JP2006200402A - Variable valve system for internal combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To accurately control a variable valve system by building a model of the variable valve system dealing with nonlinear characteristics caused by valve spring reaction force or the like. <P>SOLUTION: Operation characteristics of the variable valve system is modeled by using spring rate calculated as "driving force of an actuator/displacement of lift characteristics" at every displacement of lift characteristics to be varied, control command value of the actuator is calculated with using the model. Change direction of lift characteristics is determined (S12, S13). Different correction values are calculated for determination results (S13, S15, S16) to correct control command value. <P>COPYRIGHT: (C)2006,JPO&NCIPI

Description

  The present invention relates to a variable valve operating apparatus for an internal combustion engine that varies the lift characteristics (lift amount, operating angle) of an intake valve or an exhaust valve of the internal combustion engine in accordance with the operating state of the engine.

As is well known, the intake valve / exhaust valve lift is used to improve fuel economy at low engine speed and low load, to ensure stable operation, and to ensure sufficient output by improving intake charging efficiency at high speed and high load. There has been proposed a variable valve operating apparatus for an internal combustion engine in which the characteristics can be varied in accordance with the operating state of the engine.
As this type of variable valve device, the one disclosed in Patent Document 1 performs control by treating a reaction torque of nonlinear characteristics acting on a control shaft of the variable valve device as a disturbance by a reaction force of a valve spring. .
JP 2001-3773 A

However, in the variable valve operating device described in Patent Document 1, the relationship between the stability of the feedback control system and the reaction torque cannot be clarified, and variations in the feedback control system and the effects of secular change can be easily handled. Absent.
In stability analysis, a transfer function that is a multiplication of the transfer function of the controlled object and the transfer function of the feedback compensator is required, but the reaction force torque is treated as a disturbance and the controlled object (variable valve gear) is controlled. Since the reaction force torque action is not incorporated in the model, the modeling is insufficient, and even if the round transfer function is obtained using the transfer function of the insufficient control target, the stability analysis cannot be performed correctly.

  The present invention has been made paying attention to such a problem, and an object of the present invention is to build a model of a variable valve apparatus corresponding to non-linear characteristics caused by a valve spring reaction force and the like and control it with high accuracy. .

  Therefore, the present invention is configured to change the lift characteristic of the intake valve or exhaust valve of the engine by controlling the driving force of the actuator, and the relationship between the driving force of the actuator and the displacement of the lift characteristic is nonlinear. In the variable valve system for an internal combustion engine having various characteristics, the operating characteristic of the variable valve system is calculated by using a spring constant calculated as “actuator driving force / lift characteristic displacement” for each variable displacement of the lift characteristic. The control command value for the actuator is calculated using this model, and the control command value is corrected by calculating a different correction value according to the change direction of the lift characteristic.

  According to the present invention, by using a spring constant calculated for each lift characteristic, a variable valve device model corresponding to a non-linear characteristic caused by a valve spring reaction force or the like over the entire variable range of the valve characteristic can be obtained. Since it can be constructed, it is possible to easily cope with characteristic differences due to individual variations and improve control performance. In addition, by performing different corrections according to the direction of change in the lift characteristics, the effect of hysteresis can be effectively eliminated from the response of the lift characteristics, and good follow-up to the response desired by the designer is achieved. it can.

Hereinafter, embodiments of the present invention will be described with reference to the drawings.
1 to 3 show an embodiment in which the variable valve operating apparatus for an internal combustion engine according to the present invention is applied to the intake valve side. In FIG. 1, the configuration on the exhaust valve side (lower side in FIG. 1) is not shown.
A drive shaft 11 that is continuous over all the cylinders is provided on the upper portion of the cylinder head 10. The drive shaft 11 has a sprocket attached to one end (not shown) and rotates in conjunction with the crankshaft of the engine via a timing chain or the like.

