JP2011085078A - Internal combustion engine control device and internal combustion engine control system - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide an internal combustion engine control device reducing vibration in an early stage of cranking while maintaining good restarting properties of an engine. <P>SOLUTION: A crankshaft is controlled in a range where intake valves of all cylinders are put in close valve states and intake valves are controlled to the minimum operation angles D1 during engine stop period in steps 1-6. When the engine is under restart conditions in step 7, large operation angles D3 of the intake valves are set to start target operation angle values in step 12 if it is determined that engine temperature T is higher than first temperature T1 and second temperature T2 in step 9, 11. After that, operation angles of the intake valves are changed over to the start target operation angle values Dt in step 14. Then, cranking is started in step 15, and complete explosion control such as fuel injection and ignition is executed in step 17 after confirmation that the operation angles became the target operation angle values Dt in step 16. <P>COPYRIGHT: (C)2011,JPO&INPIT

Description

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

  As a conventional control device for an internal combustion engine, for example, one described in Patent Document 1 below is known.

  In brief, this control device advances the center phase of the lift / operating angle of the intake valve and the lift changing mechanism that can continuously change the valve lift amount and operating angle of the intake valve as a variable valve mechanism. A phase change mechanism capable of changing to the angle side and the retard side is controlled.

  When the intake drive shaft rotates in conjunction with the crankshaft, the swing cam swings via the lift changing mechanism, and the swing cam is made between the swing cam and the intake valve using hydraulic pressure. A hydraulic lash adjuster is installed to make the gap with the cam zero. The valve lift characteristics of the intake valves are set so that the all-cylinder zero lift state in which the valve lift amounts of the intake valves of all the cylinders to which the hydraulic lash adjuster is applied is almost zero when the engine is stopped. ing.

  That is, paying attention to the fact that the rotation stop position of the crankshaft of the four-cylinder internal combustion engine is near the middle of the piston top dead center (TDC) and bottom dead center (BDC) in the compression stroke, the rotation stop position of the crankshaft Is set so that the intake valves of all the cylinders are substantially in a zero lift state.

As a result, when the engine is stopped, the hydraulic lash adjusters of all the cylinders are prevented from shrinking due to the leakage of hydraulic oil, and variations in the valve lift amount among the cylinders are eliminated, so that good restartability is obtained. It has become.
JP 2003-56316 A

  However, in the conventional control device, as described above, it is assumed that the rotation stop position of the crankshaft is in the vicinity of the middle of the compression stroke TDC and BDC. Since the opening / closing timing (valve timing) of the intake valve is set so as to be in the lift state, the peak lift phase of the intake valve is in the vicinity of TDC or BDC as shown in FIGS. Become. Therefore, it is necessary to largely operate the phase changing mechanism at the time of cranking, but this phase changing mechanism has poor operation response at extremely low rotation speed, and takes time until the first explosion of the engine. As a result, good restartability cannot be obtained.

  The present invention has been devised in view of the technical problem of the conventional control device, and is good by controlling the lift characteristics of the intake valve in accordance with the state of the engine valve prior to cranking rotation. The purpose is to obtain restartability.

  According to the first aspect of the present invention, in particular, when the internal combustion engine is stopped, the variable valve device controls the valve closing period in which the intake valves of all the cylinders are in the non-lift state, and the crank position changing mechanism. The crankshaft is controlled so that the stop position of the crankshaft is within the valve closing period, and when the internal combustion engine is restarted, the variable valve device is set to a start lift characteristic of the intake valve in accordance with the state of the engine prior to cranking rotation. It is characterized by controlling to become.

  According to a second aspect of the present invention, when the internal combustion engine is stopped, the variable valve device and the crank position changing mechanism are controlled so that the spring load from the valve spring does not act on the control shaft, and the internal combustion engine is restarted. In some cases, prior to cranking rotation, the position of the control shaft is controlled to be a position suitable for starting.

  The invention according to claim 3 relates to a control system for an internal combustion engine. When the internal combustion engine is stopped, the variable valve device controls the intake valve of all the cylinders to be in a non-lift state so as to be closed. The crankshaft changing mechanism is controlled so that the crankshaft rotation stop position is within the valve closing period, and when the internal combustion engine is restarted, the variable valve gear is set in accordance with the state of the engine prior to cranking rotation. It is characterized by controlling toward the starting lift characteristic of the intake valve.

1 is a schematic view of an internal combustion engine provided for an embodiment of a variable valve operating apparatus according to the present invention. It is a perspective view which shows the intake VEL and intake VTC which are provided to this embodiment. A and B are operation explanatory views at the time of small lift control by the variable lift mechanism. A and B are operation explanatory diagrams at the time of maximum lift control by the variable lift mechanism. It is a valve lift amount of an intake valve in this embodiment, an operating angle, and a valve timing characteristic view. It is a longitudinal cross-sectional view of the intake VTC provided for this embodiment. FIG. 7 is a cross-sectional view taken along line AA of FIG. 6 showing a maximum retard angle control state by the intake VTC. FIG. 7 is a cross-sectional view taken along line AA of FIG. 6 showing a maximum advance angle control state by the intake VTC. It is a characteristic view which shows the relationship between the crank angle in this embodiment, and the opening / closing timing of the intake valve of each cylinder. It is a flowchart figure which shows the control by the controller of this embodiment. It is a characteristic view which shows the relationship between the crank angle in 2nd Embodiment, and the opening / closing timing of the intake valve of each cylinder. It is operation | movement explanatory drawing of the intake VEL of 2nd Embodiment.

Embodiments of an internal combustion engine control apparatus according to the present invention will be described below in detail with reference to the drawings.
[First Embodiment]
The first embodiment shows a so-called four-cycle four-cylinder internal combustion engine of gasoline specification applied to the intake valve side.

  First, the overall configuration of the internal combustion engine according to the present invention will be briefly described with reference to FIG. 1. The internal combustion engine can be applied to a so-called idling stop and can be applied to a so-called hybrid vehicle. .

  Piston 01 provided in a cylinder bore formed in the cylinder block SB so as to be slidable up and down, an intake port IP and an exhaust port EP formed in the cylinder head SH, and slidable in the cylinder head SH. Are provided with a pair of intake valves 4, 4 and exhaust valves 08, 08 for each cylinder that opens and closes the open ends of the intake and exhaust ports IP, EP.

  The piston 01 is connected to the crankshaft 02 via a connecting rod 03, and forms a combustion chamber 04 between the crown surface and the lower surface of the cylinder head SH.

  A throttle valve SV for controlling the intake air amount is provided in the upstream side of the intake manifold Ia of the intake pipe I connected to the intake port IP, and a fuel injection valve (not shown) is provided on the downstream side. It has been. In addition, a spark plug 05 is provided substantially at the center of the cylinder head SH.

