JP4931740B2 - Control device for internal combustion engine - Google Patents

Description

  The present invention, for example, stops the operation of intake and exhaust valves of some cylinders of an internal combustion engine for automobiles, so-called cylinder deactivation, thereby improving fuel efficiency and suppressing torque shock when shifting to cylinder deactivation. The present invention relates to a control device for an internal combustion engine.

  Various conventional control devices for an internal combustion engine that can deactivate some cylinders are provided, and one of them is described in Patent Document 1 below.

  In brief, a hydraulic cylinder deactivation mechanism that switches between full cylinder operation that activates all of the plurality of cylinders and deactivation operation that deactivates some of the plurality of cylinders, and supply or interruption of hydraulic pressure to the cylinder deactivation mechanism And a control means for controlling the drive of the solenoid valve, wherein the control means sets the drive timing of the solenoid valve based on the operating state of the engine detected by the operating state detecting means. It is like that.

As a result, each cylinder is deactivated and deactivated by operating the cylinder deactivation mechanism at an appropriate timing.
JP-A-10-82334

  However, in the control device for the conventional internal combustion engine, when the switching between the full cylinder operation and the partial cylinder deactivation operation of each cylinder is performed in a stepwise manner, the opening amount of the throttle valve is greatly changed. Therefore, it is considered that the generation of the torque step is suppressed. This is because it is necessary to absorb a large change in the throttle valve opening amount that the intake air amount greatly changes due to the change in the number of operating cylinders. For this reason, during all-cylinder operation in which the throttle valve opening is controlled to be small, a pumping loss is likely to occur due to negative pressure in the intake manifold downstream of the throttle valve, resulting in a significant reduction in fuel efficiency. End up.

  In addition, when shifting to the cylinder deactivation operation, in order to increase the output torque by the cylinder that is operating, the amount of intake air per cylinder is increased by controlling the amount of opening of the throttle valve relatively large, Suppresses the generation of torque steps as an engine.

  However, in this case as well, the engine load is constantly changing, and in order to cope with this load fluctuation, the throttle valve must be controlled to an intermediate opening amount instead of the fully open state. For this reason, the pumping loss cannot be sufficiently reduced. Therefore, it is considered that the fuel efficiency cannot be sufficiently improved even in this driving state.

  In addition, the torque lift step between the all cylinder operation and the partial cylinder deactivation operation, but the internal pressure of the intake manifold changes greatly before and after the throttle valve opening change, so for a moment until this change ends There is a possibility that a torque step will occur and a torque shock will remain.

The present invention has been devised in view of the technical problem of the conventional variable valve operating device, and the invention according to claim 1 is a valve stop control of the intake valve and the exhaust valve of the first cylinder group. When performing, the valve overlap of the intake valve and the exhaust valve of the second cylinder group is set to be substantially the same as the valve overlap before the valve stop, and the intake valve of the second cylinder group is closed. By controlling the timing so that the lift amount is increased while controlling the timing to the retarded angle side before the valve stop, the engine torque of the second cylinder group is compared with the first cylinder group before the valve stop and the first cylinder group before the valve stop. The engine torque is configured to be almost equal to the engine torque of the second cylinder group.

  In the case of the present invention, at the time of the valve stop transition, the torque step generation is not controlled depending on the throttle valve opening amount, but is controlled mainly by the lift amount (opening amount) of the intake valve.

  That is, for example, when applied to a V-type internal combustion engine, for example, during no-load or light-load operation such as idling operation, all cylinder operation is performed in both the left and right banks to suppress destabilization of the engine speed. Performs control to reduce the intake air amount by controlling the lift amount of each intake valve of all the cylinders to be small by setting the opening amount of the throttle valve to a large opening state.

  On the other hand, when the engine load rises due to low / medium load operation such as steady operation, for example, the intake / exhaust valve on the right bank side is stopped and only the intake / exhaust valve on the left bank side is operated. However, in this case as well, the throttle valve opening amount is set to the large opening state, and the lift amount of each intake valve of the cylinder on the left bank side is controlled from a small lift to a relatively large middle lift to increase the intake air amount. I do.

  Accordingly, since the throttle valve has a large opening range during any engine operation, the pumping loss can be sufficiently reduced, and as a result, fuel consumption can be improved.

  In addition, since the intake air amount per cylinder that operates due to the difference in the lift amount can be relatively reduced in all-cylinder operation and relatively large in partial deactivation operation, the torque of the entire engine can be It is possible to get closer. Furthermore, if the throttle valve opening is maintained at a large opening when switching from all cylinder operation to partial cylinder deactivation operation, it is possible to suppress changes in the intake manifold internal pressure and reduce transient torque shocks.

The invention according to claim 2 is characterized in that the valve stop control is performed when the engine speed increases or when the engine load increases.

When the valve stop control of the intake valve according to the first aspect of the invention is performed, the throttle valve maintains a large opening state before and after the switching, so that the pressure in the intake manifold is substantially the same before and after the throttle valve. As mentioned above, the pumping loss is reduced and the fuel efficiency is improved in both the all-cylinder operation and the partial rest operation (reduced cylinder operation) as described above. Even when the engine is increased or the engine load is increased, the pressure difference in the intake manifold before and after switching is small as described above, so the transient torque step due to the transient change in intake pipe pressure is reduced and transient shock is reduced. it can.
That is, after the valve stop transition, the intake valve opening timing is constant, that is, the time area of the lift amount can be increased with constant valve overlap, so that it is not easily affected by residual gas fluctuations and torque steps are unlikely to occur.

