JP2010185400A - Control device of internal combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a control device of an internal combustion engine, inhibiting discharge of unburned fuel by generating a swirl flow of proper strength in a cylinder. <P>SOLUTION: The internal combustion engine is equipped with a valve gear varying a valve opening characteristic of an intake valve by varying a lift amount of the intake valve. The control device of the internal combustion engine includes a detection means (S102) detecting the strength of the swirl flow in the cylinder from an operation angle of the intake valve and engine rotation speed, a target value set means (S103) setting a target value of the strength of the swirl flow, and a variation means (S106) varying the strength of the swirl flow detected by the detection means close to the target value set by the target value set means by varying the operation angle of the intake valve. <P>COPYRIGHT: (C)2010,JPO&INPIT

Description

  The present invention relates to a control device for an internal combustion engine.

  A technique is known in which the strength of gas flow near the intake valve and the strength of gas flow due to piston behavior are estimated from the opening / closing timing of the intake valve and the lift amount, and this is reflected in fuel injection control (for example, patents) Reference 1). However, it is difficult to accurately estimate the strength of the gas flow because the change in the intake flow rate passing through the intake valve cannot be captured.

  Here, when the internal combustion engine is under a low load, the swirl ratio is lowered by delaying the opening timing of the intake valve to suppress fuel diffusion, and the effective compression ratio is increased by increasing the closing timing of the intake valve. Yes. In this case, the opening timing of the intake valve is brought close to the intake stroke top dead center, and the closing timing of the intake valve is brought close to the intake stroke bottom dead center.

  Thus, when the operating angle of the intake valve is reduced, the lift amount of the intake valve is reduced. Therefore, the flow rate of the intake air when passing through the intake valve increases. Thereby, a swirl ratio will become high. That is, although the intake valve opening timing is delayed in an attempt to lower the swirl ratio, the intake air flow rate increases so as to cancel this. Therefore, the effect of lowering the swirl ratio and the effect of increasing the effective compression ratio are weakened, and unburned fuel may be discharged.

JP 2007-327345 A JP 2006-283659 A JP 2008-255815 A

  The present invention has been made in view of the above-described problems, and in a control device for an internal combustion engine, a technique for suppressing discharge of unburned fuel by generating a moderately strong swirling flow in a cylinder. The purpose is to provide.

In order to achieve the above object, an internal combustion engine control apparatus according to the present invention employs the following means. That is, the control device for an internal combustion engine according to the present invention provides:
In a control device for an internal combustion engine provided with a valve gear that changes a valve opening characteristic of the intake valve by changing a lift amount of the intake valve,
Detecting means for detecting the strength of the swirling flow in the cylinder from the operating angle of the intake valve and the engine speed;
Target value setting means for setting a target value of the strength of the swirling flow;
Changing means for changing the working angle of the intake valve to bring the strength of the swirling flow detected by the detecting means closer to the target value set by the target value setting means;
It is characterized by providing.

In such an internal combustion engine, the operating angle increases as the lift amount of the intake valve increases, and the operating angle decreases as the lift amount decreases. Here, there is a correlation between the operating angle of the intake valve, the engine speed, and the strength of the swirling flow in the cylinder. For this reason, the strength of the swirling flow can be detected by storing in advance the relationship between the operating angle of the intake valve, the engine speed, and the strength of the swirling flow. The strength of the swirling flow may be an angular velocity of intake air per unit time, a swirl ratio, or a flow velocity of intake air in the cylinder. The actual strength of the swirling flow is the decrease in the strength of the swirling flow by delaying the opening timing of the intake valve, and the increase in the strength of the swirling flow by increasing the flow velocity of the intake air by decreasing the lift amount. The value is the sum of minutes and minutes. That is, the detection means estimates the final value of the combined increase and decrease of the strength of the swirling flow.

  If the operating angle of the intake valve is controlled so that the strength of the swirling flow approaches the target value, the strength of the swirling flow can be made more appropriate, and the amount of unburned fuel discharged can be reduced. This can be either feedback control or feedforward control.

  Further, in the present invention, when the strength of the swirl flow in the cylinder is reduced at low load, the operating angle is increased while increasing the amount of change in the closing timing rather than the amount of change in the opening timing of the intake valve. By doing so, the strength of the swirling flow can be reduced.

