JP2012036864A - Variably operated valve apparatus for internal combustion engine, start system for internal combustion engine, and start control apparatus for internal combustion engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a variably operated valve apparatus that achieves stabilization of top dead center compression by a piston, thereby capable of improving startability.SOLUTION: In a step 1 when it is confirmed that an engine stop condition is satisfied, in a step 2, a conversion signal into most delayed angle is output by an intake VTC, and a conversion signal into maximum operating angle is output by an intake VEL. In a step 3, actual positions of closing timing and opening timing of an intake valve by the intake VEL and the intake VTC are detected, and in a step 4, it is discriminated whether or not an absolute value of actual closing timing IC-IC3 is less than a predetermined minute angle ΔIC. If less than the predetermined minute angle ΔIC, in a step 5, stop position control of a crank angle Z (piston) is performed (Z0±α) by a drive motor, and controlled to a Z0 on an advance angle side than closing timing IC3. This achieves stabilization of compression by the piston at the engine starting.

Description

  The present invention relates to a variable valve operating apparatus for an internal combustion engine that can improve startability by controlling at least a relationship between a closing timing of an intake valve and a stop position of a piston.

  As a conventional variable valve operating apparatus for an internal combustion engine, a technique described in Patent Document 1 below is known.

  When the engine start condition is satisfied, the variable valve device reduces the compression at the top dead center of the cylinder in the piston compression stroke by retarding the closing timing of the intake valve by the valve timing control device (decompression). ), The cranking rotation is accelerated and good self-starting is realized.

JP 2006-125276 A (see FIG. 3 and the like)

  However, in the conventional variable valve operating apparatus described in Patent Document 1, the stop position of the piston (crankshaft) of the cylinder in the compression stroke when the engine is stopped is increased by the piston being pushed down by the compression in the cylinder. In many cases, the position is approximately halfway between the dead center and the bottom dead center. In this state, the atmospheric pressure enters the cylinder from the gap between the piston and the cylinder bore until the next engine restart. For this reason, when cranking at the time of restarting is started, the compression stroke piston rises from the crank position when the engine is stopped, and the compression becomes large.

  The crank position (piston position) when the engine is stopped is likely to vary. If this crank position is earlier than the intake valve closing timing, the intake valve closing timing is set to atmospheric pressure, and the compression increases as the piston rises thereafter. If the intake valve is later than the closing timing of the intake valve, the crank position is set to the atmospheric pressure, and the compression increases as the piston rises thereafter.

  In other words, both the case where the stop position of the crank varies and the compression is determined by the closing timing of the intake valve and the case where the compression is determined by the stop position of the crank occur. There is a possibility that stable engine startability may not be obtained due to variations in start cranking characteristics.

  According to the first aspect of the present invention, in particular, at least the closing timing of the intake valve can be variably controlled, and when the engine is stopped, the closing timing of the intake valve is set to be more delayed than the position most delayed in the stop position control range of the piston. Further, the present invention is characterized in that it is configured to be mechanically stabilized at a retarded position.

  The invention according to claim 2 is configured such that at least the closing timing of the intake valve can be variably controlled, and when the engine is stopped, the closing timing of the intake valve is stopped at a fixed position that is retarded from the bottom dead center. And the crank position control means for controlling the stop position of the piston to a position advanced from the closing timing of the intake valve when the engine is stopped.

  According to a third aspect of the present invention, when the engine is stopped, the variable valve device controls the closing timing of the intake valve to be a fixed position on the retard side of the bottom dead center, and the crank position control means controls the piston. The stop position is controlled to a position advanced from the closing timing of the intake valve.

  According to the present invention, it is possible to stabilize the top dead center compression by the piston, thereby stabilizing the cranking characteristic and improving the startability.

1 is a schematic view of an internal combustion engine and a variable valve operating apparatus according to a first embodiment of the present invention. It is a perspective view which shows the intake VEL and intake VTC which are provided to this embodiment. It is a partial cross section figure which shows the maximum lift control state by the drive mechanism with which this embodiment is provided. It is a partial cross section figure which shows the minimum lift control state by the drive mechanism with which this embodiment is provided. A and B are operation explanatory views at the time of the minimum control by the intake air VEL of the present embodiment. A and B are operation explanatory views at the time of maximum control by the intake air VEL of the present embodiment. It is a characteristic view of the valve lift and opening / closing timing of the intake valve by the intake VEL and the intake VTC. It is a front view of the state which removed the front cover of the intake VTC provided for this Embodiment, and has shown the maximum retardation control state It is a longitudinal cross-sectional view which shows the whole structure of the same intake VTC. It is a characteristic view which shows the opening-and-closing timing characteristic of the intake valve and exhaust valve in a prior art, and the crank angle (piston stroke position) between cylinders. It is a characteristic view which shows the opening / closing timing characteristic of the intake valve and the exhaust valve in this embodiment, and the crank angle (piston stroke position) between cylinders. It is a time chart figure of the combustion cycle between each cylinder in this embodiment. It is a control flowchart figure by the controller at the time of the engine stop in this embodiment. It is a control flowchart figure by the controller at the time of engine starting in this embodiment. It is a characteristic view which shows the opening / closing timing characteristic of the intake valve and the exhaust valve in 2nd Embodiment, and the crank angle (piston stroke position) between cylinders. It is a time chart figure of the combustion cycle between each cylinder in this embodiment in this embodiment. It is a control flowchart figure by the controller at the time of engine starting in this embodiment, and after starting.

  Embodiments of a variable valve operating apparatus for an internal combustion engine according to the present invention will be described below in detail with reference to the drawings. This embodiment shows what is applied to a so-called gasoline specification four-cycle four-cylinder internal combustion engine.

  First, the overall configuration of the internal combustion engine according to the present invention will be described with reference to FIG. 1. A piston 01 provided in a cylinder bore formed in a cylinder block SB and slidable up and down, and an interior of the cylinder head SH. And a pair of intake valves 4 per cylinder that are slidably provided on the cylinder head SH and open and close the open ends of the intake and exhaust ports IP and EP, respectively. 4 and exhaust valves 5 and 5.

  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 butterfly 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. An injection valve is provided to inject fuel into the intake port IP. In addition, an ignition plug 05 is provided substantially at the center of the cylinder head SH.

