the internal combustion engine.

SUMMARY OF THE INVENTION

According to U.S. Pat. No. 6,502,535, at cold start of the internal combustion engine, the variable valve actuating system may set the exhaust valve closing timing excessively advanced. This may increase the quantity of residual burned gas in the cylinder excessively high, and thereby adversely affect combustion in the cylinder.
It is desirable to provide a variable valve actuating apparatus or system for an internal combustion engine which is capable of preventing the problems described above, and to provide a controller for a variable valve actuating system which is capable of preventing the problems described above.
According to one aspect of the present invention, a variable valve actuating apparatus for an internal combustion engine, comprises: a first valve actuating mechanism arranged to vary at least an exhaust valve operating angle of the internal combustion engine, wherein when the internal combustion engine is at rest, the first valve actuating mechanism sets the exhaust valve operating angle relatively small; and a second valve actuating mechanism arranged to vary an exhaust valve opening timing and an exhaust valve closing timing of the internal combustion engine, wherein when the internal combustion engine is at rest, the second valve actuating mechanism sets the exhaust valve opening timing and the exhaust valve closing timing relatively retarded, wherein when the internal combustion engine is at rest, the first and second valve actuating mechanisms set the exhaust valve closing timing at or close to top dead center so as to enable startup of the internal combustion engine.
According to another aspect of the present invention, a variable valve actuating apparatus for an internal combustion engine, comprises: a first valve actuating mechanism arranged to vary at least an exhaust valve operating angle of the internal combustion engine, wherein when de-energized, the first valve actuating mechanism mechanically sets the exhaust valve operating angle at or close to a minimum setpoint; and a second valve actuating mechanism arranged to advance or retard an exhaust valve opening timing and an exhaust valve closing timing of the internal combustion engine, wherein when de-energized, the second valve actuating mechanism mechanically sets the exhaust valve opening timing and the exhaust valve closing timing at or close to a most retarded setpoint.
According to a further aspect of the present invention, a controller for a variable valve actuating apparatus for an internal combustion engine, the variable valve actuating apparatus comprising a first valve actuating mechanism arranged to vary at least an exhaust valve operating angle of the internal combustion engine, and a second valve actuating mechanism arranged to vary an exhaust valve opening timing and an exhaust valve closing timing of the internal combustion engine, the controller comprises: a section connected for signal communication therewith to the first and second valve actuating mechanisms, wherein at startup of the internal combustion engine, the section sets by the first valve actuating mechanism the exhaust valve operating angle relatively small, and sets by the second valve actuating mechanism the exhaust valve opening timing and the exhaust valve closing timing relatively retarded, so as to set the exhaust valve closing timing at or close to top dead center for enabling startup of the internal combustion engine.
According to a still further aspect of the present invention, a variable valve actuating apparatus for an internal combustion engine, comprises: a first valve actuating mechanism arranged to vary at least an exhaust valve operating angle of the internal combustion engine, wherein when the internal combustion engine is at rest, the first valve actuating mechanism is automatically held in a default operating position to set the exhaust valve operating angle smaller by a predetermined amount than a maximum setpoint; a second valve actuating mechanism arranged to vary an exhaust valve lift phase of the internal combustion engine, wherein when the internal combustion engine is at rest, the second valve actuating mechanism is automatically held in a default operating position to set the exhaust valve lift phase more retarded by a predetermined amount than a most advanced setpoint; and a controller connected for signal communication therewith to the first and second valve actuating mechanisms, wherein at an initial stage of startup of the internal combustion engine, the controller outputs a control signal so as to maintain the first valve actuating mechanism in or close to the default operating position, and maintain the second valve actuating mechanism in or close to the default operating position.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic diagram showing an internal combustion engine system including a variable valve actuating system or apparatus according to first and second embodiments of the present invention.

FIG. 2 is a schematic diagram showing a perspective view of an exhaust valve operating angle varying mechanism, an exhaust valve timing varying mechanism, and an intake valve timing varying mechanism in the variable valve actuating system according to the first embodiment.

FIGS. 3A and 3B are diagrams illustrating how the exhaust valve operating angle varying mechanism operates when controlled to be in a state of small valve lift.

FIGS. 4A and 4B are diagrams illustrating how the exhaust valve operating angle varying mechanism operates when controlled to be in a state of maximum valve lift.

FIG. 5 is a graphic diagram showing how the lift, operating angle, and maximum lift phase of an exhaust valve of the engine are controlled by the variable valve actuating system.

FIG. 6 is a sectional view of the exhaust valve timing varying mechanism.

FIG. 7 is a sectional view, taken along a line VII-VII shown in FIG. 6, of the exhaust valve timing varying mechanism under a condition that the exhaust valve timing varying mechanism is controlled to be in a most retarded state.

FIG. 8 is a sectional view, taken along the line VII-VII shown in FIG. 6, of the exhaust valve timing varying mechanism under a condition that the exhaust valve timing varying mechanism is controlled to be in a most advanced state.

FIG. 9 is a front view of the intake valve timing varying mechanism under a condition that the intake valve timing varying mechanism is controlled to be in a most advanced state and a front cover is removed from the intake valve timing varying mechanism.

FIG. 10 is a front view of the intake valve timing varying mechanism under a condition that the intake valve timing varying mechanism is controlled to be in a most retarded state and the front cover is removed from the intake valve timing varying mechanism.

FIG. 11 is a schematic diagram showing characteristics of operation of intake valves and exhaust valves of the engine at an initial stage of engine start under control of the variable valve actuating system.

FIG. 12 is a schematic diagram showing characteristics of operation of the intake valves and exhaust valves at low or middle engine load under control of the variable valve actuating system.

FIG. 13 is a schematic diagram showing characteristics of operation of the intake valves and exhaust valves at high engine load under control of the variable valve actuating system.

FIG. 14 is a flow chart showing a control process which is performed by a controller of the variable valve actuating system.

FIG. 15 is a schematic diagram showing a perspective view of an exhaust valve operating angle varying mechanism, an exhaust valve timing varying mechanism, and an intake valve timing varying mechanism in the variable valve actuating system according to the second embodiment.