A cylindrical bearing portion 18a of a swing cam 18 that drives an intake valve (or exhaust valve) 19 is fitted on the outer periphery of the drive shaft 11 so as to be relatively rotatable. The swing cam 18 has a thin plate shape having a tip (cam nose) 18b, and a cam surface 18c that slides on an upper surface 19b of a valve lifter 19a as a transmission member provided at the upper end of the intake valve 19 is formed on the outer periphery thereof. Has been.
A ring-shaped eccentric cam 12 is fixed to the outer periphery of the drive shaft 11 by press fitting or the like. The center (axial center) C2 of the eccentric cam 12 is eccentric by a predetermined amount with respect to the center (axial center) C1 of the drive shaft 11. A base portion 13a of a ring-shaped link 13 is fitted on the outer periphery of the eccentric cam 12 so as to be relatively rotatable via a bearing or the like. Note that the swing center (axial center) of the swing cam 18 coincides with the center C1 of the drive shaft 11.

A control shaft 14 extends in the cylinder row direction substantially parallel to the drive shaft 11 obliquely above the drive shaft 11. The control shaft 14 is rotated by a predetermined angle according to the operating state of the engine by a drive unit 20 described later and held.
A ring-shaped control cam 15 is fixed to the outer periphery of the control shaft 14 by press fitting or the like. The center (axial center) C4 of the control cam 15 is eccentric by a predetermined amount with respect to the center (axial center) C3 of the control shaft 14. A cylindrical central base of the rocker arm 16 is fitted on the outer periphery of the control cam 15 so as to be relatively rotatable. One end portion 16a of the rocker arm 16 and the small-diameter tip portion 13b of the ring-shaped link 13 are coupled to each other via a first pin 29a that passes through both the 16a and 13b.

  The other end 16 b of the rocker arm 16 and the swing cam 18 are linked by a rod-shaped link 17. More specifically, the other end portion 16b of the rocker arm 16 and the one end portion 17a of the rod-shaped link 17 are coupled to each other via a second pin 29b that passes through both the 16b and 17a so as to be relatively rotatable. Further, the other end 17b of the rod-shaped link 17 and the swing cam 18 are coupled to each other via a third pin 29c through which both the ends 17b and 18 are inserted.

Next, the configuration of the drive unit 20 that rotates and holds the control shaft 14 will be described.
As shown in FIG. 1, the control shaft 14 extends into a case 22 fixed to the cylinder head 10, and a worm wheel 21 is fixed to one end thereof. An electric motor 26 driven by a control signal from an ECU (Engine Control Unit) 50 is attached to the case 22, and an output shaft 26 a of the electric motor 26 rotates into the case 22 via a roller bearing 25. It extends as possible. A worm gear 24 that meshes with the worm wheel 21 is fixed to the output shaft 26a. In order to increase the motor torque between the worm gear 24 and the worm wheel 21, the gear ratio is set appropriately large. Further, a rotation angle sensor 23 for detecting the rotation angle of the control shaft 14 (worm wheel 21) is attached to the case 22, and the output of the rotation angle sensor 23 is input to the ECU 50, and the rotation angle sensor The electric motor 26 is feedback-controlled based on the rotation angle of the control shaft 14 detected at 23, that is, the detected value of the operating angle (lift amount) of the intake valve 19.

  With such a configuration, when the drive shaft 11 rotates in conjunction with the rotation of the engine, the ring-shaped link 13 moves in translation via the eccentric cam 12, and the rocker arm 16 swings the center C4 of the control cam 15 accordingly. The swing cam 18 swings as a moving center, and the swing cam 18 swings through the rod-shaped link 17. At this time, the cam surface 18c of the swing cam 18 is in sliding contact with the upper surface of the valve lifter 19a as a transmission member provided at the upper end of the intake valve 19, and the valve lifter 19a is pressed against the reaction force of a valve spring (not shown). As a result, the intake valve 19 opens and closes in conjunction with the rotation of the engine.

  Further, when the output shaft 26a of the electric motor 26 is rotationally driven in accordance with the operating state of the engine, the control shaft 14 rotates through the worm gear 24 and the worm wheel 21, and the control cam that becomes the rocking center of the rocker arm 16 is rotated. The position of the center C4 of 15 changes, and the lift characteristic of the intake valve 19 changes continuously. More specifically, the valve lift amount is increased as the control shaft 14 is largely rotated (that is, the operating angle of the control shaft 14 is increased) and the distance between the center C4 of the control cam 15 and the center C1 of the drive shaft 11 is reduced. In addition, the valve operating angle increases.