  The crankshaft 02 is rotationally driven by an electric drive motor 07 via a pinion gear mechanism 06 when the engine is started. The drive motor 07 is configured not only to perform the start cranking but also as a crank position control means for controlling the crank angle (crank position) when the engine is stopped and the sliding position of the piston 01.

  The intake valves 4 and 4 are variably controlled in their valve lift amount, operating angle and lift phase (opening / closing timing) by a variable valve operating device.

  That is, as shown in FIGS. 1 and 2, the variable valve operating apparatus includes an intake VEL1 that is a second variable mechanism that controls the valve lift and the operating angle (valve opening period) of both intake valves 4 and 4, and the intake air. The intake VTC 2 is a first variable mechanism that controls the opening / closing timing of the valves 4, 4, and the exhaust VTC 3 is a third variable mechanism that controls the opening / closing timing of the exhaust valves 5, 5.

  Since the intake VEL1 has the same configuration as that described in, for example, Japanese Patent Application Laid-Open No. 2003-172112 previously filed by the present applicant, in brief, the intake VEL1 is rotatable on a bearing on the upper part of the cylinder head SH. A hollow drive shaft 6 supported on the drive shaft 6, a drive cam 7 which is an eccentric rotary cam fixed to the drive shaft 6 by press fitting or the like, and an outer peripheral surface of the drive shaft 6 are swingably supported. Two swing cams 9, 9 that slide openly contact the upper surfaces of the valve lifters 8, 8 disposed at the upper ends of the intake valves 4, 4 to open the intake valves 4, 4, and the drive cam 7 swing. And a transmission mechanism linked to the cams 9 and 9 for transmitting the rotational force of the drive cam 7 as the swinging force of the swinging cams 9 and 9.

  The drive shaft 6 receives a rotational force from the crankshaft 02 via a timing sprocket 30 and a timing chain provided at one end, and this rotational direction is set in the clockwise direction (arrow direction) in FIG. Has been.

  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 internal axis direction. The shaft center of the cam main body has a diameter from the shaft center of the drive shaft 6. Offset by a certain amount in the direction.

  As shown in FIGS. 2 and 5 and the like, both the swing cams 9 have substantially the same raindrop shape and are integrally provided at both ends of the annular cam shaft 10. A shaft 10 is rotatably supported on the drive shaft 6 via an inner peripheral surface. The cam surface 9a formed on the lower surface of each swing cam 9 includes a base circle surface on the shaft side of the camshaft 10, a ramp surface extending in an arc shape from the base circle surface toward the cam nose portion, and the ramp surface. A lift surface connected to the top surface of the maximum lift on the tip side of the cam nose portion is formed, and the base circle surface, the ramp surface, and the lift surface correspond to the position of the swing cam 9 according to the swing position of each valve lifter 8. It comes in contact with a predetermined position on the upper surface.

  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 portion 11 b is rotatably connected to one end portion 13 a of the link rod 13 via a pin 15.

  The link arm 12 has a fitting hole in which the cam body of the drive cam 7 is rotatably fitted at the center position of a relatively large-diameter annular base 12a, while the protruding end 12b is the pin. 14 is connected to one end 11a of the rocker arm.

  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 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 of 11 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.

  As shown in FIGS. 2 to 4, the drive mechanism 19 includes an electric motor 20 fixed to one end of the housing 19 a and a rotational driving force of the electric motor 20 provided in the housing 19 a to control the control shaft. And a ball screw transmission means 21 for transmission to the motor 17.

  The electric motor 20 is composed of a proportional DC motor, and is driven by a control signal from an electronic controller 22 (ECU) that detects an engine operating state.

  The electronic controller 22 includes a crank angle sensor for detecting the engine speed, an air flow meter for detecting the intake air amount, an engine cooling water temperature sensor 61, an intake air temperature sensor, a knocking sensor 62, a vehicle speed sensor, an accelerator opening sensor, and the like. The present engine operating state is detected by detection signals from the various sensors, and a control signal is output to the throttle valve SV, the fuel injection valve, the electric motor 20, and the like.

  The ball screw transmission means 21 includes a ball screw shaft 23 disposed substantially coaxially with the drive shaft of the electric 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.

  In the ball screw shaft 23, a ball circulation groove 23a having a predetermined width is continuously formed in a spiral shape on the entire outer peripheral surface excluding both end portions, and one end portion thereof is coupled to the drive shaft 20a of the electric motor 20 and applied. The rotation transmits the rotational driving force of the electric motor 20 to the ball screw shaft 23 and allows the ball screw shaft 23 to move slightly in the axial direction.

  The ball nut 24 is formed in a substantially cylindrical shape, and a guide groove 24a that rotatably holds a plurality of balls 27 together with the ball circulation groove 23a is continuously formed in a spiral shape on an inner peripheral surface. At the same time, an axial moving force is applied to each ball 27 while converting the rotational motion of the ball screw shaft 23 to the ball nut 24 into linear motion.

  The ball nut 24 is urged to the opposite side of the electric motor 20 by the spring force of the coil spring 31 so that the backlash gap with the ball screw shaft 23 disappears, and the spring The intake valves 4 and 4 are always urged by the force in the direction of maximum lift and maximum operating angle via the control shaft 17.

  Accordingly, after the engine is stopped, the spring force of the coil spring 31 described above is surely urged toward the maximum operating angle side (the most retarded angle side) to maintain a stable state.

  Hereinafter, the operation of the intake air VEL1 will be briefly described.

  When the engine is stopped, energization of the electric motor 20 from the electronic controller 22 is cut off, so that the ball nut 24 is urged to the right in the drawing by the spring force of the coil spring 31 as shown in FIG. Then, it moves linearly in the maximum right direction (direction away from the electric motor 20), whereby the control shaft 17 rotates in one direction via the link member 39 and the linkage arm 25. That is, when the conversion power by the electric motor 20 does not work, the lift and operation angle characteristics of the intake valves 4 and 4 are stably maintained at the maximum operation angle mechanically, and the maximum operation angle becomes a so-called default position. . The control shaft 17 is configured such that the left and right maximum rotation positions are restricted by a rotation restriction stopper (not shown).

  Accordingly, as shown in FIGS. 6A and 6B (rear view), the control cam 18 rotates about the axis of the control shaft 17 with the same radius, and the thick portion is spaced downward toward the drive shaft 6. Moving. As a result, the other end portion 11b of the rocker arm 11 and the pivot point of the link rod 13 move downward with respect to the drive shaft 6. For this reason, each swing cam 9 is connected to the cam nose portion side via the link rod 13. The whole is forcibly pulled down and rotated clockwise as shown in FIG.