DETAILED DESCRIPTION OF THE INVENTION

FIG. 1 schematically shows an internal combustion engine system including a variable valve actuating system or apparatus according to first and second embodiments of the present invention. In these embodiments, the internal combustion engine system includes a four-cycle spark-ignition gasoline internal combustion engine. As shown in FIG. 1, the engine includes a cylinder block 01, a cylinder head 02, a piston 03, and an ignition plug 05, where a combustion chamber 04 is defined between cylinder head 02 and piston 03 in a cylinder bore formed in cylinder block 01. Ignition plug 05 is mounted in cylinder head 02, and located substantially at the center of combustion chamber 04. Cylinder block 01 is provided with a coolant temperature sensor 06 for measuring the temperature of engine coolant flowing in a water jacket. Cylinder head 02 is provided with a fuel injection valve 07 for injecting fuel directly into combustion chamber 04.
In cylinder head 02, an intake port 08 and an exhaust port 09 are formed. Cylinder head 02 is provided with intake valves 4, 4 and exhaust valves 5, 5 which are slidably mounted for opening and closing respective ones of intake port 08 and exhaust port 09.
As shown in FIGS. 1 and 2, the variable valve actuating system includes an exhaust valve operating angle varying mechanism (exhaust valve lift varying mechanism, exhaust valve event and lift varying mechanism, or exhaust VEL) 1 as a first valve actuating mechanism for continuously varying (increasing or reducing) the lift and operating angle (operating period, or period when a valve is open) of exhaust valves 5, and an exhaust valve timing varying mechanism (exhaust valve phase varying mechanism, exhaust valve timing control mechanism, or exhaust VTC) 2 as a second valve actuating mechanism for continuously varying (advancing or retarding) a phase (lift is phase, or maximum lift phase) of exhaust valves 5 so as to vary (advance or retard) the opening and closing timings of exhaust valves 5 (an exhaust valve opening timing EVO and an exhaust valve closing timing EVC), while holding constant the operating angle of exhaust valves 5. The variable valve actuating system further includes an intake valve timing varying mechanism (intake valve phase varying mechanism, intake valve timing control mechanism, or intake VTC) 3 as a third valve actuating mechanism for continuously varying (advancing or retarding) a phase (lift phase, or maximum lift phase) of intake valves 4 so as to vary (advance or retard) the opening and closing timings of intake valves 4 (an intake valve opening timing IVO and an intake valve closing timing IVC), while holding constant the operating angle of intake valves 4. Operation of the exhaust VEL 1, exhaust VTC 2, and intake VTC 3 is controlled by a controller 22 according to engine operating state, as described in detail below.
The exhaust VEL 1 has a construction similar to that of a corresponding device adapted to intake valves in Japanese Patent Application Publication No. 2003-172112. As shown in FIGS. 2 and 3A, the exhaust VEL 1 includes a hollow drive shaft 6 which is rotatably supported by bearings on an upper part of cylinder head 02; a drive cam 7 which is an eccentric rotary cam fixedly mounted on drive shaft 6 by press fitting in this example; a pair of swing cams 9 which are swingably mounted on drive shaft 6, and arranged to open the exhaust valves 5, respectively, by sliding on top surfaces of valve lifters 8 provided in the upper ends of exhaust valves 5; and a linkage or motion transmitting mechanism arranged to transmit rotation of drive cam 7 to swing cams 9 for swing motion.
Drive shaft 6 is arranged to receive rotation from a crankshaft through a rotation transmitting mechanism which, in this example, is a chain drive including a timing sprocket 33 provided on one end of drive shaft 6, a driving sprocket provided on the crankshaft, and a timing chain not shown. When driven by the crankshaft, the drive shaft 6 rotates in the clockwise direction as shown by an arrow in FIG. 2.
Drive cam 7 is shaped like a ring, and formed with a drive shaft receiving hole extending in the axial direction of drive cam 7. Drive cam 7 is fixedly mounted on drive shaft 6 extending through the drive shaft receiving hole. The axis of drive cam 7 is offset in the radial direction from the axis of drive shaft 6 by a predetermined distance.
As shown in FIGS. 2 and 3A, swing cams 9 are formed integrally at both ends of an annular camshaft 10. Camshaft 10 is hollow and rotatably mounted on drive shaft 6. Each swing cam 9 has a lower surface including a cam surface 9 a. Cam surface 9 a includes a base circle surface region on the cam shaft's side, a ramp surface region extending like a circular arc from the base circle surface region toward a cam nose, and a lift surface region extending from the ramp surface region toward an apex of the cam nose. The cam surface 9 a abuts on the top surface of the corresponding valve lifter 8 at a predetermined position, and the contact point of the cam surface 9 a shifts among the base circle surface region, ramp surface region and lift surface region in dependence on the swing position of the swing cam 9.
The above-mentioned linkage or motion transmitting mechanism includes a rocker arm 11 disposed above drive shaft 6; a link arm 12 connecting a first end portion 11 a of rocker arm 11 with drive cam 7; and a link rod 13 connecting a second end portion 11 b of rocker arm 11 with one swing cam 9.
Rocker arm 11 includes a tubular central base portion formed with a support hole, and rotatably mounted on a control cam 18 through the support hole. The first end portion 11 a of rocker arm 11 is connected rotatably with link arm 12 by a pin 14, and the second end portion 11 b is connected rotatably with a first end portion 13 a of link rod 13 by a pin 15.
Link arm 12 includes a relatively large annular base portion 12 a and a projection 12 b projecting outward from the base portion 12 a. Base portion 12 a is formed with a center hole in which the cam portion of the drive cam 7 is rotatably fit. The projection 12 b is connected rotatably with the first end portion 11 a of rocker arm 11 by pin 14.
Link rod 13 includes a second end 13 b which is connected rotatably with the cam nose of swing cam 9 by a pin 16.
Control shaft 17 extends in parallel to drive shaft 6 in the longitudinal direction of the engine, and is rotatable supported by the same bearings at a position just above drive shaft 6. Control cam 18 is fixedly mounted on control shaft 17 and fit slidably in the support hole of rocker arm 11 to serve as a fulcrum for the swing motion of rocker arm 11. Control cam 18 is shaped like a hollow cylinder, and the axis of control cam 18 is offset from the axis of the control shaft 17 by a predetermined distance. Rotation of control shaft 17 is controlled by a drive mechanism 19.
Drive mechanism 19 includes an electric motor 20 which is fixed to one end of a housing; and a transmission mechanism 21 to transmit rotation of the electric motor 20 to the control shaft 17. In this example, the transmission mechanism 21 is a ball screw transmission mechanism.
Electric motor 20 of this example is a proportional type DC motor. Electric motor 20 is controlled by a controller 22 in accordance with a measured operating state of the engine.
Ball screw transmission mechanism 21 includes a ball screw shaft 23, a ball nut 24, a connection arm 25 and a link member 26. Ball screw shaft 23 and the drive shaft of electric motor 20 are arranged end to end and aligned with each other so that their axes form a substantially straight line. Ball nut 24 serves as a movable nut screwed on the ball screw shaft 23 and arranged to move axially in accordance with the rotation. Connection arm 25 is connected with one end portion of control shaft 17. Link member 26 links the connection arm 25 and ball nut 24.
Ball screw shaft 23 is formed with an external single continuous ball circulating groove extending, in the form of a helical thread, over the outside surface of ball screw shaft 23. Ball screw shaft 23 and the drive shaft of electric motor 20 are connected end to end by a coupling member which transmits a rotational driving force from electric motor 20 to ball screw shaft 23. Ball nut 24 is approximately in the form of a hollow cylinder. Ball nut 24 is formed with an internal guide groove designed to hold a plurality of balls in cooperation with the ball circulating groove of ball screw shaft 23 so that the balls can roll between the guide groove and the circulating groove. This guide groove is a single continuous helical thread formed in the inside circumferential surface of ball nut 24. Ball nut 24 is arranged to translate the rotation of ball screw shaft 23 into a linear motion of ball nut 24 and produce an axial force.
A coil spring 30 as a biasing device is disposed around ball screw shaft 23 between ball nut 24 and a spring seat provided at the tip of ball screw shaft 23, so as to urge the ball nut 24 axially toward electric motor 20. It is to be understood from the following description that coil spring 30 serves to bias the ball nut 24 in the direction to reduce the lift and operating angle of exhaust valves 5. Accordingly, when the engine is stopped or at rest, then ball nut 24 is moved along ball screw shaft 23 toward a position for a minimum lift and minimum operating angle of exhaust valves 5 by the elastic force of the coil spring 30.
Controller 22 of this example is a common control unit or control section used for controlling all of the exhaust VEL 1, the exhaust VTC 2, and the intake VTC 3. Controller 22 is built in an engine control unit (ECU). Controller 22 is connected with various sensors to collect information on an operating state of the engine. Controller 22 receives data signals outputted from the sensors, and identifies the engine operating state on the basis of the data signals. The sensors include a crank angle sensor for sensing the rotation angle of the crankshaft and sensing an engine speed N (rpm), an accelerator opening sensor, a vehicle speed sensor, a gear position sensor, the coolant temperature sensor 06, a drive shaft angle sensor 28 for sensing the rotation angle of drive shaft 6, and a potentiometer (control shaft angle sensor) 29 for sensing the rotation angle of control shaft 17. Controller 22 measures the relative rotational position between timing sprocket 33 and drive shaft 6, and the lift and operating angle of exhaust valves 5, on the basis of the data signals from crank angle sensor 27, drive shaft angle sensor 28, and potentiometer 29.