  Here, since the lift amount increases by increasing the operating angle of the intake valve, the strength of the swirling flow in the cylinder decreases. Since the influence of the strength of the swirling flow is large at the time of low load, the discharge amount of unburned fuel can be reduced by reducing the strength of the swirling flow at the time of low load. At this time, if the opening timing of the intake valve is advanced, the strength of the swirl flow increases due to this influence. Therefore, the amount of change in the opening timing of the intake valve is reduced. The opening timing of the intake valve need not be changed. On the other hand, even if the closing timing of the intake valve is delayed, the influence on the strength of the swirling flow and the effective compression ratio is small. Therefore, the operating angle is changed by increasing the amount of change in the closing timing of the intake valve. Thereby, discharge | emission of unburned fuel can be suppressed.

  According to the control apparatus for an internal combustion engine according to the present invention, it is possible to suppress the discharge of unburned fuel by generating a swirling flow having an appropriate strength in the cylinder.

1 is a diagram for explaining a system configuration of Embodiment 1. FIG. FIG. 2 is a perspective view for explaining a configuration of a variable valve operating apparatus in the system shown in FIG. 1. It is a figure for demonstrating the structure of the variable valve mechanism in the variable valve apparatus shown in FIG. It is the figure which showed the lift amount of the intake valve with respect to a crank angle. It is the figure which showed the relationship between the opening time of an intake valve, and a swirl ratio. It is the figure which showed the relationship between the closing timing of an intake valve, and an effective compression ratio. It is the figure which showed the relationship between the working angle of an intake valve, and a swirl ratio. 3 is a flowchart illustrating a control flow of swirl at a low load according to the first embodiment.

  Hereinafter, specific embodiments of a control device for an internal combustion engine according to the present invention will be described with reference to the drawings.

[System configuration]
FIG. 1 is a diagram for explaining the system configuration of the first embodiment. The system according to the first embodiment includes an internal combustion engine 1. The internal combustion engine 1 has a plurality of cylinders 2. FIG. 1 shows only one cylinder among a plurality of cylinders.

  The internal combustion engine 1 includes a cylinder block 4 having a piston 3 therein. The piston 3 is connected to the crankshaft 5 via a crank mechanism. A crank angle sensor 6 is provided in the vicinity of the crankshaft 5. The crank angle sensor 6 is configured to detect the rotation angle of the crankshaft 5 (that is, the crank angle).

  A cylinder head 8 is assembled to the upper part of the cylinder block 4. The cylinder head 8 includes an intake port 12 that communicates with the cylinder 2. An intake passage 26 is connected to the cylinder head 8. The intake passage 26 is provided with a throttle 27. The intake port 12 communicates the intake passage 26 and the cylinder 2. An intake valve 14 is provided at a connection portion between the intake port 12 and the cylinder 2. The system of the first embodiment includes a plurality of intake valves 14 corresponding to a plurality of intake ports 12 provided for each cylinder 2. FIG. 1 shows one intake port 12 and one intake valve 14. A variable valve gear 18 is provided between the intake valve 14 and the intake cam 16 provided on the intake camshaft 15. The variable valve operating device 18 is configured such that the valve opening characteristic of the intake valve 14 can be mechanically changed. The details of the variable valve operating device 18 will be described later.

  An intake pulley 31 is provided at the end of the intake camshaft 15. Furthermore, a variable rotation phase mechanism (hereinafter referred to as “intake side VVT”) 32 that can change the relative rotation phase between the intake camshaft 15 and the intake side pulley 31 is provided. The intake side VVT 32 controls the relative rotation phase between the intake camshaft 15 and the intake side pulley 31 in accordance with a command from the ECU 60 described later. The rotational drive of the intake camshaft 15 is performed by the driving force of the crankshaft 5.

  Further, the cylinder head 8 includes an exhaust port 28 that communicates with the cylinder 2. An exhaust valve 29 is provided at a connection portion between the exhaust port 28 and the cylinder 2.

  In addition, the system of this embodiment includes an ECU 60 as an electronic control device. To the output side of the ECU 60, the variable valve device 18, the throttle 27, and the intake side VVT 32 are connected. In addition to the crank angle sensor 6, an accelerator opening sensor 24 that outputs an electrical signal corresponding to the amount by which the driver has depressed the accelerator pedal 23 is connected to the input side of the ECU 60. The ECU 60 executes overall control of the internal combustion engine 1 such as fuel injection control and ignition timing control based on the output of each sensor.

  Further, the ECU 60 calculates the engine speed based on the output of the crank angle sensor 6. Further, the ECU 60 calculates the engine load based on the output of the accelerator opening sensor 24.