  Therefore, when the drive cam 7 rotates and pushes up the one end portion 11 a 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. As a result, the valve lift amount of the intake valves 4 and 4 becomes a large lift (L3) as shown by the valve lift characteristic of FIG. 7, and the operating angle D3 becomes the maximum. For this reason, the closing timing P3 (IC3) of the intake valves 4 and 4 is controlled to the retard side. Here, the operating angle refers to the crank rotation angle (double the rotation angle of the drive shaft 6) during the open period of the intake valves 4 and 4.

  Next, when the engine is started, the ignition switch is turned on, and the drive motor 07 is rotationally driven to start cranking rotation of the crankshaft 02. At the initial stage of cranking, the biasing force of the coil spring 31 is used. In the valve lift, the maximum lift and the operating angle D3 are also maintained at the maximum, and the closing timing (IC) of the intake valves 4 and 4 is retarded from the bottom dead center.

  Then, when the cranking rotation is increased to a predetermined rotation, the electric motor 20 is reversely rotated by a control signal from the electronic controller 22. When this rotational torque is transmitted to the ball screw shaft 23 and rotates, the ball nut 24 linearly moves from a position outside the figure toward the electric motor 20 against the spring force of the coil spring 31 along with this rotation ( (Near an intermediate position between FIG. 3 and FIG. 4). Thereby, the control shaft 17 is rotationally driven by a predetermined amount in the clockwise direction from the position of FIG.

  For this reason, the control cam 18 is held at a rotational angle position in the vicinity of the intermediate position in FIGS. For this reason, 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 cams 9 and the valve lifter 8 via the link member 13, thereby The lift amount of the valves 4 and 4 changes from the large lift (D3) shown in FIG. 6 to the medium lift (L2), and the operating angle also decreases from D3 to D2. As a result, the closing timing of the intake valves 4 and 4 is controlled so as to approach the piston bottom dead center.

  When the start is completed and the operation is shifted to the normal operation, the controller 22 controls the small lift (L1) to the medium lift (L2) and the small operation angle (D1) to the medium operation angle (D2). The Further, the lift phase is, for example, advanced-angle controlled by an intake VTC 2 described later. As a result, the closing timing of the intake valves 4 and 4 is advanced, and the valve overlap with the exhaust valves 5 and 5 is increased, and the pumping loss is reduced, so that the fuel efficiency is improved.

  When the accelerator is depressed to shift from the normal operation to the high load / high rotation operation region, the electric motor 20 is rotated in one direction by the control signal from the controller 22, and the control shaft 17 causes the control cam 18 to rotate counterclockwise. To rotate the shaft center downward as shown in FIGS. For this reason, the entire rocker arm 11 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 to place the entire swing cam 9. Rotate clockwise by a fixed 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 increases continuously by L3 as shown in FIG.

  That is, the lift amount of the intake valves 4 and 4 changes continuously from the small lift L1 to the maximum lift L3 in accordance with the operating state of the engine. Also changes continuously from the small lift D1 to the maximum lift D3.

  The intake VTC 2 is of a so-called vane type and is substantially the same as the structure shown in, for example, Japanese Patent Application Laid-Open No. 2007-198367 previously filed by the present applicant. explain.

  That is, a timing sprocket 30 that transmits a rotational force to the drive shaft 6, a vane member 32 that is fixed to the end of the drive shaft 6 and is rotatably accommodated in the timing sprocket 30, and the vane member 32 is hydraulically operated. And a hydraulic circuit 33 that rotates forward and backward.

  The timing sprocket 30 includes a housing 34 that rotatably accommodates the vane member 32, a disk-shaped front cover 35 that closes the front end opening of the housing 34, and a substantially disk 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 central positions so as to penetrate in the axial direction. A U-shaped seal member 38 and a leaf spring (not shown) that presses the seal member 38 inwardly are fitted and held in a holding groove that is cut out along the axial direction at the position.

  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. The front cover 35 is not shown at a position corresponding to each bolt insertion hole of the housing 34 on the outer periphery. These four bolt holes are drilled.

  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.

  Three of the vanes 32b are formed in a relatively long and narrow rectangular shape, and the other one is formed in a relatively large trapezoidal shape. The three vanes 32b are substantially the same in width and length. In contrast, the width of one vane 32b is set to be larger than that of the three vanes 32, and the weight balance of the entire vane member 32 is achieved.

  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 grooves 32c are formed on one side surface of each vane 32b in the rotational direction of the timing sprocket 30, respectively.

  A pair of coil springs that are biasing means for rotating and biasing the vane member 32 to the retarded angle side between the groove 32c of each vane 32b and the facing surface 10b of each shoe 34a facing the groove 32c. 55 and 56 are arranged, respectively. That is, when the engine is stopped or the like, when the hydraulic pressure from the hydraulic circuit 33 is not supplied and the conversion power by the hydraulic pressure does not work, the vane member 32 is mechanically stably biased to the most retarded position. It has become.

  The two coil springs 55 and 56 are formed independently and in parallel with each other, and the length in the axial direction (coil length) of each of the coil springs 55 and 56 is opposite to the one side surface of the vane 32b and the shoe 34a. It is set larger than the length between the surfaces, and both are set to the same length.

  The coil springs 55 and 56 are arranged in parallel with an inter-axis distance that does not contact each other at the time of maximum compression deformation, and a thin plate-like retainer (not shown) whose one end fits into the groove 32c of each shoe 34a. Are connected through.

  Further, four advance chambers 41 and retard chambers 42 are respectively defined between both sides of each vane 32b and both sides of each shoe 34a.

  As shown in FIG. 9, the hydraulic circuit 33 includes a first hydraulic passage 43 that supplies and discharges hydraulic oil pressure to and from each advance chamber 41, and hydraulic oil pressure to each retard chamber 42. There are two systems of hydraulic passages, a second hydraulic passage 44 for supplying and discharging gas, and a supply passage 45 and a drain passage 46 are connected to both of the hydraulic passages 43 and 44 via an electromagnetic switching valve 47 for switching the passage. 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 32d, three annular seal members 60 for separating and sealing between one end sides of the respective hydraulic passages 43 and 44 are fitted. It is fixed.