According to a third aspect of the present invention, the first cylinder group and the second cylinder group have the same number of cylinders, and the intake valve and the exhaust valve of the first cylinder group are controlled by valve stop control of the first cylinder group. The lift amount of the intake valve of the second cylinder group is substantially equal to the time area of the lift amount of the intake valve of the first cylinder group before the valve stop control. The second cylinder group is controlled to be substantially the same as the sum of the lift amount of the intake valve and the time area.

  According to the present invention, for example, in a light load region such as during idling, the lift amount of the intake valve is small, and the choking phenomenon occurs in the region where the flow velocity of the intake air into the cylinder is close to the speed of sound. The time area of the lift amount is proportional to the intake air amount.

  Therefore, when the operation range is changed and the entire cylinder operation is switched to the reduced cylinder operation, the intake air amount of the entire engine can be made substantially equal, so that a torque step hardly occurs.

According to a fourth aspect of the present invention, in the light load operation of the engine, the control is switched to the all cylinder operation in which the valve is not stopped.

  According to the present invention, it is possible to avoid deterioration of engine rotation fluctuation due to reduced-cylinder operation in a light load region such as idling operation. As a result, the engine can be stabilized at a light load.

The invention according to claim 5 is characterized in that the air-fuel ratio is controlled to the lean side in the case of light load operation of the engine.

  According to the present invention, the lift amount of the intake valve can be increased in a light load region such as idling operation, so that variations in combustion due to variations in lift amounts between cylinders can be reduced. Further, the fuel efficiency is improved because the thermal efficiency is improved by improving the specific heat ratio of the air-fuel mixture by leaning the air-fuel ratio. In addition, the increase in NOx due to leaning is light, so there is no real harm.

According to a sixth aspect of the present invention, the first lift amount variable mechanism has a first control unit that continuously varies the lift amount of the intake valve by applying a driving force, and the driving force is not applied. In this case, the lift amount of the intake valve is mechanically held at an intermediate lift position that is not zero lift.

  According to this invention, even if there is a failure in the electrical system, the intake valve of the first cylinder group is mechanically held at the intermediate lift position that is not in the valve stop (zero lift) state. Good restartability of the engine is ensured, and mechanical fail-safe is obtained.

According to a seventh aspect of the present invention, the second lift amount variable mechanism includes a second control unit that continuously and variably controls the lift amount of the intake valve of the second cylinder group by applying a driving force. In the case where the driving force is not applied, there is provided a stopper portion that restricts the lift amount of the intake valve to a position where the lift amount becomes a small lift.

  According to this invention, since the second lift amount variable mechanism that does not perform zero lift control mechanically holds the lift amount of the intake valve at a predetermined small lift by the stopper portion, including the effect of reducing the friction, Good restartability can be secured, and mechanical fail-safe can be obtained.

Embodiments of an internal combustion engine control apparatus according to the present invention will be described below in detail with reference to the drawings. In this embodiment, the present invention is applied to a gasoline V-type 6-cylinder internal combustion engine.
[First embodiment]
First, the overall configuration of the V-type internal combustion engine will be briefly described with reference to FIG. 1. The right cylinder RB of three cylinders, which is the first cylinder group, is a cylinder group that can be deactivated, while the second cylinder The three-cylinder left bank LB, which is a group, is a cylinder group that is always operating.

  The internal combustion engine is provided with two intake valves 4 and 4 and exhaust valves 5 and 5 for each cylinder, and the right bank RB is an intake side lift amount for variably controlling the valve lift amount of the intake valves 4 and 4. A first intake VEL1 that is a variable mechanism and a first intake VTC2 that is an intake-side lift phase variable mechanism that variably controls the opening and closing timing of the intake valves 4 and 4 are provided, and the valve lifts of the exhaust valves 5 and 5 are respectively provided. An exhaust VVL3 that is an exhaust-side lift variable mechanism that variably controls the amount is provided. The first intake VEL1, the first intake VTC2, and the exhaust VVL3 perform valve stop control of the intake valves 4, 4 and the exhaust valves 5, 5 in accordance with the engine operating state.

  On the other hand, the left bank LB includes a second intake VEL1 ′ and a second intake VTC2 ′ that variably control the valve lift amount and opening / closing timing of the intake valves 4 and 4, respectively, and the lift amount is fixed on the exhaust valves 5 and 5 side. This is a typical valve gear 8.

  Hereinafter, a specific configuration of the internal combustion engine will be described. A piston 01 provided in a cylinder bore formed in the cylinder block SB so as to be vertically slidable, an intake port 02 formed in the cylinder head SH, and The exhaust port 03 includes the intake valves 4 and 4 and the exhaust valves 5 and 5 that are slidably provided on the cylinder head SH and open and close the open ends of the suction and exhaust ports 02 and 03. .

  The piston 01 is connected to the crankshaft 04 via a connecting rod 05, and forms a combustion chamber 06 between the crown surface and the lower surface of the cylinder head SH.

  Inside the intake pipe 010 upstream of the intake manifold 09 for diverting intake air (air mixture) to each intake port 02, a throttle valve 07 for auxiliary control of the intake air amount is provided mainly for safety. ing. The cylinder head SH is provided with a fuel injection valve 08 that injects fuel directly into the combustion chamber 06.

  The first intake VEL1 of the right bank RB is configured to continuously control the valve lift amount of the intake valves 4 and 4 from zero lift (valve stop) to the maximum lift amount. The configuration is the same as that described in, for example, Japanese Patent Application Laid-Open No. 2004-76618 previously filed by the present applicant, and will be briefly described.