Next, the operational characteristics of the variable valve device (mechanism) will be discussed with reference to FIGS.
A reaction force F1 generated by a valve spring reaction force of the intake valve 19 acts on the other end portion 16b of the rocker arm 16 via the swing cam 18, the rod-shaped link 17, the second pin 29b, and the like. A reaction force F2 generated as a reaction acts on the one end 16a of the rocker arm 16 via the eccentric cam 12, the ring-shaped link 13, the first pin 29a, and the like. Therefore, the combined reaction force F3 of the reaction forces F1 and F2 substantially acts on the rocking center C4 of the rocker arm 16.

As a result, a torque T1 that is the product of the arm length r1 from the center C3 of the control shaft 14 to the direction line of the combined reaction force F3 and the combined reaction force F3 acts on the control shaft 14. Therefore, in order for the drive unit 20 to hold the control shaft 14 at a predetermined angle, at least a reverse torque that matches the torque T1 is required.
In a state where the control shaft 14 is held at a predetermined rotation angle, as shown in FIG. 4, when the swing cam 18 is pushed down to the highest lift side, that is, swings most counterclockwise in FIG. Sometimes the combined reaction force F3 becomes maximum. The direction of the resultant reaction force F3 at this time is substantially parallel to the first line L1 connecting the center C1 of the drive shaft 11 and the center C3 of the control shaft 14.

  Here, as shown in FIG. 6, the resultant reaction force F3 increases as the lift amount (and the operating angle) increases [(A): minimum operating angle, (B): intermediate position, (C): The maximum operating angle] increases as the arm length r1 increases from the minimum operating angle to the intermediate position as the eccentric cam 12 rotates, but then decreases. Therefore, the torque T1, which is the product of the combined reaction force F3 and the arm length r1, has a nonlinear characteristic with respect to the operating angle θ.

FIG. 7 shows the relationship between the operating angle θ _CS of the control shaft 14 and the drive current i _CS (proportional to the torque T1). In this embodiment, as shown in FIG. 7, the spring constant K (θ _CS ) for each operating angle θ _{CS is} expressed by the following equation with respect to the nonlinear “operating angle θ _CS −driving current i _CS characteristics”. Positioning control of the operating angle θ _CS is performed while modeling the operating characteristics of the variable valve mechanism to be controlled using the spring constant K (θ _CS ).

A (θ _CS ) = i (θ _CS ) / θ _CS (1)
K (θ _CS ) = A (θ _CS ) · K _T (2)
Where K _T is the torque constant of the electric motor 26.
That is, as shown in FIG. 8, a value obtained by multiplying A (θ _CS ) obtained by dividing the drive current i (θ _CS ) by the operating angle θ _CS and the torque constant K _T of the electric motor 26 at each operating angle θ _CS . The spring constant K (θ _CS ) is set (see FIG. 8), and the calculated spring constant K (θ _CS ) is assigned to each operating angle θ _CS to create an “operating angle-spring constant map” shown in FIG. By defining the spring constant in this way, the operating angle and driving current having non-linear characteristics can be set as a continuous and unambiguous function, and the variable valve mechanism can be modeled as described later. .

FIG. 10 is a block diagram showing a positioning control system for the operating angle θ _CS executed by the ECU 50, which is roughly divided into a dynamic characteristic compensator 101, a responsiveness compensator 102, an offset compensator 103, and a variable motion that is a control target. A valve mechanism 104 is included.
First, the transfer function Gp (s) of the variable valve mechanism 104 can be expressed as a product of a dynamic characteristic and a static characteristic in a zero order / second order as shown in the following equation (modeling of the variable valve mechanism) ). Here, the term on the left side of the right side of the formula below represents dynamic characteristics, and the term on the right side represents static characteristics.