  In the crankshaft 02, an outer ring gear 09 is always meshed with the gear of the pinion gear mechanism 06, and the pinion gear mechanism 06 is rotationally driven by an electric motor 07, whereby the crankshaft 02 is cranked. At the same time, the rotational position is controlled. That is, the electric motor 07 and the pinion gear mechanism 06 constitute a part of the crank position changing mechanism.

  The intake valves 4 and 4 are biased in the direction of closing the open ends of the intake ports IP via valve springs 5 and 5, respectively.

  In addition, as shown in FIGS. 1 and 2, this internal combustion engine is an intake valve VEL1 that is a variable lift mechanism that controls the valve lift and valve opening period (operating angle) of both intake valves 4, 4 as a variable valve operating device. And intake air VTC2 which is a lift phase variable mechanism for changing and controlling the opening / closing timing of the intake valves 4 and 4, that is, the center phase of the peak lift. In this embodiment, there is no exhaust VTC or the like on the exhaust valves 08 and 08 side, and the opening and closing timing is fixed.

  Since the intake air VEL1 has the same configuration as that described in, for example, Japanese Patent Application Laid-Open No. 2003-172112 previously filed by the applicant, as shown in FIG. 2 and FIG. A hollow drive shaft 6 rotatably supported by a bearing above the cylinder head SH, a drive cam 7 fixed to the outer peripheral surface of the drive shaft 6 by press-fitting or the like, and a swing on the outer peripheral surface of the drive shaft 6 Two swing cams 9, which are supported in a freely movable manner so as to open the intake valves 4, 4 by slidingly contacting the upper surfaces of the valve lifters 8, 8 disposed at the upper ends of the intake valves 4, 4. 9 and a drive mechanism that is interposed between the drive cam 7 and the swing cams 9 and 9 and converts the rotational force of the drive cam 7 into a swing motion and transmits it to the swing cams 9 and 9 as a swing force. And.

  The drive shaft 6 is transmitted with rotational force from the crankshaft by a timing chain (not shown) via a timing sprocket 33 provided at one end, and this rotational direction is clockwise (arrow direction) in FIG. Is set to

  The drive cam 7 has a substantially ring shape, and is fixed to the drive shaft 6 through a drive shaft insertion hole formed in the inner axial direction. The shaft center of the cam body extends from the shaft center of the drive shaft 6. Offset by a predetermined amount in the radial direction.

  As shown in FIGS. 2 and 3 and the like, both the swing cams 9 are integrally provided at both ends of a cylindrical camshaft 10, and the camshaft 10 is interposed via an inner peripheral surface. The drive shaft 6 is rotatably supported. Further, a cam surface 9a composed of a base circle surface, a ramp surface, and a lift surface is formed on the lower surface, and the base circle surface, the ramp surface, and the lift surface are arranged in accordance with the swing position of the swing cam 9. 8 is brought into contact with a predetermined position on the upper surface of 8.

  The transmission mechanism includes a rocker arm 11 disposed above the drive shaft 6, a link arm 12 linking the one end 11 a of the rocker arm 11 and the drive cam 7, the other end 11 b of the rocker arm 11, and a swing cam 9. And a link rod 13 that cooperates with each other.

  The rocker arm 11 has a cylindrical base portion at the center thereof rotatably supported by a control cam, which will be described later, via a support hole, and one end portion 11 a is rotatably connected to the link arm 12 by a pin 14. On the other hand, the other end 11 b is rotatably connected to one end 13 a of the link rod 13 via a pin 15.

  In the link arm 12, the cam body of the drive cam 7 is rotatably fitted in a fitting hole at the center position of the annular base end 12a, while the protruding end 12b protrudes from the base end 12a. Is connected to the rocker arm one end 11a by the pin 14.

  The other end portion 13 b of the link rod 13 is rotatably connected to the cam nose portion of the swing cam 9 via a pin 16.

  A control shaft 17 is rotatably supported by the same bearing member at a position above the drive shaft 6, and is slidably fitted into the support hole of the rocker arm 11 on the outer periphery of the control shaft 17. A control cam 18 serving as a swing fulcrum is fixed.

  The control shaft 17 is disposed in the longitudinal direction of the engine in parallel with the drive shaft 6 and is rotationally controlled by a drive mechanism 19. On the other hand, the control cam 18 has a cylindrical shape, and the axial center position is deviated from the axial center of the control shaft 17 by a predetermined amount.

  The drive mechanism 19 includes a drive motor 20 fixed to one end of a housing (not shown), and a ball screw transmission means 21 provided inside the housing for transmitting the rotational driving force of the drive motor 20 to the control shaft 17. It is composed of

  The electric motor 20 is constituted by a proportional DC motor and is driven by a control signal from a controller 22 which is a control mechanism for detecting an engine operating state.

  The ball screw transmission means 21 includes a ball screw shaft 23 disposed substantially coaxially with the drive shaft of the drive motor 20, a ball nut 24 that is a moving member screwed onto the outer periphery of the ball screw shaft 23, and the control It is mainly comprised from the linkage arm 25 connected with the one end part of the axis | shaft 17 along the diameter direction, and the link member 26 which links this linkage arm 25 and the said ball nut 24. As shown in FIG.

  The ball screw shaft 23 is rotationally driven by a drive motor 20 in which a ball circulation groove having a predetermined width is continuously formed spirally on the entire outer peripheral surface excluding both ends, and a drive shaft is connected to one end. It has become so.

  The ball nut 24 is formed in a substantially cylindrical shape, and a guide groove for continuously holding a plurality of balls is formed in a spiral manner in cooperation with the ball circulation groove on the inner peripheral surface. An axial moving force is applied to the ball nut 24 while converting the rotational motion of the ball screw shaft 23 into a linear motion via each ball. The ball nut 24 is urged toward the drive motor 20 (minimum lift side) by the spring force of the coil spring 30 as the second urging means. Therefore, when the engine is stopped, the ball nut 24 is moved to the minimum lift side along the axial direction of the ball screw shaft 23 by the spring force of the coil spring 30.

  The controller 22 detects the crank angle signal from the crank angle sensor that detects the current engine speed N (rpm), the engine speed signal, the accelerator opening sensor, the vehicle speed sensor, the gear position sensor, and the temperature of the engine body. The current engine operating state is detected from various information signals from an engine coolant temperature sensor or the like. Further, a detection signal from the drive shaft angle sensor 28 for detecting the rotation angle of the drive shaft 6 and a detection signal from the potentiometer 29 for detecting the rotation position of the control shaft 17 are inputted, and the sprocket 33 and the drive shaft 6 are input. And the valve lift amount and operating angle of each of the intake valves 4 and 4 are detected.

  Hereinafter, the basic operation of the intake VEL 1 will be described. For example, in a predetermined operation region such as a low rotation and low load, the ball screw shaft 23 is rotated by the rotational torque of the drive motor 20 that is driven to rotate in one direction by the control current from the controller 22. Is rotated in one direction, the ball nut 24 linearly moves in one direction (direction approaching the drive motor 20) while being assisted by the spring force of the coil spring 30, whereby the control shaft 17 and the link member 39 are moved. It rotates in one direction via the linkage arm 25.