The thus-constructed exhaust VEL 1 is controlled to operate as follows. When the engine is operating in a predetermined engine operating region, the controller 22 acts to move the ball nut 24 rectilinearly toward electric motor 20, by sending a control current to electric motor 20 and rotating the ball screw shaft 23 with electric motor 20. The movement of ball nut 24 is assisted by the elastic force of coil spring 30. With this movement of ball nut 24, the control shaft 17 is rotated in one direction by the link member 26 and connection arm 25. Accordingly, control cam 18 rotates about the axis of control shaft 17 so that the axis of control cam 18 rotates about the axis of control shaft 17, as shown in FIGS. 3A and 3B (in the form of rear view), and a thick wall portion of control cam 18 is shifted upwards from drive shaft 6. As a result, the pivot point between the second end portion 11 b of rocker arm 11 and link rod 13 is shifted upwards relative to the drive shaft 6. Therefore, each swing cam 9 is rotated in the counterclockwise direction as viewed in FIGS. 3A and 3B, and the cam nose is pulled upwards by link rod 13. Accordingly, drive cam 7 rotates and pushes up the first end portion 11 a of rocker arm 11 through link arm 12. Though a movement for valve lift is transmitted through link rod 13 to swing cam 9 and valve lifter 8, the valve lift is decreased sufficiently to a small lift L1 shown by a valve lift curve in FIG. 5, and the operating angle (valve opening period) D is decreased to a small value D1.
There is a valve clearance between swing cam 9 and valve lifter 8. Therefore, the actual valve lift is smaller by the valve clearance than the lift of swing cam 9. Accordingly, the valve operating angle is defined as extending from the timing when the valve is actually opened to the timing when the valve is actually closed.
When the engine is operating in another predetermined engine operating region, the controller 22 drives electric motor 20 in a reverse rotational direction, and thereby rotates the ball screw shaft 23 in the reverse direction. With this reverse rotation of ball screw shaft 23, the ball nut 24 moves in the axial direction away from electric motor 20 against the elastic force of coil spring 30, and control shaft 17 is rotated in the counterclockwise direction as viewed in FIGS. 3A and 3B by a predetermined amount. Therefore, the control cam 18 is held at the angular position at which the axis of control cam 18 is shifted downward by a predetermined amount from the axis of control shaft 17, and the thick wall portion of control cam 18 is shifted downwards. Rocker arm 11 is moved in the clockwise direction from the position of FIGS. 3A and 3B, and the end of rocker arm 11 pushes down the cam nose of swing cam 9 through link member 13, and swing cam 9 rotates in the clockwise direction slightly. Accordingly, drive cam 7 rotates and pushes up the end 11 a of rocker arm 11 through link arm 12. A movement for valve lift is transmitted through link member 13 to swing cams 9 and valve lifters 8. In this case, the valve lift is increased to a medium lift L2, and the operating angle is increased to a medium angle D2. By this control operation, the variable valve actuating system can shift the exhaust valve closing timing EVC on the retard side toward bottom dead center. By so doing, the variable valve actuating system can reduce the quantity of residual burned gas in combustion chamber 04, and thereby improve the startability of the engine at cold start.
When the engine operating point enters a high speed and high load region, this variable valve actuating system can rotate electric motor 20 in the reverse direction by sending the control signal from controller 22, to rotate control cam 18 further in the counterclockwise direction with control shaft 17 to the position at which the axis is rotated downwards as shown in FIGS. 4A and 4B.
Therefore, rocker arm 11 moves to a position closer to the drive shaft 6, and the second end 11 b pushes down the cam nose of swing cam 9 through link rod 13, so that the swing cam 9 is further rotated in the clockwise direction by a predetermined amount. Accordingly, drive cam 7 rotates and pushes up the first end 11 a of rocker arm 11 through link arm 12. A movement for valve lift is transmitted through link rod 13 to swing cam 9 and valve lifter 8. In this case, the valve lift is increased continuously from L2 to L3 as shown in FIG. 5. In this way, this system can improve the exhaust efficiency and the engine output in the high speed region.
In this way, the exhaust VEL 1 varies the lift of exhaust valves 5 continuously from the small lift L1 to the large lift L3, and also, the operating angle of exhaust valves 5 continuously from the small angle (angular distance) D1 to the large angle D3.
When the engine is at rest, the ball nut 24 is moved toward electric motor 20 and held in the position for small lift L1 and small operating angle D1 by the elastic force of coil spring 30. This reduces the level of friction of moving parts such as exhaust valves 5, and thereby enhances the startability of the engine, at start of the engine.
As shown in FIGS. 6, 7 and 8, the exhaust VTC 2 of this example is a vane type mechanism including the timing sprocket 33 for transmitting rotation to drive shaft 6; a vane member 32 as a movable member which is fixed to one end of drive shaft 6 and received rotatably in the timing sprocket 33; and a hydraulic circuit to rotate the vane member 32 in the forward and reverse directions by the use of an oil pressure.
Timing sprocket 33 includes a housing 34 receiving the vane member 32 rotatably; a front cover 35 shaped like a circular disk and arranged to close a front opening of housing 34; and a rear cover 36 shaped approximately like a circular disk and arranged to close a rear opening of housing 34. Housing 34 is sandwiched between front and rear covers 35 and 36, and joined with these covers to form a unit, by four small diameter bolts 37 extending in the axial direction of drive shaft 6. Housing 34 thus rotates in synchronization with the crankshaft.
Housing 34 is in the form of a hollow cylinder having the front and rear openings. Housing 34 includes a plurality of shoes 34 a projecting radially inwards from the inside circumferential surface and serving as a partition. In this example, four of the shoes 34 a are arranged at intervals of about 90 degrees.
Each shoe 34 a has an approximately trapezoidal cross section. A bolt hole 34 b is formed approximately at the center of each shoe 34 a. Each bolt hole 34 b passes axially through one of shoes 34 a, and receives the shank of one of the axially extending bolts 37. Each shoe 34 a includes an inner end surface. A retaining groove extends axially in the form of cutout in the inner end surface of each shoe 34 a at a higher position. A U-shaped seal member 38 is fit in each retaining groove, and urged radially inwards by a leaf spring not shown fit in the retaining groove.
Front cover 35 includes a center support hole 35 a having a relatively large inside diameter; and four bolt holes not shown each receiving one of the axially extending bolts 37. These four bolt holes are arranged around the center support hole 35 a, facing respective ones of the bolt holes 34 b of shoes 34 a.
Rear cover 36 includes a toothed portion 36 a formed integrally on the rear side, and arranged to engage with the before-mentioned timing chain; and a center bearing hole 36 b having a relatively large inside diameter and extending axially through rear cover 36.
Vane member 32 includes a central vane rotor 32 a and a plurality of vanes 32 b projecting radially outwards from the vane rotor 32 a. In this example, four of the vanes 32 b are arranged at angular intervals of approximately 90 degrees circumferentially around vane rotor 32 a. Vane rotor 32 a is annular and includes a center bolt hole 14 a at the center. Vanes 32 b are integral with vane rotor 32 a. Vane member 32 is fixed to the front end of drive shaft 6 by a fixing bolt 139 extending axially through the center bolt hole 14 a of vane rotor 32 a.
The vane rotor 32 a includes a front side small diameter tubular portion supported rotatably by the center support hole 35 a of front cover 35, and a rear side small diameter tubular portion supported rotatably by the bearing hole 36 b of rear cover 36.
Three of the four vanes 32 b are smaller vanes shaped approximately like a relatively long rectangle, and the remaining one is a larger vane shaped like a relatively large trapezoid. The smaller vanes 32 b are approximately equal in circumferential width whereas the larger vane 32 b has a larger circumferential width greater than that of each of the smaller vanes 32 b so that a weight balance is attained as a whole of vane member 32. The four vanes 32 b of vane member 32 and the four shoes 34 a of housing 34 are arranged alternately in the circumferential direction around the center axis, as shown in FIGS. 7 and 8. Each vane 32 b includes an axially extending retaining groove receiving a U-shaped seal member 40 in sliding contact with the inside cylindrical surface of housing 34, and a leaf spring not shown for urging the seal member 40 radially outwards and thereby pressing the seal member 40 to the inside cylindrical surface of housing 34. Moreover, in one side of each vane 32 b facing in the direction opposite to the rotational direction of drive shaft 6, there are formed two circular recesses 32 c.
An advance fluid pressure chamber 41 and a retard fluid pressure chamber 42 are formed on both sides of each vane 32 b. Accordingly, there are four of the advance fluid pressure chambers 41 and four of the retard fluid pressure chambers 42.