  As shown in FIG. 2, a hollow drive shaft 6 rotatably supported by a bearing on the upper part of the cylinder head SH, and a drive cam 7 which is an eccentric rotary cam fixed to the drive shaft 6 by press fitting or the like. The intake valves 4, 4 are slidably contacted with the upper surfaces of the valve lifters 8, 8 supported on the outer peripheral surface of the drive shaft 6 and arranged at the upper ends of the intake valves 4, 4. The two swing cams 9 and 9 to be transmitted, and a transmission mechanism linked between the drive cam 7 and the swing cams 9 and 9 for transmitting the rotational force of the drive cam 7 as the swing force of the swing cams 9 and 9. And.

  The drive shaft 6 receives a rotational force from the crankshaft of the engine, and the rotational direction is set in the clockwise direction (arrow direction) in FIG.

  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. 3 and 4 (both rear views), both the swing cams 9 have substantially the same raindrop shape and are integrally provided at both ends of the annular camshaft 10. In addition, the camshaft 10 is rotatably supported on the drive shaft 6 via the inner peripheral surface. A cam surface 9a is formed on the lower surface, 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 to the cam nose portion side, and from the ramp surface to the distal end side of the cam nose portion. A lift surface connected to the top surface of the maximum lift is formed, and the base circle surface, the ramp surface, and the lift surface are in contact with predetermined positions on the upper surface of each valve lifter 8 according to the swing position of the swing cam 9. It comes to touch.

  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 18 (described later) through 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 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 an electric motor 20 that is fixed to one end of a housing (not shown), and a ball screw transmission means 21 that is provided inside the housing and transmits the rotational driving force of the electric 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 will be described later, which is an ECU that detects the state of the vehicle including the engine operating state.

  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 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 is coupled to the drive shaft of the electric motor 20, and by this coupling, The rotational driving force of the electric motor 20 is transmitted to the ball screw shaft 23, and the ball screw shaft 23 is allowed to move slightly in the axial direction.

  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. A moving force in the axial direction is applied through each ball while converting the rotational motion of the ball screw shaft 23 into the linear motion of the ball nut 24. Further, as shown in FIG. 2, the ball nut 24 includes a first coil spring 27 elastically mounted on the tip end side of the ball screw shaft 23 and a second coil elastically mounted on the rear end side of the ball screw shaft 23. Both opposing spring forces with the spring 28 act. Therefore, even if the ball nut 24 stops at any position after a failure or a normal engine stop, the coil springs 27 and 28 hold the ball nut 24 at an intermediate position in the axial direction that is not zero lift, thereby starting the engine. Can be ensured, and since it is not the maximum lift control position, the friction of the valve operating device can be reduced and the cranking torque can be reduced, so that even better startability can be obtained.

  The controller 22 receives a crank angle signal, an engine speed signal, an accelerator opening signal, a vehicle speed signal, and a gear position signal from each sensor, as well as the rotation angle of the control shaft 17 (left and right banks first and second intake VEL1). Angle detection sensors 29a, 29a ′ for detecting the actual position of 1 ′, and angle detection sensors 29b for detecting the rotation angle of the drive shaft 6 (actual positions of the first and second intake VTCs 2, 2 ′ of the left and right banks), The current engine operating state is detected based on the input signal from 29b ', and the intake valve lift control current is output to the electric motors 20, 20'. In addition, a control current is output to a switching control valve 43 described later of the exhaust VVL, an electric motor 20 ′ described later of the second intake VEL1 ′, and an electromagnetic coil of the second intake VTC2 ′ of the exhaust VVL according to the engine operating state. ing.

  Hereinafter, an example of the operation of the first intake VEL1 will be briefly described. First, for example, at a light load such as an idling operation after the engine is started, the rotational torque transmitted to the electric motor 20 by a control signal from the controller 22 Is transmitted to the ball screw shaft 23 and rotated, the ball nut 24 moves linearly in one direction, whereby the control shaft 17 rotates in one direction via the link member 26 and the linkage arm 25.

  Therefore, as shown in FIGS. 3A and 3B, the control cam 18 rotates about the axis of the control shaft 17 with the same radius, and the thick portion moves away from the drive shaft 6 upward. As a result, the pivotal support point of the other end portion 11 b of the rocker arm 11 and the link rod 13 moves upward with respect to the drive shaft 6. Therefore, each swing cam 9 is connected to the cam nose portion 21 via the link rod 13. The side is forcibly pulled up and the whole is rotated clockwise.

  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 valve lift amount is transmitted to the swing cam 9 and the valve lifter 16 via the link rod 13. The lift amount is a small lift (Lr2) as shown on the left side of FIG.

  When the engine is in a low / medium load condition such as a steady operation, the rotational torque transmitted to the electric motor 20 by the control signal from the controller 22 is transmitted to the ball screw shaft 23 and rotated in the same direction. Further movement in the direction causes the control shaft 17 to further rotate in one direction via the link member 26 and the linkage arm 25.

  Therefore, as shown in FIGS. 3A and 3B, the control cam 18 has its axis rotated around the axis of the control shaft 17 with the same radius, and the thick portion moves further away from the drive shaft 6 in the upward direction. . As a result, the other end portion 11 b of the rocker arm 11 and the pivot point of the link rod 13 move further upward with respect to the drive shaft 6. Therefore, each swing cam 9 is connected to the cam nose portion via the link rod 13. The 21 side is pulled up and the whole further rotates clockwise.

  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 valve lift amount is transmitted to the swing cam 9 and the valve lifter 16 via the link rod 13. The lift amount becomes zero lift as shown in the upper right side of FIGS. 3A and 3B and FIG. As a result, the intake valves 4 and 4 are in a valve stop state. Here, the valve stop is a state in which no fuel is injected from the combustion injection valve 08 into the cylinder, and no output torque is generated without ignition of the spark plug. At this time, the exhaust valve is also controlled to zero lift by the exhaust VVL3.