However, J: Moment of inertia of the variable valve mechanism, D: Viscosity resistance of the variable valve device, K _T : Torque constant of the electric motor, K (θ _CS ): Spring constant (from the above “operating angle-spring constant map” Calculated), θ _CS : the operating angle of the variable valve gear.
Based on the above, each element of the control system shown in FIG. 10 will be described.
The dynamic characteristic compensator 101 corresponds to a feedforward compensator according to the present invention, and a target operating angle calculator whose response (damping ratio ζ, frequency ω) desired by the designer is expressed by the following equation (4). Given the transfer function G _T (s) of 102a,

G _FF (s) = G _T (s) / G _P (s) so that the operating angle θ _CS follows the input value and the final target value, which is the final operating value θ _CST , by the dynamic characteristic G _T (s). ), The dynamic characteristic compensation output i _CSFF that is the feedforward portion of the drive current is calculated as shown in the following equation (5) using the transfer function GFF (s) of the dynamic characteristic compensator 101 obtained from the relationship of The result is output to the adding unit 111. That is, the dynamic characteristic compensator 101 is composed of a secondary / secondary filter.

The responsiveness compensator 102 includes a target operating angle calculator 102a, a subtractor 110, and a dynamic characteristic compensation output compensator 102b. The target operating angle calculator 102a corresponds to the target lift characteristic setting means according to the present invention, and the dynamic characteristic compensation output compensator 102b corresponds to the feedback compensator according to the present invention.
The target operating angle calculator 102a receives the reaching operating angle θ _CST set according to the operating state of the engine, and calculates the target operating angle θ _CSM that is the response of the operating angle desired by the designer based on the equation (6). Calculate and output to the subtractor 110 and the offset compensator 103. This target operating angle θ _CSM is a transient operating angle until the operating angle θ _CS reaches the final reached operating angle θ _CST .

As shown in the equation (7), the subtraction unit 110 subtracts the actual operating angle θ _CS from the target operating angle θ _CSM to calculate the deviation amount θ _ERR and outputs it to the dynamic characteristic compensation output compensator 102b.
θ _ERR = θ _CSM −θ _CS … (7)
The dynamic characteristic compensation output compensator 102b calculates the dynamic characteristic compensation output correction value i _CSFB from the deviation amount θ _ERR using a filter having an integral characteristic and whose stability is compensated for the change in the parameter to be controlled. And output to the first adder 111. As an example of the filter having the above-mentioned integral characteristic, the equation (8) is given. Here, P and I are gains whose stability is compensated. In this case, the dynamic characteristic compensation output compensation value i _CSFB is calculated by the equation (9).

G _FB (s) = (Ps + I) / s (8)
i _CSFB = [(Ps + I) / s] · θ _ERR (9)
As shown in the equation (10), the first addition unit 111 includes the dynamic characteristic compensation output i _CSFF calculated by the dynamic characteristic compensation unit 101 and the dynamic characteristic compensation output correction value i calculated by the dynamic characteristic compensation output compensator 102b. _The basic current command value i _CMD is calculated by adding _CSFB and output to the second addition unit 112. The basic current command value i _CMD corresponds to the control command value according to the present invention.

i _CMD = i _CSFB + i _CSFF (10)
The offset compensator 103 calculates the offset correction amount i _OFFSET based on the input target operating angle (normative response) θ _CSM and outputs it to the second adder 112. The offset correction amount i _OFFSET corresponds to a correction value according to the present invention, and details of the calculation process of the offset correction amount i _OFFSET will be described later (see FIGS. 11 to 14).

Then, the second adder 112 calculates the final current command value i _CSC by adding the input basic current command value i _CMD and the offset correction amount i _OFFSET , as shown in the equation (11). It outputs to the valve device 104 (the electric motor 26 thereof).
i _CSC = i _CMD + i _OFFSET (11)
Next, the offset compensator 103 and the offset correction amount i _OFFSET calculation process executed in the offset compensator 103 will be described.