The hydraulic circuit includes a first fluid passage 43 leading to the advance fluid pressure chambers 41 to supply and drain an advance fluid pressure of an operating oil to and from advance fluid pressure chambers 41; a second fluid passage 44 leading to the retard fluid pressure chambers 42 to supply and drain a retard fluid pressure of the operating oil to and from retard fluid pressure chambers 42; and a directional control valve or selector valve 47 connecting the first fluid passage 43 and second fluid passage 44 selectively with a supply passage 45 and a drain passage 46. A fluid pump 49 is connected with supply passage 45, and arranged to draw the hydraulic operating fluid or brake fluid or oil from an oil pan 48 of the engine, and to force the fluid into supply passage 45. Pump 49 is a one-way type pump. The downstream end of drain passage 46 is connected to oil pan 48, and arranged to drain the fluid to oil pan 48.
First and second fluid passages 43 and 44 include sections formed in a cylindrical portion 39 which is inserted, from a first end, through the small diameter tubular portion of vane rotor 32 a, into the support hole 32 d of vane rotor 32 a. A second end of the cylindrical portion 39 is connected with directional control valve 47.
Between the outside circumferential surface of the cylindrical portion 39 and the inside circumferential surface of support hole 32 d, there are provided three annular seal members 127 fixedly mounted on the cylindrical portion 39 near the forward end and arranged to seal the first and second fluid passages 43 and 44 off from each other.
First fluid passage 43 includes a passage section 43 a serving as a pressure chamber, and four branch passages 43 b connecting the passage section 43 a, respectively, with the four advance fluid pressure chambers 41. Passage section 43 a is formed in an end portion of support hole 32 d on the side of drive shaft 6. The four branch passages 43 b are formed in vane rotor 32 a and extend radially in vane rotor 32 a.
Second fluid passage 44 includes an axially extending passage section extending axially in the cylindrical portion 39 to a closed end; an annular chamber 44 a formed around the axially extending passage section near the closed end; and an L-shaped passage section 44 b connecting the annular chamber 44 a with each retard pressure chamber 42.
Directional control valve 47 of this example is a solenoid valve having four ports and three positions. A valve element inside the directional control valve 47 is arranged to alter the connection between first and second fluid passages 43 and 44 and the supply and drain passages 45 and 46 under the control of the controller 22.
When no control current is supplied to directional control valve 47 of exhaust VTC 2, directional control valve 47 is in its default position to hydraulically connect the supply passage 45 to second fluid passage 44 leading to retard fluid pressure chamber 42, and hydraulically connect the drain passage 46 to first fluid passage 43 leading to advance fluid pressure chamber 41. Directional control valve 47 includes a coil spring that mechanically holds a valve element of directional control valve 47 in such a default position that directional control valve 47 is in the default position described above.
The exhaust VTC 2 includes a lock mechanism disposed between vane member 32 and housing 34 for locking the vane member 32 in a predetermined rotational position with respect to housing 34 or allowing the rotation of vane member 32 with respect to housing 34. Specifically, this lock mechanism is disposed between rear cover 36 and the larger vane 32 b. The lock mechanism includes a slide hole 50, a lock pin 51, a lock recess 52 a, a spring retainer 53, and a coil spring 54, as shown in FIGS. 6 and 7. Slide hole 50 is formed in the larger vane 32 b, extending in the axial direction of drive shaft 6. Lock pin 51 is cup-shaped, disposed in slide hole 50, and slidably supported on slide hole 50. Lock recess 52 a is formed in a portion 52 fixed to a hole defined in rear cover 36, and arranged to receive a tip portion 51 a of lock pin 51. The tip portion 51 a is tapered. Spring retainer 53 is fixed to a bottom portion of slide hole 50. Coil spring 54 is retained by spring retainer 53, and arranged to bias the lock pin 51 toward the lock recess 52 a.
The lock recess 52 a is hydraulically connected to retard fluid pressure chamber 42 or pump 49 through a fluid passage not shown, and receives the hydraulic pressure in retard fluid pressure chamber 42 or the discharge pressure of the pump.
When vane member 32 is in its most retarded position with respect to housing 34, the lock pin 51 is biased by coil spring 54 toward lock recess 52 a so that the tip portion 51 a of lock pin 51 is fit in lock recess 52 a. The relative rotation between timing sprocket 33 and drive shaft 6 is thus locked. When lock recess 52 a receives the hydraulic pressure in retard fluid pressure chamber 42 or the discharge pressure of the oil pump, then lock pin 51 moves away from lock recess 52 a, so as to release drive shaft 6 with respect to timing sprocket 33.
Between one side surface of each vane 32 b and a confronting side surface of an adjacent one of the shoes 34 a, there are disposed a pair of coil springs 55 and 56 serving as biasing means for urging the vane member 32 in the retard rotational direction. In other words, coil springs 55 and 56 serve as a biasing device arranged to bias the exhaust VTC 2 in a direction to retard the opening timing and the closing timing of exhaust valves 5.
Though the two coil springs 55 and 56 are overlapped in FIGS. 7 and 8, the two coil springs 55 and 56 extend separately in parallel to each other. The two coil springs 55 and 56 have an equal axial length (coil length) which is longer than the spacing between the one side surface of the corresponding vane 32 b and the confronting side surface of the adjacent shoe 34 a. The two coil springs 55 and 56 are spaced with such an interaxis distance that the springs 55 and 56 do not contact each other even when the springs 55 and 56 are compressed to the maximum extent. The two coil springs 55 and 56 are connected through a retainer shaped like a thin sheet and fit in the recesses 32 c of the corresponding shoe 34 a.
The thus-constructed exhaust VTC 2 is controlled to operate as follows. When the engine is stopped or at rest, the controller 22 stops the output of the control current to directional control valve 47, so that the valve element of directional control valve 47 is placed in the default position as shown in FIG. 7 so as to allow fluid communication between supply passage 45 and second fluid passage 44 leading to retard fluid pressure chamber 42, and allow fluid communication between drain passage 46 and first fluid passage 43. Also, when the engine is at rest, the supplied fluid pressure is equal to zero, because oil pump 49 is also inoperative. Accordingly, vane member 32 is biased by coil springs 55, 56, so as to rotate in the counterclockwise direction about the axial direction of drive shaft 6 as viewed in FIG. 7. As a result, vane member 32 is brought into a position such that the larger vane 32 b is in contact with one confronting side surface of shoe 34 a. Drive shaft 6 is thus in the most retarded position with respect to timing sprocket 33. Simultaneously, the tip portion 51 a of lock pin 51 is inserted into lock recess 52 a, so as to prevent drive shaft 6 from rotating with respect to timing sprocket 33. The exhaust VTC 2 is thus mechanically and stably held in its default position for most retarded exhaust valve opening timing EVO and exhaust valve closing timing EVC. As described in detail below, when the exhaust VTC 2 is in the default position, exhaust valve closing timing EVC is at or close to top dead center, or in such a position with respect to top dead center as to enable startup of the engine. The default position is defined as a position where a subject mechanism is mechanically and stably held when de-energized or when no control signal is issued.
When the engine is started by turning on the ignition switch and cranking the crankshaft with a starter motor, then directional control valve 47 starts to receive a control signal from controller 22. However, immediately after the engine start, vane member 32 is still held in the most retarded position by means of the lock mechanism and coil springs 55, 56, because the discharge pressure of oil pump 49 is not yet sufficiently high. At this moment, directional control valve 47 allows fluid communication between supply passage 45 and second fluid passage 44, and between drain passage 46 and first fluid passage 43. Then, the oil pressure from oil pump 49 is raised and supplied through second fluid passage 44 to retard fluid pressure chambers 42, while the advance fluid pressure chambers 41 are held in a low pressure state in which no oil pressure is supplied, and the oil pressure is drained through drain passage 46 into oil pan 48.
After the discharge pressure of oil pump 49 is increased sufficiently, the controller 22 can control the position of vane member 32 by means of directional control valve 47. When the hydraulic pressure in retard fluid pressure chamber 42 rises, then the hydraulic pressure in lock recess 52 a of the lock mechanism rises so as to move the lock pin 51 out of lock recess 52 a. This allows rotation of vane member 32 with respect to housing 34.
For example, when the engine is at idle after warmed up, the directional control valve 47 is controlled to allow fluid communication between supply passage 45 and second fluid passage 44 and between drain passage 46 and first fluid passage 43. Accordingly, the oil pressure discharged by pump 49 is supplied through second fluid passage 44 to retard fluid pressure chamber 42, while the oil pressure is drained from advance fluid pressure chamber 41 through first fluid passage 43 and drain passage 46 to oil pan 48 so that advance fluid pressure chamber 41 remains in a low-pressure state. Accordingly, vane member 32 is rotated in the counterclockwise direction by the increased pressures in retard fluid pressure chambers 42 and the elastic forces of coil springs 55 and 56, as viewed in FIG. 7. Consequently, drive shaft 6 rotates to the retard side, relative to timing sprocket 33, retarding the exhaust valve opening timing EVO and exhaust valve closing timing EVC.