  Therefore, as shown in FIGS. 3A and 3B (front view), the control cam 18 rotates around the axis of the control shaft 17 with the same radius, and the thick portion is upward from the drive shaft 6. Move away. As a result, the other end portion 11b of the rocker arm 11 and the pivot point of the link rod 13 move upward with respect to the drive shaft 6. Therefore, each swing cam 9 is connected to the cam nose portion side via the link rod 13. The whole is forcibly pulled up and rotated clockwise as shown in FIG.

  Therefore, when the drive cam 7 rotates and pushes up the one end portion 11a of the rocker arm 11 via the link arm 12, the lift amount is transmitted to the swing cam 9 and the valve lifter 16 via the link rod 13, thereby The intake valves 4 and 4 have a small lift (L1) as shown by the valve lift curve in FIG. 5, and the operation angle D1 (the valve opening rotation angle on the drive shaft is the valve opening crank). Half of the corner) becomes smaller.

  Here, there is a valve clearance between the swing cam 9 and the valve lifter 16, and the valve lift is smaller than the cam lift by the valve clearance. Further, the operating angle is from the opening timing to the closing timing of the valve lift in consideration of the valve clearance.

  In another operation state, when the drive motor 20 is rotated in the other direction by the control signal from the controller 22 and this rotational torque is transmitted to the ball screw shaft 23 and rotated, the ball nut 24 is rotated with the coil spring. It moves linearly in the opposite direction against the spring force of 30. Thereby, the control shaft 17 is rotationally driven by a predetermined amount in the clockwise direction in FIG.

  For this reason, the shaft center of the control cam 18 is held at a rotational angle position that is lower than the shaft center of the control shaft 17 by a predetermined amount, and the thick portion moves downward. Therefore, the entire rocker arm 11 moves counterclockwise from the position shown in FIG. 3, whereby each swing cam 9 is forcibly pushed down on the cam nose portion side via the link member 13, and the entire rocker arm 11 is counterclockwise. Turn slightly in the direction.

  Therefore, when the drive cam 7 rotates and pushes up the one end portion 11a of the rocker arm 11 via the link arm 12, the lift amount is transmitted to each swing cam 9 and the valve lifter 8 via the link rod 13, and the intake valve As shown in FIG. 5, the lift amount of 4 and 4 becomes the middle lift (L2), and the operating angle D2 also increases. As a result, the closing timing of the intake valves 4 and 4 is controlled in the vicinity of the bottom dead center on the retard side, so that the effective compression ratio is increased and combustion is improved. Moreover, the charging efficiency of fresh air is increased and the combustion torque is increased.

  Further, for example, in the case of shifting to the high rotation / high load region, the drive motor 20 is further rotated in the other direction by the control signal from the controller 22, and the control shaft 17 further rotates the control cam 18 in the clockwise direction, As shown in FIGS. 4A and 4B, the shaft center is rotated downward. Therefore, the entire rocker arm 11 further moves toward the drive shaft 6, and the other end portion 11 b presses the cam nose portion of the swing cam 9 downward via the link rod 13, thereby moving the entire swing cam 9. Is further rotated counterclockwise by a predetermined amount.

  Therefore, when the drive cam 7 rotates and pushes up the one end portion 11a of the rocker arm 11 via the link arm 12, the lift amount is transmitted to the swing cam 9 and the valve lifter 8 via the link rod 13. The valve lift amount continuously increases from L2 to L3 as shown in FIG. As a result, it is possible to increase the intake charging efficiency in the high rotation range, thereby improving the output.

  That is, the lift amount of the intake valves 4 and 4 changes continuously from the small lift L1 to the large lift L3 according to the operating state of the engine. Also changes continuously from a small lift D1 to a large lift D3.

  When the engine is stopped, the ball nut 24 is stably held in the small operating angle D1 and the small lift L1 region by the spring force of the coil spring 30 as described above. As a result, valve friction is reduced and good startability is obtained.

  Further, as shown in FIG. 5, there is a slight valve clearance Δ between the base circle surface and the valve lifter 8 during the swing of the swing cam 9, and the valve lift amount L slightly decreases by that amount. As a result, the operating angle D also decreases slightly. The lift amounts L1 to L3 and the operating angles D1 to D3 are expressed as values excluding the valve clearance Δ.

  Next, the intake VTC 2 will be described. As shown in FIGS. 6 to 8, this is a so-called vane type, which is rotationally driven by the crankshaft 02 and transmits the rotational driving force to the drive shaft 6, A vane member 32 fixed to the end of the drive shaft 6 and rotatably accommodated in the timing sprocket 33, and a hydraulic circuit for rotating the vane member 32 forward and backward by hydraulic pressure are provided.

  The timing sprocket 33 includes a housing 34 that rotatably accommodates the vane member 32, a disc-shaped front cover 35 that closes the front end opening of the housing 34, and a substantially disc that closes the rear end opening of the housing 34. The housing 34, the front cover 35, and the rear cover 36 are integrally fastened together by four small-diameter bolts 37 from the axial direction of the drive shaft 6.

  The housing 34 has a cylindrical shape in which both front and rear ends are formed, and shoes 34a that are four partition walls project inward at a position of about 90 ° in the circumferential direction of the inner peripheral surface.

  Each of the shoes 34a has a substantially trapezoidal cross section, and four bolt insertion holes 34b through which the shaft portions of the respective bolts 37 are inserted are formed at substantially the center position, and shafts are formed on the inner end surfaces. A U-shaped seal member 38 and a leaf spring (not shown) that presses the seal member 38 inward are fitted and held in a holding groove that is notched along the direction.

  The front cover 35 is formed in the shape of a disk plate, and a support hole 35a having a relatively large diameter is formed in the center, and at the position corresponding to each bolt insertion hole 34b of each shoe 34a on the outer periphery. Four bolt holes (not shown) are formed.

  The rear cover 36 is integrally provided with a gear portion 36a meshing with the timing chain on the rear end side, and a large-diameter bearing hole 36b is formed in the axial direction so as to penetrate therethrough.

  The vane member 32 includes an annular vane rotor 32a having a bolt insertion hole in the center, and four vanes 32b that are integrally provided at approximately 90 ° in the circumferential direction of the outer peripheral surface of the vane rotor 32a.

  In the vane rotor 32a, a small-diameter cylindrical portion on the front end side is rotatably supported by the support hole 35a of the front cover 35, while a small-diameter cylindrical portion on the rear end side is freely rotatable in the bearing hole 36b of the rear cover 36. It is supported.

  The vane member 32 is fixed to the front end portion of the drive shaft 6 from the axial direction by a fixing bolt 39 inserted through the bolt insertion hole of the vane rotor 32a from the axial direction.

  Of the vanes 32b, three of them are formed in a relatively long and narrow rectangular shape, and the three vanes 32b are set to have substantially the same width and length. The vane 32b is formed in a trapezoidal shape having a large width, and the width is set to be larger than the above three to balance the weight of the vane member 32 as a whole.