  Further, when the low / medium load region is shifted to the high load region, the electric motor 20 is rotated in reverse by the control signal from the controller 22, and this rotation torque is transmitted to the ball screw shaft 23 and rotated. Accordingly, the ball nut 24 linearly moves in the opposite direction. Therefore, the control shaft 17 rotates the control cam 18 counterclockwise to rotate the center of the control cam downward as shown in FIGS. 4A and 4B as an example. For this reason, the entire rocker arm 11 moves in the direction toward the drive shaft 6 and the other end 11b presses the cam nose portion of the swing cam 9 downward via the link rod 13 so that the swing cam 9 The whole is rotated clockwise by a predetermined amount.

  Therefore, when the drive cam 7 rotates and pushes up the one end 11a of the rocker arm 11 via the link arm 12, the valve lift amount is transmitted to the swing cam 9 and the valve lifter 8 via the link rod 13. The lift amount increases rapidly from the zero lift to the maximum lift side.

  That is, the lift amount of the intake valves 4 and 4 is switched from L1 to zero lift from the light load (no load) region of the engine to the low / medium load, and further switched to the maximum lift side in the high load region. However, these series of controls are performed continuously.

  The throttle valve 07 is normally maintained in a substantially fully opened state, and the required air amount to the combustion chamber 06 is mainly controlled by the first intake air VEL1 and the second intake air VEL1 '.

  As the first intake VTC 2, a known one using a hysteresis brake described in Japanese Patent Application Laid-Open No. 2004-156508, for example, previously filed by the present applicant is used.

  This is because an assembly angle changing means for changing the assembly angle between the drive ring on the crankshaft side and the driven shaft member on the drive shaft 6 side is interposed, and the assembly change means is operated in the engine. The controller 22 outputs a control current to the electromagnetic coil according to the state of the vehicle including the state to control the operation of the hysteresis brake to advance the lift phase of the intake valves 4, 4, that is, the opening / closing timing of the intake valves 4, 4. It is designed to control the angle or retard side. The VTC may be a vane type that converts the phase by hydraulic pressure instead of the hysteresis brake.

  The exhaust VVL3 of the right bank RB has the same structure as that described in, for example, Japanese Patent Laid-Open No. 10-8935, and briefly described, as shown in FIGS. 5 and 6, the exhaust camshaft 30 A maximum lift high-speed cam 31 provided for each cylinder, and provided on both sides of the high-speed cam 31 and supported by a zero-lift cylindrical cams 32 and 32 and a rocker shaft 33 so as to be swingable. An integral main rocker arm 34 which is disposed at a position corresponding to both the cylindrical cams 32, 32, and whose lower end of each tip is in contact with the stem ends of the exhaust valves 5, 5, and a position corresponding to the high speed cam 31. A sub-rocker arm 35 capable of lost motion, a lost motion mechanism 36 provided at a lower portion of the sub-rocker arm 35, and a support shaft 37 fixed to the main rocker arm 34. A lever member 38 that is supported so as to be swingable and engages / disengages with the lower end portion of the sub rocker arm 35 to synchronize or cancel the sub rocker arm 35 and the main rocker arm 34, and the lever member 38. A hydraulic plunger 39 and a return spring 40 that are engaged and disengaged are provided.

  The hydraulic plunger 39 is moved backward by an oil pressure supplied from an oil pump 46 to an oil chamber 41 formed on the outer peripheral side via hydraulic passages 41a and 41b formed in the rocker shaft 32 and the sub rocker arm. It moves forward by the spring force of the coil spring 42 mounted inside. Further, the electromagnetic switching control valve 43 switches the conduction between the hydraulic passages 41 a and 41 b and the drain passage 44 or the discharge hydraulic pressure of the oil pump 46. The switching control valve 43 is switched by the control current output from the controller 22.

  Hereinafter, the operation of the exhaust VVL 3 will be briefly described. First, when the engine 22 is in a light load (no load) state such as an idling operation after the engine is started, the controller 22 that has detected it is directed to the switching control valve 43. Therefore, the hydraulic passages 41a and 41b are connected to the drain passage 44, so that the hydraulic pressure is reduced.

  Therefore, as shown in FIG. 6B, the hydraulic plunger 39 moves forward by the spring force of the coil spring 42, and rotates the lever member 38 counterclockwise against the spring force of the return spring 40. The tip of the member 38 engages with the lower jaw on the tip side of the sub rocker arm 35 during the base circle of the high-speed cam 31 to interlock the sub rocker arm 35 and the main rocker arm 34.

  As a result, the main rocker arm 34 swings in accordance with the cam profile of the high-speed cam 31, so that the exhaust valves 5 and 5 are controlled to be switched to the maximum lift (see the upper left side in FIG. 8).

  On the other hand, when the engine is shifted to the low / medium load region, the switching control valve 43 is actuated to supply the hydraulic pressure discharged from the oil pump 46 into the oil chamber 41, and the hydraulic plunger 39 is turned into the coil spring 42 as shown in FIG. It moves backward against the spring force. As a result, the lever member 38 is rotated in the opposite direction by the spring force of the return spring 40 to release the connection between the sub rocker arm 35 and the main rocker arm 34, whereby the sub rocker arm 35 enters a lost motion state. For this reason, the main rocker arm 34 does not receive the lift force of the high-speed cam 31 and only comes into sliding contact with the cylindrical cams 30 and 30, and the lift amount of the exhaust valves 5 and 5 becomes zero lift. As a result, the valve is stopped, and the intake valves 4 and 4 are also in the valve stopped state, so that the cylinder is stopped (see the upper right side in FIG. 8).