When the engine enters a predetermined low speed and middle load region thereafter, then the controller 22 operates the directional control valve 47 to the position connecting the supply passage 45 with first fluid passage 43 and connecting the drain passage 46 with second fluid passage 44. Therefore, the oil pressure in retard fluid pressure chambers 42 is decreased by return through second fluid passage 44 and drain passage 46 to oil pan 48, whereas the oil pressure in advance fluid pressure chambers 41 is increased by supply of the oil pressure. Vane member 32 rotates in the clockwise direction by the high pressure in advance fluid pressure chambers 41, against the elastic forces of coil springs 55 and 56, and thereby shifts the relative rotational phase of drive shaft 6 relative to timing sprocket 33 to the advance side, as shown in FIG. 8. Then, the relative rotational phase is held at any desired position by keeping the directional control valve 47 in its neutral position.
When the engine enters a predetermined middle and high speed region from the low speed region, then directional control valve 47 is controlled similarly as when the engine is at idle after warmed up. Accordingly, the oil pressure in advance fluid pressure chambers 41 decreases, the oil pressure in retard fluid pressure chambers 42 increases, and hence the resultant of the hydraulic pressures and the elastic forces of coil springs 55 and 56 causes the vane member 32 to shift the relative rotational phase of drive shaft 6 relative to timing sprocket 33 to the retard side, as shown in FIG. 7. Then, directional control valve 47 is controlled to be in its neutral position, so that the vane member 32 is fixed relative to housing 34.
The following describes the intake VTC 3. As shown in FIGS. 9 and 10, the intake VTC 3 of this example is of a vane type like the exhaust VTC 2. The intake VTC 3 includes a timing sprocket 60 for transmitting rotation from the crankshaft to an intake camshaft 59; a vane member 61 which is fixed to one end of intake camshaft 59 and received rotatably in the timing sprocket 60; and a hydraulic circuit to rotate vane member 61 in the forward and reverse directions by the use of an oil pressure.
Timing sprocket 60 includes a housing 62 receiving the vane member 61 rotatably; a front cover shaped like a circular disk and arranged to close a front opening of housing 62; and a rear cover shaped approximately like a circular disk and arranged to close a rear opening of housing 62. Housing 62 is sandwiched between the front and rear covers, and joined with these covers to form a unit, by four small diameter bolts 63 extending in the axial direction of intake camshaft 59 Housing 62 is in the form of a hollow cylinder having the front and rear openings. Housing 62 includes a plurality of shoes 62 a projecting radially inwards from the inside circumferential surface and serving as a partition. In this example, four of the shoes 62 a are arranged at intervals of about 90 degrees. The rear cover includes a toothed portion 60 a formed integrally on the rear side, and arranged to engage with a timing chain, as in the case of the exhaust VTC 2.
Vane member 61 includes a central vane rotor 61 a and a plurality of vanes 61 b projecting radially outwards from the vane rotor 61 a. In this example, four of the vanes 61 b are arranged at angular intervals of approximately 90 degrees circumferentially around vane rotor 61 a. Vane rotor 61 a is annular and includes a center bolt hole at the center. Vanes 61 b are integral with vane rotor 61 a. Vane member 61 is fixed to the front end of intake camshaft 59 by a fixing bolt 64 extending axially through the center bolt hole of vane rotor 61 a. An advance fluid pressure chamber 65 and a retard fluid pressure chamber 66 are formed on both sides of each vane 61 b. Accordingly, there are four of the advance fluid pressure chambers 65 and four of the retard fluid pressure chambers 66.
The hydraulic circuit of the intake VTC 3 has a construction identical to the construction of the hydraulic circuit of the exhaust VTC 2, except that a directional control valve corresponding to directional control valve 47 has three positions reversed with respect to a vertical line as viewed in FIG. 6. The hydraulic circuit includes a first fluid passage leading to the advance fluid pressure chambers 65 to supply and drain an advance fluid pressure of an operating oil to and from advance fluid pressure chambers 65; a second fluid passage leading to the retard fluid pressure chambers 66 to supply and drain a retard fluid pressure of the operating oil to and from retard fluid pressure chambers 66; and the directional control valve connecting the first fluid passage and second fluid passage selectively with a supply passage and a drain passage. The directional control valve includes a movable valve element inside, and operates under control of controller 22.
The directional control valve of the intake VTC 3 is arranged to connect the supply passage to the first fluid passage leading to advance fluid pressure chambers 65, and connect the drain passage to the second fluid passage leading to retard fluid pressure chambers 66, when no control current is supplied to the directional control valve. The directional control valve includes a coil spring arranged to mechanically bias the valve element toward this default position.
The intake VTC 3 includes a lock mechanism disposed between vane member 61 and housing 62 for locking or allowing the rotation of vane member 61 with respect to housing 62. Specifically, this lock mechanism is disposed between the rear cover 36 and the larger vane 62 b. The lock mechanism includes a slide hole, a lock pin 67, a lock recess, a spring retainer, and a coil spring, similarly as in the case of the exhaust VTC 2. When the engine is at rest, and the vane member 61 is located in the most advanced position shown in FIG. 9, then the lock pin 67 is inserted and fitted in the lock recess under the elastic force of the coil spring, so as to prevent the vane member 61 from rotating relative to housing 62, and thus stably hold the vane member 61.
Between one side surface of each vane 62 b and a confronting side surface of an adjacent one of the shoes 62 a, there are disposed a pair of coil springs 68 and 69 serving as biasing means for urging the vane member 61 in the advance rotational direction. In other words, coil springs 68 and 69 serve as a biasing device arranged to bias the intake VTC 3 in a direction to advance the intake valve opening timing IVO and intake valve closing timing IVC. When the oil pump supplies no hydraulic pressure or a lower hydraulic pressure below a predetermined level, for example, when the engine is at rest, or immediately after the engine is started, then the vane member 61 is biased in the clockwise direction as viewed in FIG. 9, so as to rotate the intake camshaft 59 to the most advanced position.
The following describes how the variable valve actuating system operates. When the engine is at rest before started up after stopped, then the vane member 61 of the intake VTC 3 is mechanically and stably positioned and held at the position shown in FIG. 9 by the elastic force of coil springs 68, 69 and the lock mechanism. Accordingly, the intake VTC 3 is positioned so that the intake valve opening timing IVO and intake valve closing timing IVC are most advanced and held mechanically and stably. On the other hand, when the engine is at rest, then the vane member 32 of the exhaust VTC 2 is mechanically and stably positioned and held at the position shown in FIG. 7 by the elastic force of coil springs 55, 56 and the lock mechanism. Accordingly, the exhaust VTC 2 is positioned so that the exhaust valve opening timing EVO and exhaust valve closing timing EVC are most retarded and held mechanically and stably. Moreover, the exhaust VEL 1 is positioned by the elastic force of coil spring 30 so that the operating angle and lift of exhaust valves 5 are set to the small operating angle D1 and small lift L1.
When the exhaust VEL 1, exhaust VTC 2 and intake VTC 3 are located in such default operating positions, no valve overlap is produced between the exhaust valve closing timing EVC, which is equal to EVC1, and intake valve opening timing IVO, which is equal to IVO1, as shown in FIG. 11. The exhaust valve closing timing EVC is thus mechanically held sufficiently close to top dead center, or in a position slightly advanced from top dead center by a small angle of θEVC1. The exhaust valve opening timing EVO is mechanically held in a position EVO1 which is sufficiently retarded away from bottom dead center. On the other hand, the intake valve opening timing IVO is mechanically held most advanced (equal to IVO1) which is slightly retarded from top dead center by an angle of θIVO1, and the intake valve closing timing IVC is mechanically held close to bottom dead center, or slightly advanced from bottom dead center. In FIG. 11, al indicates the phase (maximum lift phase) of exhaust valves 5, and β1 indicates the phase of intake valves 4. These mechanically stable characteristics of operation of intake valves 4 and exhaust valves 5 are presented, when the variable valve actuating system is de-energized, or no control signal is outputted
When the engine is started under the condition described above, then the phase of exhaust valves 5 is still held most retarded by the elastic force of coil springs 55, 56 and the lock mechanism, and the phase of intake valves 4 is still held most advanced by the elastic force of coil springs 68, 69 and the lock mechanism, because the hydraulic pressures are still low.