The basic configuration of the control system is to output the basic current command value i _CMD to the variable valve operating device. As shown in FIG. 11A, the actual operation of the variable valve operating device is performed. The relationship between the angle θ _CS and the drive current i _CS is that when the operating angle θ _CS increases (when the lift amount increases) due to the effects of friction, valve springs, etc., it decreases (the lift amount decreases). has a hysteresis drive current i _CS increases at the same operating angle theta _CS than one time) in the direction of, and, in offset0_line in the basic current instruction value i _CMD (FIG when operating angle theta _CS increases It is confirmed that a drive current smaller than the basic current command value i _CMD is sufficient when a larger drive current is required and the operating angle θ _CS decreases.

Therefore, simply outputting the basic current command value i _CMD may cause a difference between the command value and the actual operation of the variable valve operating system, resulting in poor responsiveness or supplying a larger drive current than necessary. May increase power consumption.
Therefore, in the present embodiment, a table as shown in FIG. 11B is created in advance based on the actual machine data (operation angle-current value) shown in FIG. By referring to this table, even when the operating angle is the same, an offset correction amount i _OFFSET that differs depending on whether the lift amount is increasing (lifting up) or the lift amount decreasing (lifting down) is calculated. The final current command value i _CSC is calculated by correcting the (basic) current command value i _CMD with the calculated offset correction amount i _OFFSET .

FIG. 12 shows a main routine for calculating the offset correction amount i _OFFSET . In FIG. 11, in S1, it is determined whether or not the minimum value θ _MIN <(current) operating angle θ _CS <the maximum value θ _MAX . Here, the maximum value θ _MAX indicates the maximum operating angle, and at this time, the valve lift amount becomes maximum. The minimum value θ _MIN indicates the minimum operating angle, and at this time, the valve lift is minimized. If θ _MIN <θ _CS <θ _MAX , that is, if the operating angle θ _CS is not the maximum value θ _MAX or the minimum value θ _MIN , the process proceeds to S2 and first offset correction amount calculation control is performed (see FIG. 13). . On the other hand, if θ _CS = θ _MAX or θ _CS = θ _MIN , that is, if the operating angle θ _CS is the maximum operating angle or the minimum operating angle, the process proceeds to S3, and second offset correction amount calculation control is performed. (See FIG. 14).

FIG. 13 shows the contents of the first offset correction amount calculation control executed in S2 of FIG. In S11, the previous value θ _{CSM (−1)} is subtracted from the inputted target operating angle θ _CSM to calculate the change amount Grad (= θ _CSM −θ _{CSM (−1)} ).
In S12, it is determined whether or not the calculated change amount Grad is positive (Grad> 0). If Grad> 0, that is, if the lift amount increases (lift lift direction), the process proceeds to S13. If Grad ≦ 0, the process proceeds to S14.

In S13, the difference “deg_i_up (> 0)” between “up_offset_line (solid line)” and “offset0_line (one-dot chain line)” in FIG. 13B is calculated as the current correction value X _LPF based on the target operating angle θ _CSM . To do. Note that “up_offset_line” is a current value required to change the operating angle θ _CS of the variable valve device when increasing the lift amount, and “offset0_line” is the basic current command value as described above. It corresponds to i _CMD .

In S14, it is determined whether or not the calculated change amount Grad is negative (Grad <0). If Grad <0, that is, if the valve lift is to be lowered (decrease the lift amount), the process proceeds to S15. If Grad <0, that is, if Grad = 0, the process proceeds to S16.
In S15, the difference “deg_i_dn (<0)” between “dn_offset_line (dashed line)” and “offset0_line (dashed line)” in FIG. 13B is calculated as the current correction value X _LPF based on the target operating angle θ _CSM . To do. Note that “dn_offset_line” is a current value necessary to change the operating angle θ _CS of the variable valve apparatus when the lift amount is decreased.

In S16, the current correction value X _LPF = X _{LPF (−1)} is set and the previous value is held.
In S17, as shown in the equation (12), the current correction value X _LPF calculated in S13, S15, or S16 is subjected to a low-pass filter to calculate the offset correction amount i _OFFSET .
i _OFFSET = [1 / (1 + Ts)]. X _LPF (12)
FIG. 14 shows the contents of the second offset correction amount calculation control executed in S3 of FIG. In S21, it is determined whether or not the operating angle θ _CS is the minimum value θ _MIN . If θ _CS = θ _MIN , the process proceeds to S22, and if θ _CS ≠ θ _MIN , the process proceeds to S23.