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

  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. The annular chamber 44a and a second oil passage 44b communicating with each retarded angle chamber 42 are provided.

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

  This electronic 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 30 and the drive shaft 6 by signals from the crank angle sensor and the intake drive shaft angle sensor. Detected.

  Then, by the switching operation of the electromagnetic switching valve 47, hydraulic oil is supplied to the retarding chamber 42 when the engine is started, and thereafter hydraulic oil is supplied to the advance chamber 41.

  A lock mechanism is provided between the vane member 32 and the housing 34 to restrain the rotation of the vane member 32 relative to the housing 34 and to release the restraint.

  That is, as shown in FIG. 9, this locking mechanism is provided between one vane 32b having a large width and the rear cover 36, and is formed along the axial direction of the drive shaft 6 inside the vane 32b. The sliding hole 50, a covered cylindrical lock pin 51 slidably provided inside the sliding hole 50, and a cross-shaped cup-shaped engagement fixed in the fixing hole provided in the rear cover 36. It is provided in the hole constituting portion 52 and is held by an engagement hole 52a in which the tapered tip portion 51a of the lock pin 51 is engaged and disengaged, and a spring retainer 53 fixed to the bottom surface side of the sliding hole 50. The spring member 54 is configured to urge the pin 51 in the direction of the engagement hole 52a.

  The engagement hole 52a is supplied with oil pressure in the retard chamber 42 or oil pressure of the oil pump 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 30 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 chamber 42 into the engagement hole 52a or the hydraulic pressure of the oil pump, and the engagement with the engagement hole 52a is released.

  FIG. 7 showing the valve lift characteristics of the intake valves 4 and 4 shows a case where the opening and closing timing of the intake valves 4 and 4 is controlled to the most retarded angle side by the intake VTC2. Here, the peak lift phase is slightly inclined toward the advance side as the lift increases, but this does not have a significant effect.

  The operation of the intake VTC 2 is substantially the same as that of the Japanese Patent Application Laid-Open No. 2007-198367.

  The exhaust VTC3 has exactly the same configuration as the intake VTC2, and when the hydraulic pressure of the hydraulic circuit 33 does not act when the engine is stopped, the vane member 32 is rotated by the spring force of the coil springs 55 and 56 on the maximum retard side. It is designed to stably bias the moving position. Further, the exhaust VTC 3 inputs an exhaust valve lift phase control signal which is an opening / closing valve timing of the exhaust valves 5 and 5 from the electronic controller 22 based on an information signal from an exhaust drive shaft angle sensor, etc. It comes to control.

  Hereinafter, the operation of the present embodiment will be described with reference to FIG. 11. Prior to this, the problems of the prior art will be described with reference to FIG. In each figure, the first cylinder (# 1 cylinder) in the compression stroke and the third cylinder (# 3 cylinder) in the intake stroke are representatively shown. In this state, the second cylinder (# 2 cylinder) Is the expansion stroke, and the fourth cylinder (# 4 cylinder) is the exhaust stroke.

  Considering the crank angle when the engine is stopped, if the # 1 cylinder is a cylinder in the compression stroke, the rotational phase (crank angle) of the crankshaft 02 of the # 1 cylinder, that is, the position of the piston 01 is lower than the top dead center TDC. There are many cases of stopping near the middle of the dead center BDC. This is because when it approaches the top dead center, it is returned to the bottom dead center side by the compression pressure acting on the piston 01, and the crankshaft 02 rotates counterclockwise in the figure. Since the # 2 cylinder in the stroke approaches the top dead center (the # 1 cylinder is the bottom dead center), the crankshaft 02 is returned in the clockwise direction by the compression pressure (compression) acting on the piston 01. Due to this operational balance, the crank angle of the # 1 cylinder often stops in the middle of the top dead center and the bottom dead center.

  After the engine stops, the atmosphere immediately enters the combustion chamber 04 from between the piston ring of the piston 01 and the inner wall surface of the cylinder, and the inside of the cylinder becomes atmospheric pressure.

  Next, when the engine is started, the crankshaft 02 rotates in the clockwise direction from the atmospheric pressure in the cylinder, the piston 01 rises, and the compression reaches a peak at the top dead center. This top dead center compression is larger as the initial position of the piston 01 is lower and smaller as it is higher. That is, it is influenced by the stop crank angle (stop piston position) of the # 1 cylinder.

  By the way, although it is the above-mentioned stop crank position, since it is influenced by the delicate pressure balance and friction balance when the engine is stopped, the variation is essentially large. Even if the stop crank position is positively controlled using the drive motor 07 as the crank position control means, some variation still remains, and therefore the top dead center compression essentially varies.

  For this reason, the cranking becomes unstable due to the variation in the first top dead center compression of the start cranking, or the variation becomes large. As a result, the startability becomes unstable.

  In contrast, in the present embodiment, the top dead center compression is not affected by the unstable stop crank position, and is determined by the mechanically stable closing timing of the intake valves 4 and 4, that is, the default IC. The startability can be improved by stabilizing the cranking.

  That is, as shown in FIG. 11, when the engine is stopped, the intake valves 4 and 4 are held at a mechanically stable closing timing (default IC, IC3). At that time, since the intake valves 4 and 4 are set to the most retarded angle defaults by the intake VTC2 and the maximum operating angle defaults by the intake VEL1, the default IC (IC3) is close to the top dead center. Controlled to a delayed state. On the other hand, the stop crank angle Z is controlled to the advance side from the default IC by the drive motor 07 described above. This advance amount is set so as not to be slower than the default IC even when variation in stop control is taken into consideration.

  Here, when the aim of the stop crank angle Z is Z0 and the variation is ± α, even the most retarded Z0 + α is a position advanced from the default IC (Z0 + α <default IC).

  Therefore, if the control range is set to Z0−α to Z0 + α, the intake valves 4 and 4 are set to be surely opened at the stop crank position even if the stop crank position varies.

  Then, cranking is started, and even if the piston 01 rises, the intake valves 4 and 4 are open, so the in-cylinder pressure does not immediately rise and the atmospheric pressure is maintained. The compression is generated when the intake valves 4 and 4 are closed toward the default IC. Here, the default IC is a mechanically stable position and there is almost no variation, so that the top dead center compression is stabilized. As a result, cranking is stabilized and good startability is obtained.

  Further, the advancement of the stop crank angle Z means that the position of the piston 01 of another intake stroke cylinder (# 3) rises as shown in FIG. Can be lengthened.