  Each vane 32b is disposed between the shoes 34a and has a U-shaped seal member 40 slidably contacting the inner peripheral surface of the housing 34 in an elongated holding groove formed in the axial direction of each outer surface. Leaf springs that press the seal member 40 toward the inner peripheral surface of the housing 34 are fitted and held, respectively. Further, two substantially circular concave grooves 32c are formed on one side surface of each vane 32b opposite to the rotation direction of the drive shaft 6, respectively.

  Further, four advance-side hydraulic chambers 41 and retard-side hydraulic chambers 42 are separated from both sides of each vane 32b and both sides of each shoe 34a, respectively.

  As shown in FIG. 6, the hydraulic circuit operates with respect to the first hydraulic passages 43 for supplying and discharging the hydraulic pressure of the hydraulic oil to the advance angle hydraulic chambers 41 and the retard angle hydraulic chambers 42. There are two systems of hydraulic passages, a second hydraulic passage 44 for supplying and discharging oil hydraulic pressure, and a supply passage 45 and a drain passage 46 are respectively provided in these hydraulic passages 43 and 44 for switching the passage. 47 is connected. The supply passage 45 is provided with a one-way oil pump 49 for pumping oil in the oil pan 48, while the downstream end of the drain passage 46 communicates with the oil pan 48.

  The first and second hydraulic passages 43 and 44 are formed in a cylindrical passage constituting portion 39, and one end of the passage constituting portion 39 extends from the small diameter cylindrical portion of the vane rotor 32a to the inside of the support hole 32d. The other end portion is connected to the electromagnetic switching valve 47.

  Further, between the outer peripheral surface of one end portion of the passage constituting portion 39 and the inner peripheral surface of the support hole 14d, three annular seal members 27 for separating and sealing one end side of each of the hydraulic passages 43 and 44 are fitted. It is fixed.

  The first hydraulic passage 43 is formed in an oil chamber 43a formed at the end of the support hole 32d on the drive shaft 6 side, and substantially radially inside the vane rotor 32a, and the oil chamber 43a and each advance side hydraulic chamber. And four branch passages 43 b communicating with 41.

  On the other hand, the second hydraulic passage 44 is stopped in one end portion of the passage constituting portion 39, and is bent into a substantially L shape inside the annular chamber 44a formed on the outer peripheral surface of the one end portion and the vane rotor 32. , A second oil passage 44b communicating with the annular chamber 44a and each retard side hydraulic chamber 42 is provided.

  The electromagnetic switching valve 47 is a four-port, three-position type, and an internal valve body is configured to relatively switch and control the hydraulic passages 43, 44, the supply passage 45, and the drain passage 46, and Switching operation is performed by a control signal from the controller 22.

  The controller 22 is common to the intake air VEL1 and detects the engine operating state and determines the relative rotational position of the timing sprocket 33 and the drive shaft 6 based on signals from the crank angle sensor 27 and the drive shaft angle sensor 28. Detected.

  Further, a locking mechanism is provided between the vane member 32 and the housing 34 as a fixing means for restricting the rotation of the vane member 32 relative to the housing 34 and releasing the restraint.

  This locking mechanism is provided between the one vane 32b having a large width and the rear cover 36, and has a sliding hole 50 formed along the axial direction of the drive shaft 6 inside the vane 32b. A lid-shaped cylindrical lock pin 51 slidably provided inside the sliding hole 50 and a cross-sectional cup-shaped engagement hole constituting portion 52 fixed in a fixing hole provided in the rear cover 36 are provided. The lock pin 51 is held by an engagement hole 52a in which the tapered tip 51a of the lock pin 51 engages and disengages, and a spring retainer 53 fixed to the bottom surface side of the sliding hole 50, thereby moving the lock pin 51 in the direction of the engagement hole 52a. And a spring member 54 urging the spring.

  The engagement hole 52a is directly supplied with the hydraulic pressure in the retard side hydraulic chamber 42 or the hydraulic pressure of the oil pump 49 through an oil hole (not shown).

  The lock pin 51 is located at the position where the vane member 32 is rotated to the most retarded angle side, and the tip 51a is engaged with the engagement hole 52a by the spring force of the spring member 54, so that the timing sprocket 31 and the drive shaft are engaged. Lock relative rotation with 6. Further, the lock pin 51 is moved backward by the hydraulic pressure supplied from the retard side hydraulic chamber 42 into the engagement hole 52a or the hydraulic pressure of the oil pump 49 so that the engagement with the engagement hole 52a is released. It has become.

  In addition, a pair of coil spring-like bias springs 55 for rotating and biasing the vane member 32 to the retard side are provided between one side surface of each vane 32b and the opposite surface of each shoe 34a facing the one side surface. 56 is arranged.

  The bias springs 55 and 56 appear to overlap in FIGS. 7 and 8, but are actually formed independently of each other and arranged in parallel to each other, and have their axial lengths ( The coil length) is set to be longer than the length between one side surface of the vane 32b and the opposing surface of the shoe 34a, and both are set to the same length.

  The bias springs 55 and 56 are arranged side by side with an inter-axis distance so that they do not contact each other even during maximum compression deformation, and one end of each of the bias springs 55 and 56 fits into the groove 32c of each shoe 34a. It is connected through.

  Hereinafter, the basic operation of the intake VTC 2 will be described. First, when the engine is stopped, output of the control current from the controller 22 to the electromagnetic switching valve 47 is stopped, and the valve body is mechanically operated by the bias springs 55 and 56 as shown in FIG. The supply passage 45 communicates with the retarded-side second hydraulic passage 44, and the drain passage 46 communicates with the first hydraulic passage 43. When the engine is stopped, the oil pressure of the oil pump 49 does not act and the supply oil pressure becomes zero.

  Accordingly, as shown in FIG. 7, the vane member 32 is rotated and biased to the most retarded angle side by the spring force of each of the bias springs 55 and 56, and one shoe 34a facing one end face of one wide vane 32b. Abuts on one side. At the same time, the tip 51a of the lock pin 51 of the lock mechanism is engaged in the engagement hole 52a, and the vane member 32 is stably held at the most retarded position. In other words, the intake VTC 2 is in the default position where the intake VTC 2 is mechanically stabilized at the most retarded position.

  Next, the operation of the intake VTC 2 will be briefly described. First, when the engine is started, that is, when the ignition switch is turned on to rotate the electric motor 07 to crank the crankshaft, a control signal is output from the controller 22 to the electromagnetic switching valve 47. However, since the discharge hydraulic pressure of the oil pump 49 has not yet sufficiently increased immediately after the start, the vane member 32 is moved to the most retarded angle side by the lock mechanism and the spring force of the bias springs 55 and 56. Is retained.