In S22, the minimum operating angle is the direction in which the lift amount is necessarily increased. Therefore, based on the input target operating angle θ _CSM , “up_offset_line (solid line)” and “offset0_line (in FIG. 13B) The difference “deg_i_up (> 0)” from the “dashed line” is set as the offset correction amount i _OFFSET .
In S23, the maximum operating angle is the direction in which the lift amount is necessarily reduced. Therefore, based on the input target operating angle θ _CSM , “dn_offset_line (broken line)” and “offset0_line ( The difference “deg_i_dn (<0)” from the “dashed line” is set as the offset correction amount i _OFSSET .

As described above, the variable valve system is modeled using the spring constant K (θ _CS ) calculated for each operating angle, and the current command value i _CSC is calculated and controlled using the model, thereby changing the parameter variation. In addition, it is difficult to be affected by disturbances and disturbances, and the working angle response desired by the designer can be obtained.
In particular, the change direction of the lift amount is determined, and based on actual machine data, an offset correction amount i _OFFSET that differs between the lift amount increasing direction and the lift decreasing direction is calculated, and the (basic) current command value is calculated based on this offset correction amount i _OFFSET Since the final current command value i _CSC is calculated by correcting i _CMD , the effect of hysteresis is removed, and the operating angle follows the norm response well regardless of the change direction of the lift amount (operating angle). Can be realized. Further, when the (current) operating angle θ _CS is the minimum operating angle, it is always changed in the lift increasing direction, so “deg_i_up” corresponding to the target operating angle θ _CSM is set as the offset correction amount i _OFFSET. If the (current) operating angle θ _CS is the maximum operating angle, it will always be changed in the lift amount decreasing direction, so that “deg_i_dn” corresponding to the target operating angle θ _CSM is offset. By immediately calculating the correction amount i _{OFFSET as the} correction amount, the control can be simplified, and good follow-up can be realized even when the operating angle changes from the minimum / maximum operating angle.

In the above-described embodiment, the offset compensator 103 receives the reference response (θ _CSM ) as an input, and is based on the reference response (that is, by comparison with the previous value) in the lift amount increasing direction or the lift amount decreasing direction. Although determines whether, instead of this, as shown in FIG. 15, a dynamic characteristic compensation output i _CSFF is output from the dynamic characteristic compensating unit 101 as an input, based on animal characteristics compensation output i _CSFF or determines Te (second embodiment), as shown in FIG. 16, inputs the basic current instruction value i _CMD, it may be configured to or determined based on the basic current instruction value i _CMD ( Third embodiment).

Further, as described above, the relationship between the operating angle θ _CS of the variable valve operating device and the driving current i _CS is such that when the operating angle θ _CS increases, the driving current i _{CS at} the same operating angle θ _{CS is} greater than when the operating angle θ _CS decreases. It has been confirmed that the hysteresis width has a characteristic that the hysteresis width increases as the engine rotational speed Ne increases as shown in FIG. That is, when the operating angle θ _CS increases, the driving current i _CS increases as the engine rotational speed Ne increases. When the operating angle θ _CS decreases, the driving current i _CS increases as the engine rotational speed Ne increases. Tends to be smaller.

Therefore, in the fourth embodiment, in the above first to third embodiments, a table that also considers the engine rotational speed Ne as shown in FIG. 17 (b) based on the actual machine data shown in FIG. 17 (a). An offset correction amount i _OFFSET is calculated based on the displacement amount Grad, the engine rotational speed Ne, and the target operating angle θ _CSM by using this in advance and using this instead of the table shown in FIG. In this case, in the flowcharts of FIGS. 13 and 14, “up_offset_line” is replaced with “up_offset_rpm_line” corresponding to the current engine speed Ne (rpm), and “dn_offset_line” is “dn_offset_rpm_line” corresponding to the current engine speed Ne. Will be replaced. In the figure, only 1000 rpm and 3000 rpm are shown as representative examples, but it goes without saying that the correction amount i _OFFSET can be calculated even at other rotational speeds.