  As a result, sufficient fuel atomization time can be taken and the homogeneity of the air-fuel mixture can be increased. Therefore, the combustibility (complete explosion performance) at the time of starting becomes favorable, and the starting performance can be improved. On the other hand, since the closing timing of the intake valves 4 and 4 is controlled to be retarded, the top dead center compression is sufficiently lowered, an excessive increase in load torque at the initial stage of the start is suppressed, and the occurrence of start-up vibration can be suppressed.

  Although FIG. 11 shows a state in which the crankshaft 02 and the piston 01 are directly connected, the actual piston 01 is connected to the crankshaft 02 via the connecting rod 03 as shown in FIG. Since the relationship between the angle and the vertical movement of the piston 01 is qualitatively the same, the description is simplified for easy understanding.

  FIG. 12 is a time chart showing the state of each cylinder when the engine is started in the present embodiment.

  The order of ignition is # 1 cylinder, # 3 cylinder, # 4 cylinder, # 2 cylinder, and this is repeated. When the compression stroke is the # 1 cylinder, the crank angle Z is advanced with respect to the default IC (IC3) of the retard angle, and the top dead center compression is stabilized by the above-described mechanism. On the other hand, in the # 3 cylinder in the intake stroke, the stop crank angle Z is advanced to the vicinity of the default IO (IO3) that is the opening timing of the intake valves 4, 4. A sufficient intake stroke of 01 can be secured. Therefore, sufficient fuel atomization time can be taken and the homogeneity of the air-fuel mixture can be increased. As a result, the starting combustibility (complete explosion performance) is improved and the starting performance can be further improved.

  On the other hand, since the closing timing (IC3) of the intake valves 4 and 4 is retarded to the maximum, the top dead center compression is sufficiently reduced, and an excessive increase in load torque at the beginning of the start is suppressed, so that the start-up vibration can be suppressed. .

  The next # 4 cylinder can take a sufficient intake stroke from the intake valve 4 and 4 opening timing IO3, and similarly, the closing timing (IC3) of the intake valves 4 and 4 is retarded to the maximum. The same vibration suppression effect as that of the three cylinders can be obtained.

  Further, in the next # 2 cylinder, before reaching the intake stroke, the intake VEL1 is converted to the intermediate operating angle D2 and the intake VTC2 is converted to the intermediate phase, so that the closing timing of the intake valves 4 and 4 approaches the bottom dead center (IC2). . Since the # 4 cylinder has slightly increased rotation at the third combustion, it is difficult for start-up vibration to occur, so the intake timing of the intake valves 4 and 4 closes to the bottom dead center to increase the intake charge efficiency, Increase combustion torque. This further promotes the increase in rotation.

  Further, in the next cylinder # 1, the closing timing IC of the intake valves 4 and 4 is further lowered to the bottom dead center by being converted to the small operating angle D1 by the intake VEL1 and advanced by the intake VTC2 until the intake stroke is reached. Move closer (IC1). As a result, the intake charge efficiency is further increased, and the rotational rise is accelerated, so that quick startability can be realized.

  FIGS. 13 and 14 are flowcharts showing the operation of the present embodiment, and a control flowchart by the electronic controller 22. FIG. 13 is a flowchart of the part leading to the engine stop, and FIG.

  First, in step 1 shown in FIG. 13, it is determined whether or not the ignition switch is turned off or an engine stop condition such as an idle step condition is satisfied. If it is confirmed that the stop condition is satisfied, a control signal to the default position is output in step 2. In other words, the closing timing IC of the intake valves 4 and 4 attempts to mechanically settle to the most retarded IC3, but further outputs a conversion signal to the most retarded angle side by the intake VTC2 and operates maximum by the intake VEL1. The conversion signal to the corner is output. With these conversion powers, the transition to the closing timing IC3 can be accelerated and more reliable.

  In step 3, the actual positions of the closing timing and opening timing of the intake valves 4, 4 by the intake VEL1 and the intake VTC2 are detected.

  Subsequently, in step 4, it is determined whether or not the actual closing timing IC by both 1 and 2 is sufficiently close to the target closing timing IC3. That is, it is determined whether or not the absolute value of the actual closing timing IC-IC3 is less than a predetermined minute angle ΔIC. Here, if it is determined that the number is less, the process proceeds to step 5.

  In step 5, the stop position of the crank angle Z is controlled by the drive motor 07 (Z0 ± α), and is controlled to Z0 on the advance side with respect to the closing timing IC3. Even when the control variation ± α is taken into consideration, the closing timing IC3 is set to the advance side. In this state, a process for stopping the engine in step 6 is performed.

  Therefore, as shown in FIG. 11, a desired stop crank angle and closing timing of the intake valves 4 and 4 can be obtained.

  If it is determined in step 4 that the angle is not less than the minute angle ΔIC, the process proceeds to step 7. That is, it recognizes that a so-called bleed phenomenon due to a failure or the like has occurred in the intake VEL1 or the intake VTC2, and determines that decompression by the closing timing IC cannot be performed, shifts to step 7 and delays the crank stop position with respect to Z0. Change to crank stop position control targeting Z1 on the corner side. Therefore, in this case, the initial top dead center compression is reduced not by the closing timing IC but by the stop crank angle to ensure the minimum startability. In addition, the crank stop position control itself is canceled, but the variation of the stop crank angle increases, but it shifts to the retard side, so the first top dead center compression is similarly reduced to ensure the minimum startability. .

  Next, the flowchart of FIG. 14 will be described. First, when it is determined in step 10 that the engine start condition is satisfied, it is confirmed in step 11 that the actual closing time IC is sufficiently close to the default IC 3, cranking is started in step 12, and the process proceeds to step 13.

  In step 13, as in step 2, a signal for conversion to the closing timing IC3 is output even after the start of cranking. Thereby, it is possible to further match the closing timing IC3.

  Then, in step 14, it is determined whether or not the predetermined crank angle (Za) has been reached by cranking. If it is determined that the crank angle has not been reached, the process returns to step 13, but if it is determined that it has been reached, the process proceeds to step 15. To do.