  At this time, the electromagnetic switching valve 47 causes the supply passage 45 and the second hydraulic passage 44 to communicate with each other and the drain passage 46 and the first hydraulic passage 43 communicate with each other according to the control signal output from the controller 22. As the hydraulic pressure pumped from the oil pump 49 increases, the hydraulic pressure is supplied to the retarded hydraulic chamber 42 through the second hydraulic passage 44, while the advanced hydraulic chamber 41 receives the hydraulic pressure in the same manner as when the engine is stopped. The oil pressure is released from the drain passage 46 into the oil pan 48 without being supplied, and the low pressure state is maintained.

  Here, after the hydraulic pressure has increased, the vane position can be freely controlled by the electromagnetic switching valve 47. That is, as the hydraulic pressure in the retard side hydraulic chamber 42 increases, the hydraulic pressure in the engagement hole 52a of the lock mechanism also increases, the lock pin 51 moves backward, and the tip 51a comes out of the engagement hole 52a and Since relative rotation with the vane member 32 is permitted, free vane position control is possible.

  For example, in the idling state after the warm-up is completed, the electromagnetic switching valve 47 communicates the supply passage 45 and the second hydraulic passage 44 and communicates the drain passage 46 and the first hydraulic passage 43. Therefore, the vane member 32 maintains the position of FIG. 7 together with the spring force of each bias spring 55, 56 as the internal pressure of the retard side hydraulic chamber 42 increases, and the drive shaft 6 retards the timing sprocket 33. Rotate to the side.

  Thereafter, for example, when shifting to a predetermined low-rotation load range, the electromagnetic switching valve 47 is actuated by a control signal from the controller 39 to connect the supply passage 45 and the first hydraulic passage 43, while the drain passage 46 and the second passage. The hydraulic passage 44 is communicated.

  Accordingly, the hydraulic pressure in the retard side hydraulic chamber 42 is returned to the oil pan 48 from the drain passage 46 through the second hydraulic passage 44 this time, while the inside of the retard side hydraulic chamber 42 becomes low pressure, while the advance angle is increased. The hydraulic pressure is supplied into the side hydraulic chamber 41 and becomes high pressure.

  Therefore, the vane member 32 rotates in the clockwise direction in the drawing against the spring force of the bias springs 55 and 56 due to the high pressure in the advance side hydraulic chamber 41 and relatively rotates to the position shown in FIG. The relative rotational phase of the drive shaft 6 with respect to the timing sprocket 33 is converted to the advance side. Further, by setting the position of the electromagnetic switching valve 47 to the neutral position in the middle of the relative rotation, it can be held at an arbitrary relative rotation phase.

  Further, when the engine shifts from the low rotation range to the normal medium rotation range and further to the high rotation range, the vane member 32 is controlled by performing the same control as the idling operation state after the warm-up is completed. The hydraulic pressure supplied to the advance-side hydraulic chamber 41 decreases, and conversely, the hydraulic pressure in the retard-side hydraulic chamber 42 increases, and due to the combined force with the spring force of each bias spring 55, 56, the timing sprocket 33 and The relative rotational phase of the drive shaft 6 is converted to the retard side (see FIG. 7).

  Next, the control by the controller 22 will be described. Prior to that, the relationship between the crank angle of the crankshaft 02 shown in FIG. 9 and the opening / closing timing of the intake valves 4 and 4 in each cylinder will be described. Note that the firing order of each cylinder is # 1 cylinder → # 3 cylinder → # 4 cylinder → # 2 cylinder. In this figure, it is assumed that the internal combustion engine is stopped, and the opening / closing timing of the intake valves 4 and 4 by the intake VTC 2 is stable on the default most retarded side described above.

  Now, the control shaft 17 of the intake VEL1 controls the intake valves 4 and 4 to the position of the maximum operating angle D3. For example, the crank angle of the crankshaft 02 is the top dead center position of the compression stroke in the # 1 cylinder. Suppose there was. At this time, the intake valves 4 and 4 of the # 1 cylinder are of course not opened, but the intake valves 4 and 4 (operating angle D3) of the # 3 cylinder of the next cycle are opened, and # 4 cylinder intake valves 4, 4 (operating angle D3) are also open. That is, the intake valves 4 and 4 of the two cylinders are simultaneously opened.

  Next, when it is examined whether or not all the intake valves 4 and 4 are closed at different crank angles, such a state does not exist. This corresponds to the existence of a region where lift curves overlap between two cylinders.

  That is, at any crank angle, at least the intake valves 4 and 4 of one cylinder are opened, and the reaction force of the valve springs 5 and 5 acts on the control cam 18. Here, if the control shaft 17 is to be rotated by the drive motor 20 via the ball screw transmission means 21 while the internal combustion engine is stopped, the reaction force and a large static friction coefficient will cause the control shaft 17 to rotate quickly. could not.

  Next, even if the operating angle D of the intake valves 4 and 4 is not the maximum D3 but the intermediate operating angle D2, there is still a region where the lift curves overlap in two cylinders. Even if it exists, since the intake valves 4 and 4 of at least one cylinder are opened, the spring reaction force of the control cam 18 and the valve springs 5 and 5 acts.

  If the control shaft 17 is rotated by the drive motor 20 or the like while the engine is stopped, the reaction force and a large static friction coefficient cannot be rotated as quickly as the operating angle D3.

  Next, the minimum operating angle D1 of the intake valves 4 and 4 is verified. When the crank angle is at the star A point in FIG. 9, the # 3 cylinder is still open, so the control cam 18 The spring reaction force of the valve springs 5 and 5 is acting. Therefore, if the control shaft 17 is to be rotated by the drive motor 20 or the like while the engine is stopped, even the operating angle D1 cannot be quickly rotated due to the reaction force and the static friction coefficient.

  However, at this operating angle D1, there appears a region where the lift curves do not overlap between the two cylinders. In this case, since none of the cylinders are open in the crank angle range α1, the spring reaction force of the valve springs 5 and 5 hardly acts when the crank angle is at, for example, the star B point in α1. Therefore, the control cam 18 rotates smoothly. When the control cam 18 starts to rotate, the friction coefficient of the sliding portion of the control cam 18 changes from a large static friction coefficient to a small dynamic friction coefficient. Therefore, the operating angle D1 can be smoothly changed from D1 to D2, and further to D3.

  Here, when the operating angle becomes D2 and D3, the valve starts to open even at the star point B, but at that time, it has already changed to a small dynamic friction coefficient, and the control shaft 17 has already started to rotate, A good conversion response can be maintained by the inertia.

  That is, in the present embodiment, when the internal combustion engine is stopped, the rotation stop position of the crankshaft 02 is adjusted within the range of α1 (or α2 to α4) by the electric motor 07 described above. At this time, the operating angle is controlled to the minimum operating angle.

  Next, when the engine is restarted, a control signal is output to the drive motor 20 of the intake air VEL1 in order to convert the control cam 18 to a desired operating angle before cranking is started. As a result, the conversion is started before cranking, and the conversion time to the target operating angle is shortened in combination with the effect of improving the conversion response (dynamic friction coefficient) and the rotational inertia of the control shaft 17. .