[Configuration of variable valve gear]
FIG. 2 is a perspective view for explaining the configuration of the variable valve operating device 18 in the system shown in FIG.

As shown in FIG. 2, the intake camshaft 15 is provided with two intake cams 16 and 17 per cylinder. The two intake valves 14L and 14R are arranged symmetrically about the first intake cam 16 as the main cam. Between the first intake cam 16 and the intake valves 14L and 14R, variable valve mechanisms 40L and 40R are provided, respectively, which link the lift movement of the intake valves 14L and 14R with the rotational movement of the first intake cam 16. Yes. On the other hand, the second intake cam 17 is arranged so as to sandwich the second intake valve 14R between the first intake cam 16 and the second intake cam 16. A fixed valve mechanism 70 is provided between the second intake cam 17 and the second intake valve 14R to link the lift movement of the second intake valve 14R with the rotational movement of the second intake cam 17. The variable valve operating device 18 is configured to be able to selectively switch the interlocking destination of the second intake valve 14R in conjunction with the lift between the variable valve mechanism 40R and the fixed valve mechanism 70. In the present embodiment, the description of the fixed valve mechanism 70 is omitted.

  FIG. 3 is a view for explaining the configuration of the variable valve mechanism 40 in the variable valve apparatus 18 shown in FIG. Specifically, FIG. 3 is a view of the variable valve mechanism 40 as viewed from the axial direction of the intake camshaft 15. The left and right variable valve mechanisms 40L and 40R are basically symmetrical with respect to the first intake cam 16, and therefore the left and right variable valve mechanisms 40L and 40R are not distinguished here. Will be explained. In the present specification and drawings, when the left and right variable valve mechanisms 40L and 40R are not distinguished, they are simply referred to as the variable valve mechanism 40. Similarly, the component parts of the variable valve mechanisms 40L, 40R and the symmetrically arranged parts such as the intake valves 14L, 14R are distinguished from each other except when it is necessary to distinguish between them. The symbol is not attached.

  As shown in FIG. 3, the rocker arm 35 is supported by the intake valve 14. The variable valve mechanism 40 is interposed between the first intake cam 16 and the rocker arm 35. The variable valve mechanism 40 is configured to continuously change the interlocking state between the rotational motion of the first intake cam 16 and the rocking motion of the rocker arm 35.

  The variable valve mechanism 40 has a control shaft 41 disposed in parallel with the intake camshaft 15. The control shaft 41 is configured to be rotationally driven. Further, as shown in FIG. 3, a control arm 42 is fixed to the control shaft 41 by a bolt 43. A part of the control arm 42 protrudes in the radial direction of the control shaft 41. An intermediate arm 44 is attached to the protruding portion of the control arm 42 by a pin 45. The pin 45 is disposed at a position eccentric from the center of the control shaft 41. Therefore, the intermediate arm 44 is configured to swing around the pin 45. Rollers 52 and 53, which will be described later, are rotatably provided at the distal end of the intermediate arm 44.

  A swing cam arm 50 is swingably supported on the control shaft 41. The swing cam arm 50 has a slide surface 50 a on the side facing the first intake cam 16. The slide surface 50 a is formed so as to contact the second roller 53. The slide surface 50a is formed in a curved surface such that the distance from the first intake cam 16 gradually decreases as the second roller 53 moves from the distal end side of the swing cam arm 50 toward the axial center side of the control shaft 41. ing. The swing cam arm 50 has a swing cam surface 51 on the opposite side of the slide surface 50a. The rocking cam surface 51 has a non-acting surface 51a formed so that the distance from the rocking center of the rocking cam arm 50 is constant, and the position away from the non-working surface 51a is closer to the axis of the control shaft 41. It is comprised with the action surface 51b formed so that distance may become far.

  A first roller 52 and a second roller 53 are disposed between the slide surface 50 a and the peripheral surface of the first intake cam 16. More specifically, the first roller 52 is disposed so as to contact the peripheral surface of the first intake cam 16. The second roller 53 is disposed so as to contact the slide surface 50 a of the swing cam arm 50. The first roller 52 and the second roller 53 are rotatably supported by a connecting shaft 54 fixed to the distal end portion of the intermediate arm 44. Since the intermediate arm 44 swings around the pin 45 as a fulcrum, the rollers 52 and 53 also swing along the slide surface 50 a and the peripheral surface of the first intake cam 16 while maintaining a certain distance from the pin 45.