  On the other hand, when shifting from the low / medium load region to the high load region, the controller 22 cuts off the energization to the switching control valve 43, so that the hydraulic passages 41 a and 41 b and the drain passage 44 are communicated with each other. 6B, the hydraulic plunger 39 moves forward by the spring force of the coil spring 42, whereby the lever member 38 rotates against the spring force of the return spring 40. During the base circle of the cam 31, the front end portion engages with the lower end jaw portion on the front end side of the sub rocker arm 35, and the sub rocker arm 35 and the main rocker arm 34 are interlocked.

  Therefore, the main rocker arm 34 swings according to the cam profile of the high-speed cam 31 as in the idling operation range, and the exhaust valves 5 and 5 are controlled to be switched to the maximum lift. As described above, the exhaust VVL 3 is configured to switch the lift amount of the exhaust valves 5 and 5 on and off between zero lift and high lift.

  As shown in FIG. 7, the structure of the second intake VEL1 'on the left bank LB side is basically the same as that of the first intake VEL1 on the right bank RB side. Although the detailed description is omitted, the difference is that the second coil spring for urging the ball nut 24 ′ to the large lift control side via the control shaft 17 ′ is eliminated, and the small lift Only the first coil spring 27 'biased toward the control side is used. Furthermore, a stopper mechanism 45 that restricts the rotation of the control shaft 17 ′ to a predetermined small lift that does not become zero lift on the small lift side biased by the first coil spring 27 ′ is provided. A stopper shaft 45a provided on one end surface of the control shaft 17 'and a stopper wall (not shown) which is provided on a rocker cover or the like and which comes into contact with the stopper shaft 45a to restrict the rotation of the control shaft 17 are provided. ing.

  Therefore, the second intake VEL1 ′ can not only perform the zero lift control, but can obtain the same operation and effect as the first intake VEL1, and at the time of engine stop, the stopper mechanism 45 causes the intake valve 4, via the control shaft 17 ′. Since 4 is regulated to a predetermined small lift, the fail-safe function works to ensure engine startability. That is, the intake air amount necessary for starting is ensured, and a low friction effect by a small lift is obtained.

  Since the structure of the second intake VTC 2 ′ is the same as that of the first intake VTC 2, a detailed description thereof will be omitted. In addition, the opening and closing timing of the intake valves 4 and 4 is also controlled to advance or retard.

  Hereinafter, a specific operation of the present embodiment will be described.

  First, as shown in FIG. 9, in the light load / low rotation region A including the idling operation, all the intake / exhaust valves 4 and 5 are operating in the cylinder groups of the left and right banks LB and RB ( All cylinder operation). Here, the operation means a state where fuel injection is performed and output torque is generated by combustion.

  Further, in the low / medium load region B, the cylinder group of the right bank RB is stopped (partially stopped), and only the cylinder group of the left bank LB is operated.

  In the region C where the load is high or the rotation is high from the region B, the entire cylinder operation state is again controlled to generate the engine torque.

  9A to 10C show changes in the lift amount (L) when the engine speed and load are gently increased (swept) as indicated by the arrows in FIG. That is, FIG. 10a shows the change in the valve lift amount Lr of the intake valves 4, 4 in the right bank RB indicated by the broken line and the valve lift amount Ll of the intake valves 4, 4 in the left bank LB indicated by the solid line in the sweep of FIG. Show. FIG. 10b shows the internal pressure of the intake manifold 09 downstream of the throttle valve 07. During the sweep, the opening of the throttle valve 07 is maintained at a substantially large opening, so that the pressure is close to the atmospheric pressure. The engine torque is not controlled by the throttle valve 07 but mainly by the valve lift amount of the intake valves 4 and 4. FIG. 10c shows the change of the intake valve closing timing (IVC) when sweeping. Here, the IVC timing T in the right bank indicated by the broken line sticks to the lowest position, indicating that the intake valves 4 and 4 are not opened and closed with zero lift.

  Here, the vicinity of the boundary that changes from the all-cylinder operation region A to the reduced-cylinder operation region B will be considered. As can be seen from FIG. 10a, on the side of the region A immediately before the boundary, the right bank RB side is the lift Lr2 and the left bank LB side is the lift Ll2, both of which are substantially the same value. As a result, the variation in the intake air amount between the banks RB and LB can be reduced, the variation in the friction of the valve gear can be reduced, the combustion fluctuation for each cylinder including both banks can be suppressed, and the engine rotation can be stabilized. It becomes possible to make it.

  Further, as can be seen in FIG. 10c, also regarding the intake valve closing timing (IVC), Tr2 of the right bank RB and Tl2 of the left bank LB are substantially the same value, and this further causes variation in the intake air amount for each cylinder. The engine rotation can be stabilized.

  In addition, since the intake air amount control is performed by reducing the valve lift and advancing the intake valve closing timing with respect to the bottom dead center, the throttle valve 07 can be opened to a large degree as described above. As a result, since the internal pressure of the intake manifold 09 can be increased to near atmospheric pressure, the pumping loss can be reduced and the fuel efficiency can be improved even in the region A where the reduced cylinder operation is not performed.

  Next, considering the region B immediately after the boundary, that is, after the shift to the reduced cylinder operation, the lift Lr3 of the intake valves 4 and 4 on the right bank RB side becomes 0 (valve stop), and the valve lift of the exhaust valves 5 and 5 The amount also becomes 0 (valve stop) by the exhaust VVL3, and the cylinder is deactivated. On the other hand, the valve lift L13 of the intake valves 4 and 4 on the left bank LB side increases conversely.