  In this step 15, the intake VEL1 and the intake VTC2 output signals for converting the operation angle of the intake valves 4, 4 to the intermediate operation angle and the lift phase to the intermediate phase, respectively. As a result, the closing timing IC2 (intermediate phase, intermediate operating angle) is controlled in the # 2 cylinder, which advances the intake (intake) stroke next to the # 3 cylinder and the # 4 cylinder. At this time, since it is the third combustion (inhalation), the influence of vibration is reduced because the cranking rotational speed is slightly increased. Therefore, by increasing the intake charging efficiency by bringing the closing timing IC close to the bottom dead center, it is possible to increase the torque and further increase the cranking rotation.

  Further, in step 16, it is determined whether or not the crank angle (Zb) has reached a predetermined crank angle. If it is determined that the crank angle (Zb) has not been reached, the process returns to step 5. The control signal is output by the intake VEL1 and the intake VTC2 so that the operating angle becomes a small operating angle and the lift phase becomes the advanced phase. As a result, the # 1 cylinder, which advances the intake (intake) stroke next to the # 3 cylinder, the # 4 cylinder, and the # 2 cylinder, is controlled to the closing timing IC1 (small operating angle, advanced phase). At this time, since it is the fourth combustion (inhalation), the influence of vibration is further reduced by further increasing the crank rotation speed. For this reason, the closing timing IC is made closer to the bottom dead center (IC1), and the intake charge efficiency is further improved to further increase the torque and further increase the cranking rotation. As a result, a good and agile engine startability can be realized.

  In step 18, it is determined whether or not the engine water temperature has reached a predetermined temperature. If it is determined that the engine water temperature has not yet reached, the process returns to step 17, but if it is determined that it has reached, it is determined that warm-up has been completed. If yes, go to Step 19. Here, the control shifts to normal control based on the engine speed and the load map.

If it is determined in step 11 that the actual IC is greatly deviated from the default IC 3, it is determined that the intake VEL1 and intake VTC2 are caused by a failure or the like, and steps 20 to 22 are performed. A transition is made to continue outputting a signal for conversion to the default IC3, and a transition is made to the engine failsafe control mode. That is, in step 20, a signal for starting cranking is output, and in step 21, the conversion signal to the default position (IC3) is output to the intake air VEL1 and the intake air VTC2, while the minimum IC is based on the actual IC in step 22. Fail-safe control that ensures drivability is performed.
[Second Embodiment]
FIG. 15 shows a second embodiment of the present invention, which is applied to a so-called direct injection type internal combustion engine in which fuel injection by the fuel injection valve is directly performed in a cylinder (inside the combustion chamber 04). Independently of the drive, fuel is injected and ignited and burned (self-combustion) to push down the piston 01 to start rotation (self-start). Other configurations are the same as those of the first embodiment.

  The crank position of the # 1 cylinder in the compression stroke shown in FIG. 15 is Z as in the case of FIG. 11, and the closing timing of the intake valves 4 and 4 is also the default IC3.

  Here, the cylinder in the expansion stroke is the # 2 cylinder, and the piston 01 of the # 1 cylinder has advanced to the bottom dead center side. Therefore, the piston 01 of the # 2 cylinder has advanced to the top dead center side. Accordingly, the piston stroke in the expansion stroke can be taken longer and the expansion work due to combustion can be increased, so that the startability by self-combustion can be improved by effectively depressing the piston 01.

  Further, the default timing EO1 of the opening timing (EO) of the exhaust valves 5 and 5 is the most retarded angle and is close to the bottom dead center. As a result, it is possible to suppress combustion pressure loss due to the opening of the exhaust valves 5 and 5 during combustion, further increase the expansion work, and effectively push down the piston 01.

  On the other hand, the position of the piston 01 of the # 1 cylinder in the compression stroke is the same as in FIG. 11 and is relatively close to the bottom dead center, and the closing timing (IC3) of the intake valves 4 and 4 is the most retarded position. . Therefore, the top dead center compression is determined by the IC 3 as in the first embodiment, and a low and stable compression can be obtained.

  Accordingly, the compression top dead center of the # 1 cylinder is easily realized by the self-combustion torque, and the self-startability can be enhanced from this point.

  FIG. 16 is an operation time chart of each cylinder, and self-combustion is performed in the # 2 cylinder in the expansion stroke at the stop crank angle Z, but from the stop crank angle Z to the opening timing (EO1) of the exhaust valves 5 and 5. Since the crank angle, that is, the piston stroke is expanded, the period and stroke in which the piston 01 is pushed down by combustion can be increased to increase the expansion work and increase the driving torque.

  Further, since the compression of the # 1 cylinder in the compression stroke can be stably reduced, the driving top torque can easily overcome the compression top dead center (maximum compression) of the # 1 cylinder. That is, it is not pushed back toward the top dead center, and positive rotation by self-combustion is reliably obtained.

  Further, considering the # 1 cylinder that burns next to the # 2 cylinder, the stop crank position Z is advanced in the initial compression stroke, and the closing timing (IC3) of the intake valves 4 and 4 is retarded. A part of the air in the cylinder is discharged from the stop crank angle Z to the closing timing IC3 to the closing timing IC3, and then the piston 01 rises toward the top dead center, and in the middle of the cylinder Inject fuel and ignite. Since a certain amount of compression is generated at the top dead center, when the ignition is performed, a combustion torque higher than the self-combustion of the # 2 cylinder is obtained, and an increase in rotation is promoted.

  Next, considering the # 3 cylinder that burns next to the # 1 cylinder, the crank angle Z is in the advanced position in the initial intake stroke, so that the intake stroke of the piston 01 can be made longer. Atomization is promoted and combustion is stabilized. In addition, since the closing timing IC3 of the intake valves 4 and 4 is retarded to the maximum, an abrupt increase in intake charge efficiency is suppressed, and vibrations that cause problems in the extremely low rotation range are suppressed by the decompression effect. it can.

  In the next combustion of the # 4 cylinder, the intake valves 4 and 4 can be sucked from the opening timing IO3, so that the intake stroke can be further lengthened and the atomization can be further improved, and the vibration reduction effect by the retard control of the intake valve closing timing IC3. Is maintained.

  In the next combustion of the # 2 cylinder (second time), the operating angle reduction control by the intake air VEL1 and the advance angle control by the intake air VTC2 are performed by the intake stroke, and the closing timing IC changes to the bottom dead center side. As a result, the intake charge efficiency (ηv) is improved and the torque is increased, so that the engine rotation is promoted and agile startability is obtained. Here, since the rotation has already increased to some extent, the start-up vibration is an area where there is no problem.