The setting shown in FIG. 11, in which the exhaust valve closing timing EVC is close to top dead center, is effective for reducing the quantity of residual burned gas in combustion chamber 04, and thereby allowing improved stable combustion in combustion chamber 04 even at the initial stage of engine start when the engine temperature is low.
The reduction of the quantity of residual burned gas results in increase of the quantity of fresh air (air fuel mixture) entering the cylinder, and thereby results in increase of the engine output torque. This is effective for canceling the increase of friction of the engine at cold start, and thereby further improving the combustion and startability.
Also, the setting shown in FIG. 11, in which the exhaust valve opening timing EVO is held retarded away from bottom dead center, is effective for ensuring a sufficient period of combustion, and reducing the quantity of exhaust emissions exiting a tail pipe, even when a catalytic converter is in a cold state, and is not enough activated.
Moreover, holding the operating angle and lift of exhaust valves 5 to the minimum setpoint is effective for reducing the level of friction of moving parts such as exhaust valves 5, and thereby allowing the engine to smoothly speed up during cranking operation.
Thus, the variable valve actuating system can achieve smooth engine start with improved combustion
The intake valve closing timing IVC1 close to bottom dead center is effective for increasing the fresh air charging efficiency, because of no delay of entrance of intake air.
In order to set the intake valve opening timing IVO close to top dead center while setting the intake valve closing timing IVC close to bottom dead center, an intake VEL which is constructed as the exhaust VEL 1 may be added to vary the operating angle of intake valves 4.
When the engine enters a low or middle load region after warmed up, then the lock mechanisms are released so as to allow free relative rotation of vane member 32 and vane member 61. Then, the exhaust VEL 1 is controlled to provide the medium lift L2 and medium operating angle D2, and the exhaust VTC 2 is controlled to rotate the vane member 32 to the advance side. Thus, as shown in FIG. 12, the exhaust valve closing timing EVC is set advanced to EVC2 by an angle of θEVC2 from top dead center, and the exhaust valve opening timing EVO is somewhat advanced to EVO2 with respect to bottom dead center.
Because the exhaust valve closing timing EVC is sufficiently advanced with respect to top dead center, exhaust valves 5 are closed before burned gas is completely exhausted from the cylinder, so that a large quantity of burned gas remains in combustion chamber 04. The large quantity of residual burned gas is effective for reducing the pumping loss, increasing the ratio of specific heat of the air fuel mixture in combustion chamber 04, and thereby improving the fuel efficiency. The large quantity of residual burned gas may be used to control the incylinder temperature of combustion chamber 04 so as to enable self-ignition combustion, for further improving the fuel efficiency.
The quantity of residual burned gas may be increased by significantly retarding the exhaust valve closing timing EVC beyond top dead center so as to increase the valve overlap between the exhaust valve closing timing EVC and the intake valve opening timing IVO. However, with this method, it is difficult to desirably control the quantity of residual burned gas, because the burned gas flows inversely to the intake port side, and then flows into combustion chamber 04 again. Moreover, the burned gas reentering the combustion chamber 04 has a lower temperature, which may adversely affect the combustion process.
In contrast, the method implemented by advancing the exhaust valve closing timing EVC with respect to top dead center as shown in FIG. 12 can directly control the quantity of residual burned gas by closing the exhaust valves 5 during exhaust stroke so as to prevent a part of the burned gas from flowing out from combustion chamber 04. This method is advantageous in controllability and combustion process, because the residual burned gas maintains a higher temperature.
The exhaust valve opening timing EVO2 which is set advanced with respect to bottom dead center is effective for reducing the exhaust loss so as to maintain the engine output torque, and allowing the hot burned gas as exhaust gas to reach a catalytic converter so as to rapidly heat up and activate a catalyst for full purification performance.
The exhaust valve opening timing EVO2 and the exhaust valve closing timing EVC2 shown in FIG. 12 are achieved by the operating angle D2 and the lift L2 set by the exhaust VEL 1.
On the other hand, in the intake VTC 3, vane member 61 is rotated to the retard side as shown in FIG. 10, so that the intake valve opening timing IVO and the intake valve closing timing IVC are set retarded as shown in FIG. 12. Specifically, the intake valve opening timing IVO2 is retarded from top dead center by an angle of θIVO2. Since intake valves 4 are opened when the piston is located away from top dead center and the incylinder pressure of combustion chamber 04 has fallen, the residual burned gas in combustion chamber 04 is prevented from inversely flowing into the intake port.
If θIVO2 is substantially equal to or larger than θEVC2, the residual burned gas in combustion chamber 04 is reliably prevented from inversely flowing into the intake port.
The intake valve closing timing IVC, which is set to IVC2 as shown in FIG. 12, away from bottom dead center, is effective for reducing the pumping loss so as to improve the fuel efficiency. The large quantity of residual burned gas in combustion chamber 04 is also effective for reducing the pumping loss so as to further improve the fuel efficiency.
Self ignition is possible without ignition plug 05, by increasing the quantity of residual burned gas by advancing the exhaust valve closing timing EVC2 with respect to top dead center, precisely controlling the incylinder temperature at top dead center to a predetermined temperature, and injecting fuel from fuel injection valve 07 during intake stroke. Therefore, this valve timing setting can be used for so called premixed gasoline compression ignition engines. For example, a premixed gasoline compression ignition engine is published by Michihiko Tabata, “Gasoline HCCI Combustion Technology and Combustion Sensing”, at Society of Automotive Engineers of Japan Symposium No. 17-04 “Gasoline Engine for Future”, Dec. 10, 2004. Since this engine is based on unified premixed combustion (not stratified lean combustion) with a lean air fuel ratio, it is possible to improve the fuel efficiency while minimizing the quantity of NOx.
When the load (torque) of the engine is to be desired to increase, the intake valve closing timing IVC may be retarded from IVC2 toward bottom dead center so as to increase the intake air charging efficiency, or the exhaust valve closing timing EVC may be retarded toward top dead center so as to reduce the quantity of residual burned gas. When the decrease of the quantity of residual burned gas may disable compression ignition, the system may employ a combustion mode based on spark ignition by ignition plug 05, or employ both of spark ignition and compression ignition.
When the vehicle is accelerating, or the accelerator opening is set large in response to depression of an accelerator pedal, the exhaust VEL 1 is controlled to expand the operating angle and lift of exhaust valves 5 from the medium operating angle D2 and medium lift L2 to the large operating angle D3 and large lift L3. Also, the exhaust VTC 2 is controlled to retard the peak lift phase of exhaust valves 5 slightly on the retard side of the most advanced position (from phase α2 to α3, as shown in FIGS. 12 and 13).
Specifically, as shown in FIG. 13, with the large operating angle D3 set, the exhaust valve closing timing EVC is set retarded close to top dead center (EVC3), and the exhaust valve opening timing EVO is further advanced from EVO2 to EVO3. The retarded exhaust valve closing timing EVC3 is effective for reducing the quantity of residual burned gas in combustion chamber 04, and thereby improving the fresh air charging efficiency. The advanced exhaust valve opening timing EVO3 is effective for minimizing the exhaust loss which tends to increase due to increase in the quantity of exhaust gas under high load, and thereby increasing the engine output torque.
On the other hand, the intake VTC 3 is controlled to advance the phase of intake valves 4, as shown in FIG. 13. The intake valve closing timing IVC3 close to bottom dead center is effective for increasing the intake air charging efficiency so as to increase the engine output torque. Since the combustion pressure is high so that compression ignition may cause unstable combustion, ignition may be implemented by ignition plug 05 so as to stably generate a large output torque.
FIG. 14 shows a control process performed by controller 22. First, at Step S1, controller 22 determines whether or not an ignition key switch is OFF. When the answer to step S1 is negative (NO), then controller 22 returns from this control process. On the other hand, when the answer to step S1 is affirmative (YES), then controller 22 proceeds to Step S2.
At Step S2, while the engine control unit outputs an engine stop signal for fuel injection valve 07 to be de-energized so as to cut fuel supply, controller 22 stops sending control signals, so as to de-energize the exhaust VEL 1, exhaust VTC 2 and intake VTC 3. Accordingly, the engine speed decreases.
Subsequently, at Step S3, controller 22 allows the exhaust VEL 1, exhaust VTC 2 and intake VTC 3 to move to their default operating positions, and specifically allows ball nut 24, vane member 32 and vane member 61 to mechanically moves to the respective default operating positions under the forces of coil springs 30, 55, 56, 68 and 69.
Specifically, the exhaust VEL 1 is mechanically and stably held in the position for the minimum lift and operating angle by the force of coil spring 30 and the reaction force of valve springs. The exhaust VTC 2 is moved to the most retarded position so as to maximally retard the exhaust valve opening timing EVO and exhaust valve closing timing EVC, and stably held by the lock mechanism The intake VTC 3 changes the phase of intake valves 4 (the intake valve opening timing IVO and intake valve closing timing IVC) to the most advanced position and holds the phase mechanically and stably by the lock mechanism.