  The lower stage (middle stage) of FIG. 8 is a diagram comparing valve lift curves for the lift Ll2 in the region A and the lift Ll3 in the region B of the left bank LB, and the lift Ll3 is larger than the lift Ll2, Tl3 at the intake valve closing timing (IVC) is retarded to the bottom dead center (TDC) side than Tl2. Therefore, even if the region A is shifted to B and the number of operating cylinders is halved, the intake valve lift amount of the operating cylinders increases, and the intake air amount is increased by bringing the IVC closer to the bottom dead center. The total amount of intake air can be at the same level. Therefore, the torque change at the time of shifting from the region A to B can be suppressed.

  In addition, as shown in the lowermost stage of FIG. 8, switching is performed while maintaining a large opening of the throttle valve 07, so that the internal pressure of the intake manifold 09 is maintained near atmospheric pressure before and after switching. In other words, transient torque fluctuations caused by large changes in intake manifold internal pressure during switching transitions are unlikely to occur. Therefore, torque shock at the time of switching can be effectively suppressed.

  In addition, as shown in the middle of FIG. 8, the intake valve opening timing (IVO) on the left bank LB side is substantially the same between the lift Ll2 and the lift Ll3, so that the change in valve overlap before and after switching is suppressed. Therefore, the change in the amount of residual gas taken into the combustion chamber 06 during that time can be suppressed, and the torque difference before and after switching can be reduced.

  By the way, when the valve lifts (L2, L3) of the intake valves 4, 4 in the left bank LB shown in FIG. 8 are sufficiently small, the flow velocity at which fresh air is taken into the combustion chamber 06 from the gap between the intake valves 4, 4 is It becomes faster to the sound speed level. Then, a so-called choking phenomenon occurs, and the amount of intake air into the combustion chamber 06 should be determined approximately by the lift time area S.

  Here, if the time area S13 is about twice that of S12, the occurrence of a torque difference can be prevented even when the choking occurs. In other words, the region A is shifted to B, and the number of active cylinders is halved, but the lift time area S of the active cylinders has doubled, so that the total lift time area when viewed with all cylinders is almost equal. means. That is, S13 = Sl2 × total number of cylinders / number of operating cylinders. Here, Lr2 and Ll2 during the operation of all cylinders are the same, and Sr2 is also substantially the same as Sl2.

  In the reduced-cylinder operation, the number of operating cylinders is reduced, so that the surface area in the cylinder is reduced, thereby reducing the cooling loss and improving the fuel efficiency. In addition to this effect, the above-mentioned pumping loss reduction effect can further improve the fuel efficiency. . In addition, since the IVC is before the bottom dead center, the effective compression ratio is lowered and the compression top dead center temperature is lowered. Therefore, the cooling loss is further reduced, and the fuel efficiency can be sufficiently improved.

  Next, when the engine is further swept in the reduced-cylinder operation region B and the load and rotation of the engine are increased, the entire cylinder operation is switched again.

  That is, the right bank RB is in the cylinder deactivation state in the region B, that is, the lift amount of the intake valves 4 and 4 is 0 (Lr4) and the lift amount of the exhaust valves 5 and 5 is also 0. It switches to the middle lift Lr5, and the lift amount of the exhaust valves 5 and 5 is also operated with a normal lift (maximum lift). This normal lift is the same as the fixed lift amount of the exhaust valves 5 and 5 of the normally operating cylinder group of the left bank LB.

  On the other hand, the lift amount of the intake valves 4 and 4 in the left bank LB decreases from a slightly large lift state (Ll4) to Ll5.

  Therefore, the total intake air amount when viewed in all cylinders can be suppressed at an equivalent level, and the occurrence of a torque difference can be suppressed.

  Further, the lift amount Ll5 of the left bank LB and the lift amount Lr5 of the right bank RB in the region C are equal to each other, and as shown in FIG. 10c, the intake valves 4 of the left and right banks LB, RB. , 4 at the closing timing (IVC), Tl5 and Tr5 have substantially the same timing. Therefore, the occurrence of intake swirl is suppressed and high torque is obtained.

  Next, in the region C, when the engine load / rotation further increases, the valve lift amounts Ll and Lr of the intake valves 4 and 4 of the left and right banks LB and RB maintain the substantially same lift amount and the lift amount increases. The intake valve closing timing and Tl, Tr of both of them are retarded while maintaining substantially the same timing, and a desired increase in torque can be obtained.

  In addition, since the throttle valve 07 is maintained in a large opening state (intake pipe internal pressure atmospheric pressure level) over the entire sweep, the pumping loss can be reduced, and fuel consumption and torque can be improved.

In addition, the throttle valve 07 is maintained in a large opening state at the time of switching from the region A to B and from the region B to C, so that the valve lift characteristic is greatly switched and changed. It becomes possible to prevent a transient torque shock caused by a transient change in the internal pressure of the intake manifold 09.
[Second Embodiment]
FIGS. 11 to 13 show the control characteristics of the control device of the second embodiment. First, as shown in the control map of FIG. 11, in the region A of all cylinder operation during light load operation including idling operation, the left and right banks Lean burn control is performed in which the air-fuel mixture injected from the LB and RB fuel injection valves 08 into the combustion chambers 06 is diluted more than the stoichiometric air fuel ratio. With such lean burn control, the torque for the same intake air volume decreases, so that the lift amount is increased in order to obtain the target torque.

  Here, considering the vicinity of the boundary where the swept region A changes to B, the lift amount of the intake valves 4 and 4 in both banks LB and RB is shown in FIG. As shown, Lr2 ′ and Ll2 ′ are slightly larger than Lr2 and Ll2 in the first embodiment. Therefore, it is possible to avoid instability of engine rotation caused by variations in lift between cylinders, which becomes a problem when controlling to a small lift with a light load. This is because, since the control lift becomes larger than normal (first embodiment), even if the lift variation amount is the same, the lift variation ratio decreases because the base lift becomes high. Here, in the case of lean burn, NOx is likely to be generated, but since it is a light load operation, there is no actual harm.