  Then, in the next combustion of the # 1 cylinder (second time), the closing timing IC of the intake valves 4 and 4 is close to the bottom dead center by the operation angle reduction control by the intake VEL1 and the advance angle control by the intake VTC2 until the intake stroke. Advance. As a result, the intake charge efficiency (ηv) is further improved to increase the torque, and the startability is promoted by the more agile rotation.

  Subsequently, in the next combustion of the # 3 cylinder (second time), the closing timing IC of the intake valves 4 and 4 is retarded again (IC3). This reduces the torque to suppress rotational overshoot. Alternatively, it is also possible to perform a countermeasure against vibration by performing a similar retard angle control of the closing timing IC at a rotational speed that coincides with the resonance point of the vehicle drive system, and re-decompressing.

  Further, the exhaust VTC3 advances the lift phase of the exhaust valves 5 and 5 to an intermediate phase to advance the opening timing EO of the exhaust valves 5 and 5 (EO1 → EO2). As a result, before the temperature of the combustion gas is lowered in the expansion stroke, the combustion gas is discharged to the exhaust port EP side, thereby promoting an increase in the downstream catalyst temperature and reducing exhaust emission. This is to promote exhaust emission reduction as the next step that realizes agile start. However, as described above, since the start time is shortened, the effect of reducing exhaust emission is great.

  Then, after the engine water temperature reaches a predetermined temperature, the intake VEL1, the intake VTC2, and the exhaust VTC3 are controlled by the engine rotation and the load map according to the demand for operability.

  Hereinafter, control by the controller in the present embodiment will be described with reference to a control flowchart shown in FIG.

  First, in step 20, it is determined whether or not the engine start condition is satisfied. If the engine start condition is not satisfied, the process is terminated. If the engine start condition is satisfied, the process proceeds to step 21.

  In this step 21, the cylinder in the current expansion stroke is determined by the crank angle sensor and the intake side drive shaft angle sensor. Here, it is assumed that the expansion stroke cylinder is, for example, the # 2 cylinder.

  In step 22, a signal for converting the closing timing of the intake valves 4 and 4 to IC3 by the intake VEL1 and the intake VTC2 is output. Thereby, it can be made to correspond to the said closing time IC3.

  Then, in step 23, in-cylinder injection and ignition are performed on the # 2 cylinder in the expansion stroke, thereby starting combustion. That is, self-combustion is performed.

  By this self-combustion, engine rotation is started in step 24, and fuel injection and an ignition signal are output in step 25 according to the ignition sequence.

  In step 26, it is determined whether or not the crank angle Z has been rotated to a predetermined crank angle (Za). If it has not yet rotated, the process returns. If it has rotated, the process proceeds to step 27.

  In this step 27, signals are output to the intake VEL1 and the intake VTC2 so as to have an intermediate operating angle and an intermediate phase so as to be in time for the intake stroke of the next combustion of the # 2 cylinder (second cycle). 4 is closed to IC2.

  In step 28, as in step 26, it is determined whether or not the crank angle Z has rotated to a predetermined crank angle (Zb). If it has not yet rotated, the process returns. If it has rotated, the process proceeds to step 29. .

  In step 29, a small operating angle control signal is output to the intake VEL1 and an advance phase control signal is output to the intake VTC2 so as to be in time for the intake stroke of the next combustion of the # 1 cylinder (second cycle). , 4 is converted to IC1.

  In step 30, as in step 28, it is determined whether or not the crank angle Z has rotated to a predetermined crank angle (Zc). If it has not yet rotated, the process returns. If it has rotated, the process proceeds to step 31. .

  Here, the intake valve 4 outputs the maximum operating angle control signal to the intake VEL1 and the most retarded phase control signal to the intake VTC2 so as to be in time for the intake stroke of the next combustion of the # 3 cylinder (second cycle). , 4 is converted to IC3.

  In step 32, it is determined whether or not the engine water temperature has reached a predetermined temperature. If the engine water temperature has not yet reached the predetermined temperature, the process returns, and if it has reached, the process proceeds to step 33 assuming that the warm-up is complete.

  In this step 33, control signals are appropriately output to the intake VEL1, intake VTC2 and exhaust VTC3 according to the engine rotation and load map, and the operating angle, lift phase of the intake valves 4, 4 and the lift phase of the exhaust valves 5, 5 are set. Each engine is controlled to achieve optimum engine performance according to engine operating conditions.

  The effects of these series of controls are the same as described above.

  The present invention is not limited to the configuration of the embodiment. Moreover, on the intake valves 4 and 4 side, the example in which the intake VEL1 and the intake VTC2 are used together has been shown, but only one of them may be used. Further, although an example in which the exhaust VTC 3 on the exhaust valves 5 and 5 side is used together is shown, this is not always necessary.

  In addition, the intake valve closing timing IC and opening timing IO and the exhaust valve opening timing and opening timing may be exactly the timing of starting and ending lift of the intake valve and exhaust valve, but in a minute lift region having a gradual lift gradient. The lift may be started and finished in a state excluding a certain so-called ramp region. The latter case corresponds to the start / end of substantial combustion gas suction and the start / end of combustion gas discharge.

The technical ideas of the invention other than the claims ascertained from the embodiment will be described below.
[Claim a] The variable valve operating apparatus for an internal combustion engine according to claim 1,
A variable valve operating apparatus for an internal combustion engine, wherein the closed state of the intake valve is mechanically stabilized by an urging force of an urging means.
[Claim b] In the variable valve operating apparatus for an internal combustion engine according to claim a,
A variable valve operating apparatus for an internal combustion engine, wherein the movable portion abuts against a stopper to mechanically stabilize the closed state of the intake valve.
[Claim c] In the variable valve operating apparatus for an internal combustion engine according to claim 1,
A variable valve operating apparatus for an internal combustion engine, wherein the closing timing is made variable by increasing the operating angle of the intake valve, and mechanically stabilized at the maximum operating angle when the engine is stopped.
[Claim d] In the variable valve operating apparatus for an internal combustion engine according to claim c,
A first variable mechanism that variably controls only the lift phase of the intake valve; and a second variable mechanism that variably controls the operating angle of the intake valve;
The variable valve operating apparatus for an internal combustion engine, wherein the first variable mechanism mechanically stabilizes the lift phase of the intake valve at the most retarded position.