  The swing cam arm 50 is formed with a spring seat 50b. One end of a lost motion spring 38 is hung on the spring seat 50b. The other end of the lost motion spring 38 is fixed to a stationary part of the internal combustion engine 1. The lost motion spring 38 is a compression spring. Due to the urging force received from the lost motion spring 38, the slide surface 50 a of the swing cam arm 50 is pressed against the second roller 53, and further, the first roller 52 is pressed against the first intake cam 16. As a result, the first roller 52 and the second roller 53 are positioned in a state of being sandwiched between the slide surface 50a and the peripheral surface of the first intake cam 16 from both sides.

  The rocker arm 35 is disposed below the swing cam arm 50. A rocker roller 36 is provided on the rocker arm 35 so as to face the swing cam surface 51. The rocker roller 36 is rotatably attached to an intermediate portion of the rocker arm 35. One end of the rocker arm 35 is supported by a valve shaft 14 a of the intake valve 14, and the other end of the rocker arm 35 is rotatably supported by a hydraulic lash adjuster 37. During the lift operation, the valve shaft 14a is biased by the valve spring 14b in the closing direction, that is, the direction in which the rocker arm 35 is pushed up. The rocker roller 36 is pressed against the swing cam surface 51 of the swing cam arm 50 by the biasing force and the hydraulic lash adjuster 37.

  According to the configuration of the variable valve mechanism 40 described above, the pressing force of the first intake cam 16 is transmitted to the slide surface 50 a via the first roller 52 and the second roller 53 as the first intake cam 16 rotates. Is done. As a result, when the contact point between the swing cam surface 51 and the rocker roller 36 extends from the non-operation surface 51a to the operation surface 51b, the rocker arm 35 is pushed down and the intake valve 14 is opened.

  Further, according to the configuration of the variable valve mechanism 40, when the rotation angle of the control shaft 41 is changed, the position of the second roller 53 on the slide surface 50a changes, and the swing cam arm 50 swings during the lift operation. The range changes. More specifically, when the control shaft 41 is rotated in the counterclockwise direction in FIG. 3, the position of the second roller 53 on the slide surface 50 a moves to the tip side of the swing cam arm 50. Then, the rotation angle of the swing cam arm 50 that is actually required until the rocker arm 35 starts to be pressed after the swing cam arm 50 starts swinging by transmitting the pressing force of the first intake cam 16 is: The control shaft 41 becomes larger as it rotates counterclockwise in FIG. That is, the operating angle and lift amount of the intake valve 14 can be reduced by rotating the control shaft 41 counterclockwise in FIG. Conversely, the operating angle and lift amount of the intake valve 14 can be increased by rotating the control shaft 41 in the clockwise direction. The rotation of the control shaft 41 is performed by rotating the sector gear 82 by, for example, an electric motor.

  Thus, according to the variable valve mechanism 40, the operating angle can be changed by changing the valve lift amount of the intake valve 14. Further, according to the intake side VVT 32, the valve opening timing and the valve closing timing of the intake valve 14 can be changed by the same angle. According to the variable valve mechanism 40 and the intake side VVT 32, the operating angle and the opening / closing timing of the intake valve 14 can be freely changed. In this embodiment, the variable valve gear 18 and the intake side VVT 32 correspond to the valve gear in the present invention.

[Features of Example 1]
In this embodiment, the amount of unburned fuel is reduced by lowering the swirl ratio and increasing the effective compression ratio when the internal combustion engine 1 is under a low load. That is, by delaying the opening timing of the intake valve 14 from before the intake stroke top dead center to the intake stroke top dead center side, the swirl ratio is reduced to suppress fuel overdiffusion. Further, the effective compression ratio is increased by advancing the closing timing of the intake valve 14 to the intake stroke bottom dead center side after the intake stroke bottom dead center.

  FIG. 4 is a diagram showing the lift amount of the intake valve 14 with respect to the crank angle. When the lift amount of the intake valve 14 is larger than 0, the intake valve 14 is open. The crank angle while the intake valve 14 is open corresponds to the operating angle. As shown in FIG. 4, the smaller the maximum lift amount of the intake valve 14 (the more the maximum lift amount moves in the direction of the arrow in FIG. 4), the smaller the operating angle. Further, the smaller the maximum lift amount of the intake valve 14, the smaller the area surrounded by the lift amount locus and the horizontal axis, the so-called time area. In FIG. 4, the intake side VVT 32 is controlled so that the opening timing of the intake valve 14 does not change.