  On the other hand, in the region B shown in FIG. 11, the injected air-fuel mixture remains at the stoichiometric air-fuel ratio. In this case, the lift Ll3 of the intake valves 4 and 4 does not increase as compared with the normal (first embodiment), so that the response speed for switching from Ll2 'to Ll3 is increased, so that the control accuracy can be improved.

  Further, in the region B during the reduced cylinder operation, on the high load side (BB) beyond the D line, the opening / closing timing phase of the intake valves 4, 4 is advanced by the second intake VTC 2 'on the left bank LB side. Control to the side. This creates a valve overlap with the exhaust valves 5 and 5 to increase the internal EGR (see FIG. 13). For this reason, it is possible to improve the fuel efficiency by improving the specific heat ratio of the air-fuel mixture.

  Further, in the reduced-cylinder operation, the cylinder pressure (load) of the operating cylinder on the left bank LB side increases, so that NOx is easily generated in principle. On the other hand, in this embodiment, since the internal EGR by the valve overlap is used instead of the external EGR by the EGR valve, the EGR passage is clogged like the external EGR even if EGR is applied to a large amount of exhaust gas. There is no problem. This large amount of internal EGR can reduce NOx.

  Further, since the exhaust gas is recirculated by the negative pressure in the intake manifold 09 in the external EGR, it is difficult to EGR a large amount of exhaust gas in a state where the intake manifold internal pressure is close to atmospheric pressure as in the reduced cylinder operation. Although there is a fear, EGR is possible even if the intake manifold internal pressure is high when the valve overlap expansion control is performed by the advance angle control of the second intake VTC 2 '. Therefore, since a large amount of EGR is possible, it is possible to avoid an increase in NOx, which is a problem during reduced-cylinder operation.

  If the D line is supplemented based on FIG. 12, the lift Lla of the intake valves 4 and 4 of the left bank LB in the region not exceeding the D line (FIG. 11B-a) is the stoichiometric air-fuel ratio, and the D line The lift Llb exceeding the value indicates that the internal EGR is increased (advance angle control by the second intake VTC 2 ′).

  Here, Lla <Llb is for correcting the torque drop due to the increase in internal EGR and for increasing the operating angle (valve opening period) in order to advance the intake valve opening timing (IVO).

  Note that, on the low load side of the region B, the lean burn control may be used instead of the stoichiometric air-fuel ratio. In this case, the fuel consumption is further improved. However, since the lean burn control is performed, NOx may increase, but since the engine load is small, it can be suppressed to a predetermined low level.

Further, in each of the regions A to C, the throttle valve 07 is almost fully opened as shown in the lowermost stage of FIG. 13, so that the internal pressure of the intake manifold 09 is close to the atmospheric pressure, and the pumping loss can be reduced. Thus, the same effects as those of the first embodiment can be obtained.
[Third embodiment]
14 and 15 show the control characteristics of the control device of the third embodiment. In this embodiment, as shown in the control map of FIG. 14, switching between valve stop and lift operation in the intake valves 4 and 4 of the right bank RB. Equivalent hysteresis is provided on both the boundary line and the switching boundary line of the lift sudden change switching line of the intake valves 4 and 4 of the left bank LB.

  That is, with respect to the switching line E on the engine load / rotation increasing side between the regions A and B, the switching line Ed in the load / rotation decreasing direction is slightly set in the load / rotation decreasing direction. That is, a hysteresis of ΔE is provided. Accordingly, the switching operation of the first and second intake air VEL1, 1 ′, the first and second intake air VTC2, 2 ′, and the first exhaust VVL3 in both the right bank RB and the left bank LB in the vicinity of the boundary line E is performed. It is possible to prevent the occurrence of hunting at the time.

  Similarly, with respect to the switching line F on the load / rotation increasing side in the regions B and C, the switching line Fd in the load / rotation decreasing direction is slightly set in the load / rotation decreasing direction. That is, a hysteresis of ΔF is provided.

  Accordingly, the first and second intake air VEL1, 1 ′, the first and second intake air VTC2, 2 ′, and the first exhaust gas in both the right bank RB and the left bank LB in the vicinity of the boundary line F as described above. It becomes possible to prevent the occurrence of hunting at the time of switching operation of VVL3.

  In FIG. 15, between the regions A and B, the hysteresis between the lifts Lr3 and Lr3d of the intake valves 4 and 4 on the right bank RB side is ΔE, and similarly between Lr2 and Lr2d, Ll3 on the left bank LB side Similarly, the hysteresis ΔE is between L1 and Ll3d and between Ll2 and Ll2d.

  In addition, between regions B and C, the hysteresis between Lr4 and Lr4d is ΔF, and similarly between Lr5 and Lr5d, between Ll5 and Ll5d, and between Ll4 and Ll4d is also ΔF. ing.

  Here, since ΔE is a normal range, it is slightly larger than ΔF, and unpleasant hunting is less likely to occur in the normal range.

  On the other hand, ΔF is on the high rotation side and the high load side, and since the occurrence frequency of hunting is low, it is made smaller than ΔE in order to improve the switching response.

  Also in this embodiment, the throttle valve 07 is almost fully opened in each operation region, so that the pumping loss can be reduced, thereby improving the fuel efficiency and the same effects as in the first embodiment. Is obtained.

  The present invention is not limited to the configuration of each of the above embodiments. For example, as an internal combustion engine, the present invention can be applied to a series type other than the V type, and is not a spark ignition by a spark plug but a compression ignition. It can also be applied to engines.