According to the present invention, since both of the two variable mechanisms are mechanically stabilized so that the closing timing of the intake valve is retarded, the closing timing of the intake valve can be sufficiently delayed. It will be possible to start up even faster.
[Claim e] In the variable valve operating apparatus for an internal combustion engine according to claim 1,
An exhaust valve variable mechanism that variably controls the opening timing of the exhaust valve is provided, and the exhaust valve variable mechanism is characterized in that the opening timing of the exhaust valve is stabilized on the retard side when the engine is stopped. Variable valve device.

According to the present invention, the expansion work can be sufficiently performed to increase the combustion torque, and a sufficient start of rotation can be achieved at the time of start-up with high engine friction.
[Claim f] The variable valve operating apparatus for an internal combustion engine according to claim 1,
A variable valve operating apparatus for an internal combustion engine, wherein when the engine is stopped, the closing timing of the intake valve passes through the bottom dead center and is stabilized on the top dead center side.
[Claim g] In the variable valve operating apparatus for an internal combustion engine according to claim 1,
A variable valve operating apparatus for an internal combustion engine, wherein the variable valve operating apparatus is applied to an engine that automatically stops regardless of operation of an ignition switch.
(Claim h) In the internal combustion engine start system according to claim 2,
A start system for an internal combustion engine, which is applied to a multi-cylinder engine having a cylinder whose crank angle is 180 ° different from that of one cylinder.

  According to the present invention, when the engine is stopped, the crank position of the cylinder in the compression stroke is controlled by the crank stop position control means to the advance side rather than the crank position at the constant position on the retard side. Since the crank position of the cylinder is also advanced, the intake stroke of the piston can be lengthened. Therefore, sufficient fuel atomization time can be taken and the homogeneity of the air-fuel mixture can be increased.

On the other hand, since the closing timing of the intake valve is retarded, an excessive load torque increase due to the compression at the initial stage of the start can be suppressed, and vibration at the start can be suppressed.
[Claim i] The internal combustion engine start system according to claim h,
Cylinder discriminating means for determining the cylinder of the expansion stroke when the engine is stopped;
Fuel injection means for injecting fuel into the cylinder;
Fuel ignition means for igniting the fuel in the cylinder,
An internal combustion engine start system characterized in that fuel is injected by the fuel injection means into an expansion stroke cylinder determined to be an expansion stroke by the cylinder determination means, and the engine is started by igniting the fuel by the fuel ignition means.

According to the present invention, when the cylinder in the expansion stroke is burned and started independently, the top dead center compression of the cylinder in the compression stroke is reduced and stable. Since a certain cylinder can easily overcome the top dead center, the self-starting performance is improved.
(Claim j) The internal combustion engine start system according to claim h,
One of a plurality of cylinders is controlled by the crank position control means so that the position of the piston is at the bottom dead center side,
A starting system for an internal combustion engine, wherein the closing timing of the intake valve is controlled to the top dead center side by the variable valve operating device.

According to the present invention, the closing timing of the intake valve at the time of starting the engine is stabilized and the top dead center compression is determined by the closing timing of the intake valve, so that the top dead center compression is stabilized and the start cranking characteristic is improved. As a result, the start can be stabilized.
(Claim k) The internal combustion engine starting system according to claim 2,
Before starting the engine, if the closing timing of the intake valve deviates from a predetermined position on the retard side by a predetermined amount or more, the control of the crank position control means is changed or canceled, and the internal combustion engine is started system.

According to the present invention, the top dead center compression can be reduced to some extent even when the variable valve operating apparatus has the above-described occurrence, and the minimum startability can be ensured.
[Claim 1] The internal combustion engine start system according to claim 2,
A starting system for an internal combustion engine, characterized in that it is used in an engine that automatically stops regardless of the operation of an ignition switch.
[Claim m] In the internal combustion engine start control apparatus according to claim 3,
The piston position is controlled to the bottom dead center side by the crank position control means,
An internal combustion engine start control device characterized in that the closing timing of the intake valve is controlled to the top dead center side by the variable valve operating device.
[Claim n] In the internal combustion engine start control apparatus according to claim 3,
If the closing timing of the intake valve deviates from a predetermined position on the retarded side by a predetermined amount or more immediately before stopping the engine, the control of the crank position control means is changed or canceled, and the internal combustion engine is started Control device.
[Claim o] In the internal combustion engine start control apparatus according to claim 3,
A start control device for an internal combustion engine, which is applied to an engine that automatically stops regardless of operation of an ignition switch.

01 ... Piston 02 ... Crankshaft 04 ... Combustion chamber 07 ... Drive motor (crank position control means)
1 ... Intake VEL (second variable mechanism)
2 ... Intake VTC (first variable mechanism)
3. Exhaust VTC (third variable mechanism)
DESCRIPTION OF SYMBOLS 4 ... Intake valve 5 ... Exhaust valve 6 ... Drive shaft 19 ... Drive mechanism 20 ... Electric motor 21 ... Ball screw transmission means 22 ... Electronic controller 31 ... Coil spring (biasing means to the maximum operating angle by the side of intake VEL)
55, 56 ... Coil spring (means for urging in the most retarded direction on the intake VTC side)

Claims (3)

A variable valve operating device used for an internal combustion engine capable of controlling a stop position of a piston to a position passing through a bottom dead center when the engine is stopped,
At least the intake valve closing timing can be variably controlled,
An internal combustion engine characterized in that, when the engine is stopped, the closing timing of the intake valve is mechanically stabilized at a position more retarded than a position most delayed in the stop position control range of the piston. Variable valve gear for engine. A variable valve system configured such that at least the closing timing of the intake valve can be variably controlled, and when the engine is stopped, the closing timing of the intake valve is stopped at a fixed position on the retard side from the bottom dead center;
Crank position control means for controlling the stop position of the piston to a position advanced from the closing timing of the intake valve when the engine is stopped;
An internal combustion engine starting system comprising: A variable valve gear capable of variably controlling at least the closing timing of the intake valve; and
Crank position control means for controlling the stop position of the crankshaft and the piston when the engine is stopped;
An internal combustion engine start control device comprising:
When the engine is stopped, the intake valve is controlled by the variable valve device so that the closing timing of the intake valve becomes a fixed position on the retard side of the bottom dead center, and the stop position of the piston is controlled by the crank position control means. A start control device for an internal combustion engine, wherein the start control device controls to a position advanced from the closing timing of the engine.

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