  FIG. 5 is a diagram showing the relationship between the opening timing of the intake valve 14 and the swirl ratio. The vertical axis is the average value of the swirl ratio. The abscissa represents the opening timing of the intake valve 14 and represents the time before the intake stroke top dead center by the crank angle. That is, the opening timing of the intake valve 14 is advanced in the direction of the arrow in FIG. The swirl ratio increases as the intake valve 14 opens earlier than the top dead center of the intake stroke. In this case, the amount of change in the swirl ratio with respect to the amount of change in the opening timing of the intake valve 14 is substantially constant.

  Next, FIG. 6 is a diagram showing the relationship between the closing timing of the intake valve 14 and the effective compression ratio. The vertical axis represents the effective compression ratio, and the horizontal axis represents the closing timing of the intake valve 14 and how far behind the intake stroke bottom dead center is represented by the crank angle. In other words, the intake valve 14 closes earlier in the direction of the arrow in FIG. The effective compression ratio increases as the closing timing of the intake valve 14 approaches the intake stroke bottom dead center after the intake stroke bottom dead center. In this case, the change amount of the effective compression ratio with respect to the change amount of the closing timing of the intake valve 14 becomes smaller as the intake stroke bottom dead center is approached.

  By the way, if the operating angle of the intake valve 14 is reduced by the variable valve mechanism 40, the maximum lift amount of the intake valve 14 is reduced, so that the passage area is reduced. For this reason, if the flow rate of the intake air is constant, the flow rate of the intake air is increased by the reduction of the passage area, so that the swirl ratio is increased. Therefore, if the operating angle of the intake valve 14 is made too small in order to reduce the swirl ratio, the opposite effect is obtained. That is, there is an optimum value for the operating angle of the intake valve 14.

  In contrast, in this embodiment, the swirl ratio in the cylinder 2 is estimated from the working angle of the intake valve 14 and the flow rate of the intake air, and the working angle of the intake valve 14 is changed so that this estimated value becomes the target value. . The swirl ratio is a value corresponding to the strength of the swirling flow, and instead, the flow velocity of the intake air or the strength of the tumble flow can be used. Further, the operating angle may be represented by an area surrounded by the lift amount locus shown in FIG. 4 and the horizontal axis, that is, a so-called time area. As described above, the smaller the operating angle, the smaller the time area.

  Here, FIG. 7 is a diagram showing the relationship between the operating angle of the intake valve 14 and the swirl ratio. The working angle may be a time area. The swirl ratio decreases as the operating angle increases, and the swirl ratio increases as the operating angle decreases. Further, the engine speed or supercharging pressure increases in the direction of the arrow shown in FIG. 7, and the swirl ratio increases as the engine speed increases or the supercharging pressure increases.

  For this reason, in this embodiment, the ECU 60 stores two two-dimensional maps related to the swirl ratio.

Here, the flow rate of the intake air is proportional to the value obtained by dividing the flow rate of the intake air by the time area. That is, as the intake air flow rate increases or the time area decreases, the intake air flow rate increases and the swirl ratio increases. With this relationship, a map is provided in which the estimated value of the swirl ratio decreases as the operating angle increases, and the estimated value of the swirl ratio increases as the operating angle decreases. According to this map, the required operating angle can be calculated even during a transition.

  Further, the swirl ratio increases as the intake air amount increases. For example, at the time of high rotation or high boost pressure, the intake air amount increases, so the flow rate of intake air increases. This increases the swirl ratio. Therefore, a map is provided that corrects the estimated value of the swirl ratio to increase as the intake air flow rate increases, and corrects the estimated value of the swirl ratio to decrease as the intake air flow rate decreases. The flow rate of intake air can be engine speed or supercharging pressure.

  Note that the relationship between the operating angle and the engine speed and the swirl ratio may be obtained in advance through experiments or the like and mapped.

  Further, in this embodiment, when the operating angle is too small at low load and the actual swirl ratio becomes larger than the target value, the amount of change in the closing timing is larger than the amount of change in the opening timing of the intake valve 14. However, the swirl ratio is reduced by increasing the operating angle. In this case, the operating angle may be increased by delaying the closing timing without changing the opening timing of the intake valve 14. Here, since the displacement amount of the piston 3 with respect to the change in the crank angle is small near the bottom dead center of the intake stroke, the change in the effective compression ratio is small as shown in FIG. On the other hand, when the opening angle of the intake valve 14 is changed to increase the operating angle, the swirl ratio increases as shown in FIG. That is, by delaying the closing timing of the intake valve 14 and increasing the operating angle, it is possible to suppress the change in the effective compression ratio while reducing the swirl ratio, and thus it is possible to suppress the discharge of unburned fuel. .