  By the way, the required operating angle of the intake valves 4 and 4 at the time of starting the engine varies depending on the engine temperature or the like. For example, when the engine temperature is extremely low, the intake valves 4 and 4 are used to ensure good combustion. Needs to be close to the bottom dead center of the piston, and the intermediate operating angle D2 is selected as the target operating angle.

  Conversely, when the engine temperature is high, the large operating angle D3 is selected as the target operating angle in order to suppress pre-ignition and starting vibration. According to this, since the IVC is greatly retarded with respect to the bottom dead center, the fresh air once sucked is discharged, thereby reducing the effective compression ratio and suppressing pre-ignition and starting vibration by the decompression action. It can be done.

  In the case of a general restart at which the engine temperature is neither extremely low nor high, the minimum operating angle D1 is selected. According to this, since the valve lift amount and the operating angle of the intake valves 4 and 4 are small, the valve drive friction is reduced and a smooth increase in the rotation of the engine can be obtained, so that smooth and quick startability can be realized. .

  Here, the minimum operating angle D1 of the intake valves 4 and 4 reduces the effective compression ratio and has a decompression effect. However, the minimum operation angle D1 is slightly increased by the intake agitation effect due to the delay of the valve opening timing (IVO) of the intake valves 4 and 4. Since there is a possibility of facilitating ignition, the maximum operating angle D3 is somewhat advantageous for high oil temperature.

  As described above, when the engine is stopped, the crank angle is set in advance by the crank position control mechanism such as the electric motor 07 during the all-cylinder closing period in which the intake valves 4 and 4 of all the cylinders are in the non-lifted state. Since the control signal is output to the intake air VEL1 prior to cranking rotation toward the target operating angle value corresponding to the engine state such as engine oil temperature at the time of starting, the conversion time to the target operating angle can be shortened.

  Hereinafter, a specific control flow by the controller 20 will be described with reference to FIG.

  First, in step 1, it is determined whether or not the current engine state is an engine stop condition, that is, whether or not the engine is turned off by an ignition switch. Or if it is a hybrid vehicle using an idling stop system, it will be judged automatically whether it is the conditions which an engine stops.

  If NO is determined, the process returns without processing. If YES, that is, if it is determined that the stop condition is satisfied, the process proceeds to step 2, where the intake valves 4, 4 are connected to the drive motor 20 of the intake VEL1. A switching control signal is output so that the minimum operating angle D1 is obtained.

  In step 3, a control signal is output to the electric motor 07 of the crank angle changing mechanism, and the crankshaft 02 is controlled so that the intake valves 4 and 4 of all the cylinders are closed (period), for example, α1. To do.

  In step 4, it is determined whether or not the intake valve VEL1 actually has the operating angle of the intake valves 4 and 4 set to the operating angle D1, and whether or not the crank angle when the engine is stopped is within the range α1. Here, if it is determined that it is not yet, the process returns to Step 2, but if it is determined that it is (for example, the star B in FIG. 9), the process proceeds to Step 5.

  In this step 5, an engine stop signal is outputted, and in step 6, the engine rotation is actually stopped.

  Until the next engine restart, the internal combustion engine continues to be stopped. However, as described above, in the intake VEL1, the minimum operating angle D1 of the intake valves 4 and 4 is stabilized via the spring force of the coil spring 30. In the position (default), this operating angle D1 is maintained. Further, as described above, the intake VTC 2 is in a position (default) where the opening / closing timing of the intake valves 4 and 4 is stabilized at the most retarded angle by the spring force of the bias springs 55 and 56, and the most retarded angle is maintained. Further, the crankshaft 02 is maintained at the star point B in FIG.

  Next, in step 7, it is determined whether or not the engine restart conditions, that is, restart conditions such as a reacceleration request scene of a hybrid vehicle, for example. If yes, go to Step 8.

  In step 8, for example, the engine temperature T, which is one of the current engine states, is read from a water temperature sensor or the like, and the process proceeds to step 9. In step 9, is the engine temperature T higher than a predetermined first temperature T1? Judge whether or not.

  If it is determined in step 9 that T ≦ T1, that is, if it is determined that the engine is in a cold state, the process proceeds to step 10, where the intermediate operating angle D2 is set to the target operating angle value for starting by the intake air VEL1. Set to Dt, then go to Step 14. At this time, the intake valves 4 and 4 are stable at the most retarded angle by the intake VTC2, and the IVC is also near the bottom dead center. That is, for example, the IVC of the # 1 cylinder is near the bottom dead center of the # 1 cylinder (= compression top dead center of the # 2 cylinder). Therefore, since the effective compression ratio can be set high, combustion during cold operation can be improved.

  If it is determined in step 9 that T> T1, the process proceeds to step 11, where the engine temperature is further compared and determined to determine whether the current engine temperature T is higher than the second temperature T2. . If it is determined that T ≧ T2, that is, if it is determined that the temperature is high, the process proceeds to step 12.

  In step 12, the large operating angle D3 of the intake valves 4 and 4 is set to the starting target operating angle value. Here, the intake VTC 2 is stable at the most retarded angle, and the IVC of the intake valves 4 and 4 of the # 1 cylinder is larger than the bottom dead center of the # 1 cylinder (= compression top dead center of the # 2 cylinder). is doing. Further, the IVCs of the intake valves 4 and 4 of the other cylinders are similarly retarded from the bottom dead center of the cylinders.

  Therefore, the effective compression ratio can be lowered, and therefore the occurrence of pre-ignition can be suppressed by decompression. In addition, the starting vibration may increase due to the high oil temperature (low viscosity), but the starting vibration can be sufficiently suppressed by the decompression.

  When it is determined NO in step 11, that is, when the temperature is lower than T2 and the relationship of T1 <T <T2, it is determined that the general starting condition is satisfied, and in step 13, the intake valve 4 , 4 is set to the starting target operating angle value Dt.

  Thus, after the start target operating angle value Dt is determined (steps 10, 12, 13), prior to cranking rotation, the operating angle of the intake valves 4, 4 is switched to the start target operating angle value Dt in step 14. A signal is output to the drive motor 20 of the intake air VEL1.

  Here, the above-mentioned star point B is a crank angle at which the intake valves 4 and 4 are not opened in all the cylinders, so that the spring reaction force of the valve springs 5 and 5 hardly acts on the control cam 18. For this reason, the conversion of the control cam 18 is started smoothly via the control shaft 17 by the rotational force of the drive motor 20 of the intake air VEL1. The sliding portion between the control cam 18 and the rocker arm 11 changes from the static friction coefficient (large) region to the dynamic friction coefficient (small) region. For this reason, even if the operation becomes smoother, the operation angle D2 starts to open in some cylinders, and further converted to the maximum operation angle D3, this smooth conversion operation continues. It is. Further, when the control shaft 17 starts to rotate, the smooth conversion operation is continued by its inertia.