Without coil spring 30 and coil springs 55, 56, the ball nut 24 of the exhaust VEL 1 and the vane member 32 of the exhaust VTC 2 can be moved to the respective default positions by the reaction force of the valve springs. However, coil spring 30 and coil springs 55, 56 serve to reliably and stably bias and hold ball nut 24 and vane member 32, respectively.
On the other hand, in the intake VTC 3, coil springs 68, 69 serve to move the vane member 61 to the most advanced position against the force of the valve springs.
Subsequently, at Step S4, controller 22 allows the engine to be completely stopped.
Steps S5 to S12 show a process for cases where the engine is restarted. First, at Step S5, controller 22 determines whether or not the ignition key switch is ON.
When the answer to Step S5 is NO, controller 22 returns from this control process. On the other hand, when the answer to Step S5 is YES, controller 22 proceeds to Step S6, at which controller 22 starts to crank the engine. Even if the exhaust VEL 1, exhaust VTC 2 and intake VTC 3 are deviated from the default positions, they tend to approach or move to the default positions during the cranking operation, because they are mechanically stable in the default positions.
At Step S7, controller 22 outputs control signals to electric motor 20 and directional control valve 47 so as to move the exhaust VEL 1, exhaust VTC 2 and intake VTC 3 to the default positions, and hold them in the default positions. In general, during cranking operation, hydraulic pressures in the hydraulic circuit for the exhaust VTC 2 and intake VTC 3 are low, and mechanical friction is large. However, the exhaust VEL 1, exhaust VTC 2 and intake VTC 3 can be rapidly moved to the default positions, because the exhaust VEL 1, exhaust VTC 2 and intake VTC 3 are currently already close to the default positions before Step S7.
The exhaust VEL 1, exhaust VTC 2 and intake VTC 3 may be controlled to setpoints which are slightly deviated from the default positions. For example, the exhaust VEL 1 may be in a position for larger lift and operating angle than the minimum lift and operating angle. This can be easily implemented even when the engine is in cold state, because the exhaust VEL 1 is electrically actuated or driven by electric motor 20 (not by hydraulic pressure). This active operation allows suitable control of lift and operating angle, even when stoppers are worn which restrict rotational movement of control shaft 17 within allowable maximum angular positions. The setpoint for the exhaust VEL 1 at this time may be adjusted according to engine temperature.
At Step S8, controller 22 determines whether or not a reference cranking speed is reached. When the answer to step S8 is NO, controller 22 returns to Step S7. On the other hand, when the answer to step S8 is YES, controller 22 proceeds to Step S9.
At Step S9, controller 22 performs a complete explosion control by controlling the quantity of fuel injected by fuel injection valve 07 and the timing of fuel ignition by ignition plug 05. According to the valve setting at the time, stable combustion is performed for engine start with a small quantity of residual burned gas in combustion chamber 04 and a small quantity of exhaust emissions. Also, the engine smoothly speeds up, because the level of friction of the valve system is low according to the small lift and operating angle of intake valves 4.
Subsequently, at Step S10, controller 22 determines on the basis of a timer value whether or not a reference time is elapsed. When the answer to step S10 is YES, then controller 22 proceeds to step S11. The timer value may be started to increase when the cranking operation is started, or when the complete explosion control is started. The reference time may be corrected according to the engine temperature measured by coolant temperature sensor 06. For example, the reference time is reduced when the engine temperature is high, and is increased when the engine temperature is low. When the engine temperature is low, the exhaust valve opening timing EVO may be slightly advanced, in order to promote rise of the temperature of the catalyst.
At Step S11, controller 22 measures a current engine operating state on the basis of accelerator opening, engine speed, engine temperature, etc.
After the engine temperature rises so that the engine is completely warmed up, at Step S12, controller 22 performs a variable valve control by the exhaust VEL 1, exhaust VTC 2 and intake VTC 3, according to the engine operating state, as shown in FIGS. 12 and 13.
Since controller 22 is built in the engine control unit, after the engine control unit outputs a cranking signal, controller 22 can control the exhaust VEL 1, exhaust VTC 2 and intake VTC 3 with improved response, by outputting controls signals to them with no time lag due to signal communication via lines.
As described above, when the internal combustion engine is at rest, or when the exhaust VEL 1 and exhaust VTC 2 are de-energized, the exhaust VEL 1 as the first valve actuating mechanism automatically or mechanically sets the exhaust valve operating angle relatively small, or at or close to a minimum setpoint, and the exhaust VTC 2 as the second valve actuating mechanism automatically or mechanically sets the exhaust valve opening timing and the exhaust valve closing timing relatively retarded or, at or close to a most retarded setpoint, so as to set the exhaust valve closing timing at or close to top dead center so as to enable startup of the internal combustion engine. After startup of the internal combustion engine, the exhaust VEL 1 and exhaust VTC 2 maintain their respective operating positions as at startup of the internal combustion engine. The exhaust VTC 2 includes a lock mechanism adapted to be operated by a hydraulic pressure generated by an oil pump driven by the internal combustion engine, and at startup of the internal combustion engine, the lock mechanism maintains the exhaust valve opening timing and the exhaust valve closing timing as when the internal combustion engine is at rest.
FIG. 15 schematically shows the variable valve actuating system according to the second embodiment. This variable valve actuating system is based on that of the first embodiment, and modified in that the ball screw transmission mechanism 21 of the exhaust VEL 1 further includes a coil spring 30 a for biasing the ball nut 24 away from electric motor 20 so as to increase the lift and operating angle of exhaust valves 5. When the engine is stopped or at rest, ball nut 24 is mechanically held in an intermediate position by the urging forces of coil springs 30 and 30 a. Accordingly, when the engine is stopped, the lift and operating angle of exhaust valves 5 are set and stably held at a medium setpoint.
According to the second embodiment, the exhaust valve closing timing EVC1 shown in FIG. 11 is set closer to top dead center. This is more effective for reducing the quantity of residual burned gas in combustion chamber 04, and thereby further improving the startability of the engine.
The variable valve actuating systems according to the present embodiments are applicable to premixed gasoline compression ignition engines in addition to typical spark ignition engines, as described above.
The variable valve actuating systems according to the present embodiments are applicable to compression ignition engines such as diesel engines. In such cases, the variable valve actuating system can also reduce the quantity of residual burned gas in combustion chamber 04 at engine start, and thereby improving the combustion process at cold start.
The lock mechanism of the exhaust VTC may be omitted, because the exhaust VTC 2 can be stably held in the most retarded position only by the coil springs.
In the second embodiment, the exhaust VTC 2 may further include a coil spring for biasing the vane member 32 to the advance side in addition to coil springs 55, 56, so that the vane member 32 is stably held in an intermediate position so as to set the lift and operating angle of exhaust valves 5 medium as a default setpoint.
Thus, when the internal combustion engine is at rest, the exhaust VEL 1 as the first valve actuating mechanism may be automatically held in a default operating position to set the exhaust valve operating angle smaller by a predetermined amount than a maximum setpoint, and the exhaust VTC 2 as the second valve actuating mechanism may be automatically held in a default operating position to set the exhaust valve lift phase more retarded by a predetermined amount than a most advanced setpoint. Moreover, at an initial stage of startup of the internal combustion engine, controller 22 may output a control signal so as to maintain the first valve actuating mechanism in or close to the default operating position, and maintain the second valve actuating mechanism in or close to the default operating position.
The foregoing embodiments may be modified as follows. The exhaust VEL 1 may be constructed to vary the operating angle and lift of exhaust valves 5 stepwise.
The exhaust VTC 2 and intake VTC 3 may be driven by an electronic system as shown in U.S. Pat. No. 6,502,537 corresponding to Japanese Patent Application Publication No. 2002-227615 instead of the hydraulic system.
This application is based on a prior Japanese Patent Application No. 2007-331557 filed on Dec. 25, 2007. The entire contents of this Japanese Patent Application No. 2007-331557 are hereby incorporated by reference.
Although the invention has been described above by reference to certain embodiments of the invention, the invention is not limited to the embodiments described above. Modifications and variations of the embodiments described above will occur to those skilled in the art in light of the above teachings. The scope of the invention is defined with reference to the following claims.