  Instead of the valve stop mechanism VVL3 on the exhaust side, a continuously variable lift mechanism capable of changing the lift to zero lift similar to the VEL1 on the intake side may be used. Further, the specific mechanism of VEL1 and VEL1 'is not particularly limited, and may be a continuously variable lift mechanism of another mechanism as described in JP-A-2006-200391, for example.

1 is a schematic view of an internal combustion engine according to a first embodiment of the present invention. It is a perspective view which shows the 1st intake VEL and 1st intake VTC which are provided to a present Example. A and B are explanatory diagrams of operation at the time of zero lift control by the first intake VEL. A and B are operation explanatory diagrams at the time of large lift control by the intake VEL. It is a perspective view which shows the exhaust VVL provided to a present Example. A is an operation explanatory diagram at the time of zero lift control by the exhaust VVL, and B is an operation explanatory diagram at the time of high lift control. It is a perspective view which shows the 2nd intake VEL and 2nd intake VTC which are provided to a present Example. It is a valve lift characteristic figure of the intake / exhaust valve at the time of all cylinder operation at the time of light load in this example, and at the time of reduced cylinder operation at the time of low / medium load. It is a control map which shows the relationship between the engine speed and engine torque by the controller provided for a present Example. a is an intake valve lift amount characteristic diagram during all cylinders and reduced cylinder operation in this embodiment, b is a throttle valve opening explanatory diagram during each operation, and c is an explanatory diagram showing the closing timing of the intake valve during each operation. FIG. It is a control map which shows the relationship between the engine speed and the engine torque by the controller provided for 2nd Example. a is an intake valve lift amount characteristic diagram at the time of all-cylinder and reduced-cylinder operation in this embodiment, and b is an explanatory view of the throttle valve opening during each operation. It is a valve lift characteristic figure of the intake / exhaust valve at the time of all cylinder operation at the time of light load in this example, and at the time of reduced cylinder operation at the time of low / medium load. It is a control map which shows the relationship between the engine speed and the engine torque by the controller provided for 3rd Example. All cylinders in this embodiment, intake valve lift amount characteristic diagram during reduced cylinder operation,

Explanation of symbols

RB ... Right bank (first cylinder group)
LB ... Left bank (second cylinder group)
1... First intake VEL (first control mechanism / first lift variable mechanism)
2 ... 1st intake VTC
1 ′... Second intake VEL (second control mechanism / second lift variable mechanism)
2 '... 2nd intake VTC
3. Exhaust VVL (first control mechanism / first lift variable mechanism)
DESCRIPTION OF SYMBOLS 4 ... Intake valve 5 ... Exhaust valve 6 ... Drive shaft 7 ... Drive cam 9 ... Swing cam 11 ... Rocker arm 12 ... Link arm 13 ... Link rod 17 ... Control shaft 18 ... Control cam 19 ... Drive mechanism 20 ... Electric motor 22 ... Controller 31 ... High speed cam 32 ... Cylindrical cam 33 ... Rocker shaft 34 ... Main rocker arm 35 ... Sub rocker arm

Claims (7)

A control device for an internal combustion engine comprising a first cylinder group in which an intake valve and an exhaust valve are controlled to stop, and a second cylinder group in which the intake valve is variably controlled in lift amount,
When performing valve stop control of the intake valve and exhaust valve of the first cylinder group, the valve overlap of the intake valve and exhaust valve of the second cylinder group is substantially the same as the valve overlap before the valve stop. And the second cylinder group is controlled to increase the lift amount while controlling the closing timing of the intake valve of the second cylinder group to the retard side than before the valve stop. control apparatus for an internal combustion engine, characterized in that the engine torque is configured to be substantially equivalent to the engine torque by the first cylinder group and second cylinder group before the valve stop. The control apparatus for an internal combustion engine according to claim 1,
The control apparatus for an internal combustion engine, wherein the valve stop control is performed when the engine speed increases or when the engine load increases . The control apparatus for an internal combustion engine according to claim 1 or 2,
The first cylinder group and the second cylinder group have the same number of cylinders, and the lift amount of the intake valve and the exhaust valve of the first cylinder group becomes substantially zero by the valve stop control of the first cylinder group. On the other hand, the time area of the lift amount of the intake valve of the second cylinder group is equal to the time area of the lift amount of the intake valve of the first cylinder group before the valve stop control and the intake valve of the second cylinder group. A control apparatus for an internal combustion engine, which is controlled to be substantially the same as a sum of a lift amount and a time area . The control apparatus for an internal combustion engine according to any one of claims 1 to 3,
A control device for an internal combustion engine, wherein the control is switched to all-cylinder operation without valve stop in a light load operation state of the engine. The control apparatus for an internal combustion engine according to claim 4 ,
A control apparatus for an internal combustion engine , wherein the air-fuel ratio is controlled to a lean side in a light load operation state of the engine. In the control device for an internal combustion engine according to any one of claims 1 to 5 ,
The first lift amount variable mechanism that stops the valve by setting the lift amount of the intake valve and the exhaust valve of the first cylinder group to substantially zero is a first control that continuously varies the lift amount of the intake valve by applying a driving force. And a control device for an internal combustion engine , wherein when the driving force is not applied, the lift amount of the intake valve is held at an intermediate lift position that is not mechanically zero lift . The control apparatus for an internal combustion engine according to claim 6 ,
The second lift amount variable mechanism that continuously variably controls the lift amount of the intake valve of the second cylinder group that is always operated continuously applies the drive force to the lift amount of the intake valve of the second cylinder group. And a second control unit that variably controls the control unit and a stopper unit that restricts the lift amount of the intake valve to a predetermined small lift when the driving force is not applied. apparatus.

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