  Subsequently, in step 15, cranking of the crankshaft 02 by the electric motor 07 is started. This cranking start may be performed after confirming that the target operating angle value Dt has been reached, or may be in the middle of conversion to the target operating angle value Dt, or in a state where conversion has not been confirmed.

  In the former case, since the target operating angle value Dt has been reached from the beginning of cranking, the desired start performance effect can be obtained. Further, when cranking is started by the electric motor 07, the peak current timing of the drive motor 20 of the intake air VEL1 has passed, so that a sufficient battery voltage can be supplied to the electric motor 07, and surplus cranking can be realized. .

  On the other hand, in the latter case, cranking can be started quickly without confirming the completion of conversion. Therefore, cranking is started immediately after it is determined in step 7 that the engine restart condition is satisfied, and the engine is immediately started. This is advantageous in sudden acceleration of the vehicle. This flowchart is described assuming the latter example. In step 15, cranking is started without confirming the end of conversion to the target operating angle value Dt.

  By the way, considering the initial compression in the initial stage of cranking, the # 1 cylinder is at the star B point (slightly before the compression top dead center). Therefore, atmospheric pressure has flowed into the cylinder after the engine has stopped from the gap of the piston, and compression is performed from point B to compression top dead center with atmospheric pressure as the initial condition.

  However, since the point B is close to the compression top dead center in the first place, the piston stroke to the compression top dead center is short and the compression is negligible, so the cranking rotation rises smoothly. Startability can be improved. Here, if point B is after compression top dead center, the initial compression itself does not occur, and the cranking rotation can be raised smoothly.

  Therefore, not only pre-ignition and start-up vibration can be avoided, but speedy start can be realized.

  Then, in step 16, it is determined whether or not the target operating angle value Dt has been reached. If yes, after the confirmation, complete explosion control such as fuel injection and ignition is performed in step 17, and a reliable and speedy start is performed. Is completed. If it is determined in step 16 that the target operating angle value Dt has not been reached, the process returns to step 14 to output the control signal to the target operating angle value Dt again, and in step 15 the cranking is started and continued. .

In the present embodiment, the engine temperature is targeted as the engine state, but it is also possible to include the vehicle speed as a target. With these, the required state of the period is detected in step 8, and the target operating angle is accordingly detected. Can also be set.
[Second Embodiment]
11 and 12 show a second embodiment, which is applied to an in-line two-cylinder internal combustion engine, and the basic configuration of the intake air VEL1 and the intake air VTC2 is the same as that of the first embodiment.

  As shown in FIG. 11, when the engine is stopped, the operating angle of the intake valves 4, 4 by the intake VEL1 is the default minimum operating angle D1 '.

  Here, the crank angle region where the intake valves 4 and 4 of both cylinders are closed in the vicinity of the compression top dead center of the # 1 cylinder is α1 ′, and from the closing of the intake valves 4 and 4 of the # 1 cylinder. This is the interval until the intake valves 4 and 4 of the # 2 cylinder are opened. These intervals are sufficiently expanded as compared with the α1 of the four cylinders shown in FIG. This is because when the cylinder interval is 4 cylinders, the crank angle is only 180 °, but when the cylinder interval is 2 cylinders, the angle is increased to 360 °.

  Therefore, since the control target range of the crank position changing mechanism is expanded from α1 to α1 ', the control accuracy of the crank angle position changing mechanism can be lowered.

  In addition, the minimum operating angle D1 ', that is, the period between the opening and closing of the intake valves 4 and 4 (operating angle) is assumed to be shorter than that of the first embodiment.

  The minimum operating angle D1 'is D1 shown in FIG. 5 in the first embodiment, but is D1' in the second embodiment. In the first embodiment, it means a section D1 between the point at which the valve is actually opened and the point at which the valve is closed, but in the second embodiment, the down ramp point (the minute lift ΔL) shown in FIG. This refers to the section D1 ′ between the minute lifts ΔL).

  Therefore, since the relationship of D1 ′ <D1 is satisfied, α1 ′ can be further enlarged accordingly. As a result, the demand for control accuracy of the crank position changing mechanism is eased. Alternatively, it can be said that the controllability that satisfies the target crank stop position range is improved.

  Here, the link posture at the time of the ramp lift is shown in FIG. 12, but even when the load FS is applied from the bulb springs 5 and 5, the offset ΔT of the load point with respect to the swing axis is sufficiently small. Accordingly, the moment ΔM acting on the swing cams 9 and 9 is sufficiently small, so that the load acting on the control cam 18 via the link rod 13 and the like is sufficiently small. Therefore, even when the engine is stopped, the control shaft 17 can be smoothly rotated, and switching control close to a complete zero lift can be performed.

  Incidentally, for example, in the case of the L1 lift exceeding the ramp lift, as shown in FIG. 3B, the offset of the load point is large (T), and therefore the moment acting on the swing cams 9 and 9 is large (M). Since the load acting on the control cam 18 becomes significantly large, it is difficult to smoothly rotate (convert) the control shaft 17 when the engine is stopped.

  Furthermore, if the rotational position of the crankshaft 02 is controlled by the electric motor 07 or the like with the point B ′, which is the intermediate point of the control target range α1 ′ of the crank position changing mechanism, as a target point, there will be variations in the friction of the piston 01. Can be put in the target range α1 ′ with high accuracy.

  The present invention is not limited to the configuration of each of the embodiments described above. For example, the number of cylinders is not limited, but the intake valves 4 and 4 of all the cylinders are closed as the number of cylinders decreases to 3 and 2 cylinders. Since the crank angle range is expanded, control becomes easier.

  Further, the crank position changing mechanism is configured by the electric motor 07 and the pinion gear mechanism 06, but is not limited to this, and may be provided in different forms. For example, an electric motor can be directly connected to the rear end of the crankshaft.

  Further, the vehicle can be applied to a general vehicle that also stops when the engine is stopped, and can also be applied to a hybrid vehicle that can be self-propelled by a motor even when the engine is stopped.

  Further, as the variable valve mechanism, the one that changes the attitude of the swing cam 9 by changing the angle of the control shaft 17 has been shown. However, by changing the position of the control shaft 17 in the axial direction, the swing cam 9 It can also be applied to other structures such as those that change the posture of the camera.

The technical ideas of the invention other than the claims ascertained from the embodiment will be described below.
[Claim a] The control apparatus for an internal combustion engine according to claim 1,
When the internal combustion engine is stopped, the control device for the internal combustion engine is controlled to the small operating angle side in the operating angle movable range of the intake valve by the variable valve operating device.
[Claim b] An internal combustion engine control apparatus according to claim 2,
The variable valve operating device can be controlled until the operating angle of the intake valve reaches a minimum operating angle that does not become zero, and when the internal combustion engine is stopped, the variable valve operating device is controlled so that the operating angle becomes the minimum operating angle. A control device for an internal combustion engine.