Claims

1. A variable valve actuating apparatus for an internal combustion engine, comprising:

a first valve actuating mechanism arranged to vary at least an exhaust valve operating angle of the internal combustion engine, wherein when the internal combustion engine is at rest, the first valve actuating mechanism sets the exhaust valve operating angle relatively small; and

a second valve actuating mechanism arranged to vary an exhaust valve opening timing and an exhaust valve closing timing of the internal combustion engine, wherein when the internal combustion engine is at rest, the second valve actuating mechanism sets the exhaust valve opening timing and the exhaust valve closing timing relatively retarded,

wherein when the internal combustion engine is at rest, the first and second valve actuating mechanisms set the exhaust valve closing timing at or close to top dead center so as to enable startup of the internal combustion engine.

2. The variable valve actuating apparatus as claimed in claim 1, wherein the first valve actuating mechanism is arranged to automatically set the exhaust valve operating angle relatively small when the internal combustion engine is at rest.

3. The variable valve actuating apparatus as claimed in claim 1, wherein the second valve actuating mechanism is arranged to automatically set the exhaust valve opening timing and the exhaust valve closing timing relatively retarded when the internal combustion engine is at rest.

4. The variable valve actuating apparatus as claimed in claim 3, wherein the second valve actuating mechanism includes a biasing device arranged to mechanically bias a movable member in a direction to retard the exhaust valve opening timing and the exhaust valve closing timing.

5. The variable valve actuating apparatus as claimed in claim 1, wherein the second valve actuating mechanism includes a lock mechanism adapted to be operated by a hydraulic pressure generated by an oil pump driven by the internal combustion engine, and wherein at startup of the internal combustion engine, the lock mechanism maintains the exhaust valve opening timing and the exhaust valve closing timing as when the internal combustion engine is at rest.

6. The variable valve actuating apparatus as claimed in claim 1, wherein the first valve actuating mechanism is arranged to vary an exhaust valve lift of the internal combustion engine according to variation of the exhaust valve operating angle.

7. The variable valve actuating apparatus as claimed in claim 1, wherein the internal combustion engine is a compression ignition engine.

8. The variable valve actuating apparatus as claimed in claim 1, wherein the internal combustion engine is one of a spark ignition engine and a compression ignition engine that uses both of compression ignition and spark ignition.

9. A variable valve actuating apparatus for an internal combustion engine, comprising:

a first valve actuating mechanism arranged to vary at least an exhaust valve operating angle of the internal combustion engine, wherein when de-energized, the first valve actuating mechanism mechanically sets the exhaust valve operating angle at or close to a minimum setpoint; and

a second valve actuating mechanism arranged to advance or retard an exhaust valve opening timing and an exhaust valve closing timing of the internal combustion engine, wherein when de-energized, the second valve actuating mechanism mechanically sets the exhaust valve opening timing and the exhaust valve closing timing at or close to a most retarded setpoint.

10. The variable valve actuating apparatus as claimed in claim 9, wherein the first valve actuating mechanism is arranged to mechanically set the exhaust valve operating angle at the minimum setpoint when de-energized.

11. The variable valve actuating apparatus as claimed in claim 9, wherein the second valve actuating mechanism is arranged to mechanically set the exhaust valve opening timing and the exhaust valve closing timing at the most retarded setpoint when de-energized.

12. The variable valve actuating apparatus as claimed in claim 11, wherein the second valve actuating mechanism includes a biasing device arranged to mechanically bias a movable member in a direction to retard the exhaust valve opening timing and the exhaust valve closing timing.

13. The variable valve actuating apparatus as claimed in claim 9, wherein the second valve actuating mechanism includes a lock mechanism adapted to be operated by a hydraulic pressure generated by an oil pump driven by the internal combustion engine, and wherein at startup of the internal combustion engine, the lock mechanism maintains the exhaust valve opening timing and the exhaust valve closing timing as when the internal combustion engine is at rest.

14. The variable valve actuating apparatus as claimed in claim 9, wherein the internal combustion engine is a compression ignition engine.

15. The variable valve actuating apparatus as claimed in claim 9, wherein the internal combustion engine is one of a spark ignition engine and a compression ignition engine that selectively uses one of compression ignition and spark ignition.

16. A controller for a variable valve actuating apparatus for an internal combustion engine, the variable valve actuating apparatus comprising a first valve actuating mechanism arranged to vary at least an exhaust valve operating angle of the internal combustion engine, and a second valve actuating mechanism arranged to vary an exhaust valve opening timing and an exhaust valve closing timing of the internal combustion engine, the controller comprising:

a section connected for signal communication therewith to the first and second valve actuating mechanisms, wherein at startup of the internal combustion engine, the section sets by the first valve actuating mechanism the exhaust valve operating angle relatively small, and sets by the second valve actuating mechanism the exhaust valve opening timing and the exhaust valve closing timing relatively retarded, so as to set the exhaust valve closing timing at or close to top dead center for enabling startup of the internal combustion engine.

17. The controller as claimed in claim 16, wherein after startup of the internal combustion engine, the first and second valve actuating mechanisms maintain their respective operating positions as at startup of the internal combustion engine.

18. The controller as claimed in claim 17, wherein the first valve actuating mechanism is electrically driven.

19. The variable valve actuating apparatus as claimed in claim 16, wherein the internal combustion engine is a compression ignition engine.

20. The controller as claimed in claim 16, wherein the section is built in an engine control unit for the internal combustion engine.

21. A variable valve actuating apparatus for an internal combustion engine, comprising:

a first valve actuating mechanism arranged to vary at least an exhaust valve operating angle of the internal combustion engine, wherein when the internal combustion engine is at rest, the first valve actuating mechanism is automatically held in a default operating position to set the exhaust valve operating angle smaller by a predetermined amount than a maximum setpoint;

a second valve actuating mechanism arranged to vary an exhaust valve lift phase of the internal combustion engine, wherein when the internal combustion engine is at rest, the second valve actuating mechanism is automatically held in a default operating position to set the exhaust valve lift phase more retarded by a predetermined amount than a most advanced setpoint; and

a controller connected for signal communication therewith to the first and second valve actuating mechanisms, wherein at an initial stage of startup of the internal combustion engine, the controller outputs a control signal so as to maintain the first valve actuating mechanism in or close to the default operating position, and maintain the second valve actuating mechanism in or close to the default operating position.