  FIG. 8 is a flowchart showing a control flow of swirl at a low load according to the present embodiment. This routine is executed when the load of the internal combustion engine 1 is lower than the threshold value.

  In step S101, the operating angle of the intake valve 14 is determined. That is, the operating angle of the intake valve 14 is determined so that the opening timing of the intake valve 14 is changed to the intake stroke top dead center side and the closing timing of the intake valve 14 is changed to the intake stroke bottom dead center side. For example, the opening timing of the intake valve 14 is defined as the intake stroke top dead center, and the closing timing is defined as the intake stroke bottom dead center. For these, optimum values are obtained in advance by experiments or the like. Note that the operating angle determined in step S101 is a provisional value.

  In step S102, the swirl ratio is estimated. The swirl ratio is obtained by calculating the relationship between the working angle and engine speed and the swirl ratio in advance through experiments or the like, and substituting the working angle and engine speed into the map. In the present embodiment, the ECU 60 that processes step S102 corresponds to the detection means in the present invention.

  In step S103, the target value of the swirl ratio is acquired. The target value of the swirl ratio is determined according to the engine speed, for example. The relationship between the target value and the engine speed is obtained in advance through experiments or the like and stored in the ECU 60. In this embodiment, the ECU 60 that processes step S103 corresponds to the target value setting means in the present invention.

  In step S104, it is determined whether the estimated swirl ratio is lower than the target value. In this step, it is determined whether or not the swirl ratio is too high.

  If an affirmative determination is made in step S104, the process proceeds to step S105, and if a negative determination is made, the process proceeds to step S106.

In step S105, the opening timing and operating angle of the intake valve 14, which are provisionally determined at the present time, are determined as target values. Thereafter, this routine is terminated. Then, the variable valve mechanism 40 and the intake side VVT 32 are operated so that the opening timing and the operating angle of the intake valve 14 become target values.

  In step S106, the tentatively determined operating angle is increased. At this time, the operating angle is increased while the amount of change in the closing timing is made larger than the amount of change in the opening timing of the intake valve 14. The operating angle may be increased by delaying the closing timing while maintaining the opening timing of the intake valve 14. By doing in this way, a swirl ratio can be made small, suppressing the change of an effective compression ratio. Note that the closing timing of the intake valve 14 may be delayed by a predetermined amount, or may be delayed by a value corresponding to the difference between the estimated value of the swirl ratio and the target value. Thereafter, the process returns to step S102, and the swirl ratio is estimated again according to the operating angle changed in step S106. In this case, the estimated value of the swirl ratio estimated in step S102 is smaller than the previous value. In this embodiment, the ECU 60 that processes step S106 corresponds to the changing means in the present invention.

  As described above, according to this embodiment, the swirl ratio can be estimated from the operating angle and the engine speed. Then, the operating angle can be changed by comparing the estimated value of the swirl ratio with the target value. Thereby, a swirl ratio can be approximated to an optimal value. That is, the discharge of unburned fuel can be suppressed by generating a swirling flow having an appropriate strength in the cylinder 2.

DESCRIPTION OF SYMBOLS 1 Internal combustion engine 2 Cylinder 3 Piston 4 Cylinder block 5 Crankshaft 6 Crank angle sensor 8 Cylinder head 12 Intake port 14 Intake valve 15 Intake camshaft 16 Intake cam 18 Variable valve gear 23 Accelerator pedal 24 Accelerator opening sensor 26 Intake passage 27 Throttle 28 Exhaust port 29 Exhaust valve 31 Intake side pulley 32 Variable rotation phase mechanism (intake side VVT)
40 Variable valve mechanism 60 ECU

Claims (2)

In a control device for an internal combustion engine provided with a valve gear that changes a valve opening characteristic of the intake valve by changing a lift amount of the intake valve,
Detecting means for detecting the strength of the swirling flow in the cylinder from the operating angle of the intake valve and the engine speed;
Target value setting means for setting a target value of the strength of the swirling flow;
Changing means for changing the working angle of the intake valve to bring the strength of the swirling flow detected by the detecting means closer to the target value set by the target value setting means;
A control device for an internal combustion engine, comprising:   When reducing the strength of the swirling flow in the cylinder at low load, the swirling flow strength is increased by increasing the operating angle while increasing the amount of change in the closing timing rather than the amount of change in the opening timing of the intake valve. 2. The control device for an internal combustion engine according to claim 1, wherein the control device is made smaller.

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