According to the present invention, since the initial operating angle (lift) that is not zero is converted to the target operating angle, the conversion width is reduced accordingly, so that the conversion response time can be shortened and the structure of the variable valve operating device is simplified. it can.
[Claim c] A control apparatus for an internal combustion engine according to claim 1,
When the internal combustion engine is restarted when the engine temperature is equal to or lower than a predetermined first temperature, the closing timing of the intake valve is controlled to be near the bottom dead center of the piston in the intake stroke. Control device.
[Claim d] A control apparatus for an internal combustion engine according to claim c,
When the engine temperature at the time of restart of the internal combustion engine is equal to or higher than the second temperature exceeding the first temperature, the closing timing of the intake valve is controlled to deviate from the vicinity of the piston bottom dead center of the intake stroke. A control device for an internal combustion engine.

According to the present invention, combustion can be improved when the engine temperature at the time of restart is low, and pre-ignition and start-up vibration when the engine temperature at the time of restart is high can be avoided.
(Claim e) A control device for an internal combustion engine according to claim d,
When the engine temperature at the time of restarting the internal combustion engine is equal to or lower than the first temperature, the operating angle of the intake valve is larger than when the engine temperature exceeds the first temperature and is lower than the second temperature, and A control apparatus for an internal combustion engine, characterized in that control is performed to be smaller than when the second temperature is exceeded.
[Claim f] A control device for an internal combustion engine according to claim 1, comprising:
When the internal combustion engine is stopped, the variable valve device controls the intake valve of all the cylinders to be in a non-lift state, and then the crank position control mechanism controls the crankshaft stop position. Is controlled so as to be the valve closing period.
[Claim g] A control apparatus for an internal combustion engine according to claim 1, wherein:
The control apparatus for an internal combustion engine, wherein the internal combustion engine is mounted on a vehicle having an idling stop function.
(Claim h) A control apparatus for an internal combustion engine according to claim 1, wherein
The control apparatus for an internal combustion engine, wherein the crank position control mechanism controls a position of a crankshaft by controlling an electric motor.
[Claim i] A control apparatus for an internal combustion engine according to claim 1, comprising:
The control apparatus for an internal combustion engine, wherein the internal combustion engine is mounted on a hybrid vehicle capable of running only with an electric motor while the internal combustion engine is stopped.
[Claim j] A control device for an internal combustion engine according to claim g,
The control apparatus for an internal combustion engine, wherein the crank position control mechanism controls a position of a crankshaft by controlling an electric motor used for traveling of a vehicle.
[Claim k] A control apparatus for an internal combustion engine according to claim 1, wherein:
The control apparatus for an internal combustion engine, wherein the crank position control mechanism controls a position of a crankshaft by controlling an alternator used for power generation.
[Claim 1] A control device for an internal combustion engine according to claim 1, comprising:
When the internal combustion engine is restarted, when the ignition switch is operated and the power is turned on, the variable valve operating device is controlled to start lift characteristics of the intake valve according to the engine operating state. Control device.
[Claim m] A control apparatus for an internal combustion engine according to claim l,
If forced cranking is performed by an ignition switch before the variable valve device is controlled to the start lift characteristic of the intake valve according to the engine operating state, the variable valve device is inhaled during cranking rotation. A control device for an internal combustion engine, wherein the control device is controlled toward a starting lift characteristic of the valve.
[Claim n] A control apparatus for an internal combustion engine according to claim l,
A control device for an internal combustion engine, wherein cranking is started after the variable valve operating device is controlled to a start lift characteristic of an intake valve.
(Claim o) A control device for an internal combustion engine according to claim n,
The control apparatus for an internal combustion engine, wherein the control shaft is directly driven by power of an electric motor.
[Claim p] A control device for an internal combustion engine according to claim 1,
A control apparatus for an internal combustion engine, wherein a stop position range of a crankshaft when the internal combustion engine stops is a predetermined range including a piston top dead center of a compression stroke with respect to a predetermined cylinder.
[Claim q] A control apparatus for an internal combustion engine according to claim 1, comprising:
A control apparatus for an internal combustion engine, characterized in that a stop position range of a crankshaft when the internal combustion engine stops is a range including immediately before a piston top dead center of a compression stroke with respect to a predetermined cylinder.

  According to the inventions described in claims p and q, during the initial stop, the predetermined cylinder internal pressure near the top dead center of the piston during the compression stroke or immediately before the pressure decreases to the atmospheric pressure. Can be shortened or eliminated, so the compression at the time of the first compression can be realized, and the starting vibration and pre-ignition can be further reduced.

02 ... Crankshaft 06 ... Pinion gear mechanism (Crank position changing mechanism)
07 ... Electric motor (Crank position changing mechanism)
08 ... Exhaust valve 1 ... Intake VEL (lift variable mechanism / variable valve operating device)
2 ... Intake VTC (lift phase variable mechanism / variable valve operating device)
DESCRIPTION OF SYMBOLS 4 ... Intake valve 5 ... Valve spring 6 ... Drive shaft 7 ... Drive cam 9 ... Swing cam 11 ... Rocker arm 17 ... Control shaft 18 ... Control cam 20 ... Drive motor 22 ... Controller 30 ... Coil spring (biasing member)
32 ... vane members 55, 56 ... biasing spring (biasing member)

Claims (3)

An internal combustion engine that controls a crank position changing mechanism capable of changing a rotation stop position of a crankshaft of the internal combustion engine and a variable valve operating device that changes at least operating angles of intake valves of a plurality of cylinders by changing the position of a control shaft. A control device of
When the internal combustion engine is stopped, the variable valve device controls the intake valve of all the cylinders to be in a non-lift state, and a crankshaft stop position is controlled by the crank position changing mechanism. To control the valve closing period,
A control apparatus for an internal combustion engine, wherein, when the internal combustion engine is restarted, the variable valve apparatus is controlled to have a start lift characteristic of an intake valve in accordance with an engine state prior to cranking rotation. An internal combustion engine for controlling a crank position changing mechanism capable of changing a rotation stop position of a crankshaft of an internal combustion engine, and a variable valve operating device for changing an operation angle of intake valves of a plurality of cylinders by changing a position of a control shaft A control device,
When the internal combustion engine is stopped, the variable valve device and the crank position changing mechanism are controlled so that almost no spring load from the valve spring acts on the control shaft,
A control apparatus for an internal combustion engine, wherein when the internal combustion engine is restarted, the position of the control shaft is controlled to be a position suitable for starting prior to cranking rotation. Crank position changing mechanism capable of changing rotation stop position of crankshaft of internal combustion engine, variable valve operating device for changing operating angle and lift amount of intake valves of plural cylinders, crank position changing mechanism and variable valve operating device An internal combustion engine control system comprising a control device for controlling
When the internal combustion engine is stopped, the variable valve device controls the intake valve of all the cylinders to be in a non-lift state, and a crankshaft rotation stop position is set by the crank position changing mechanism. Control to be the valve closing period,
A control system for an internal combustion engine, wherein when the internal combustion engine is restarted, the variable valve operating device is controlled toward a start lift characteristic of an intake valve in accordance with an engine state prior to cranking rotation.

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