Further, in each of the above embodiments, when the change amount Grad of the target operating angle (normative response) θ _CSM is 0, the previous current correction value X _LPF is held (that is, the final current command value I _CSC also holds the previous value). However, in the case where only the operating angle is maintained, the current value or operating angle between “up_offset_line” and “dn_offset_line” in FIG. 11B or the engine rotational speed Ne is set according to the operating angle. Accordingly, the current value between “up_offset_rpm_line” and “dn_offset_rpm_line” in FIG. 17B may be output, and can be set as appropriate within this range.

Therefore, when the change amount Grad is 0 (that is, in S16), the current correction value X _LPF = 0 and the current value on “offset0_line” is output (fifth embodiment), or the current correction value X _LPF The current value on “dn_offset_line” (or “dn_offset_rpm_line”) may be output as “deg_i_dn” (sixth embodiment). In the former case, it is possible to satisfactorily ensure responsiveness when the change direction of the lift amount (operating angle) is switched, and in the latter case, there is an advantage that current consumption can be suppressed.

The top view of the variable valve apparatus which concerns on embodiment. The principal part sectional drawing of the said variable valve apparatus. The block diagram which shows the drive part of the said variable valve apparatus. The figure for demonstrating the effect | action of the said variable valve apparatus. The block diagram which shows the control shaft and control cam of the said variable valve apparatus. The figure for demonstrating the effect | action in each rotation angle of the said control shaft. The figure which shows the relationship between the operating angle of the said variable valve apparatus, and a drive current. The figure for demonstrating calculation of the spring constant of the said variable valve apparatus. Basic characteristic map of spring constant. The control block diagram of the said variable valve apparatus. The figure which shows the relationship between the operating angle of the said variable valve apparatus, an operating angle (lift amount) change direction, and a drive current. The flowchart which shows the main routine of offset correction amount calculation. The flowchart which shows the content of 1st offset correction amount calculation control. The flowchart which shows the content of 2nd offset correction amount calculation control. The control block diagram of the variable valve apparatus which concerns on 2nd Embodiment. The control block diagram of the variable valve apparatus which concerns on 3rd Embodiment. The figure which shows the relationship between the operating angle of the variable valve operating apparatus, engine rotational speed, operating angle (lift amount) change direction, and drive current which are used in 4th Embodiment.

Explanation of symbols

      DESCRIPTION OF SYMBOLS 11 ... Drive shaft, 12 ... Eccentric cam, 13 ... Ring-shaped link, 14 ... Control shaft, 15 ... Control cam, 16 ... Rocker arm, 17 ... Rod-shaped link, 18 ... Swing cam, 19 ... Intake valve, 20 ... Drive , 50 ... ECU, 101 ... dynamic characteristic compensation part, 102 ... response correction part, 102a ... target operating angle calculator, 102b ... dynamic characteristic compensation output compensator, 103 ... offset compensator, 104 ... variable valve operating device

Claims (8)

An internal combustion engine having a configuration in which lift characteristics of an intake valve or an exhaust valve of an engine are changed by controlling a driving force of an actuator, and a relationship between a driving force of the actuator and a displacement of the lift characteristic is nonlinear. In variable valve gear,
For each variable displacement of the lift characteristic, the operating characteristic of the variable valve operating device is modeled using a spring constant calculated as “actuator driving force / lift characteristic displacement”, and the control command of the actuator is used using this model. Control command value calculating means for calculating a value;
A correction unit that calculates a different correction value according to the change direction of the lift characteristic, and corrects the control command value with the calculated correction value;
A variable valve operating apparatus for an internal combustion engine, comprising: The correction means determines whether the lift amount is in the increasing direction or decreasing direction,
2. The internal combustion engine according to claim 1, wherein the control command value is corrected to increase when the lift amount is in an increasing direction, and the control command value is corrected to decrease when the lift amount is in a decreasing direction. Variable valve gear.   The correction means decreases and corrects the control command value without performing the determination when the current lift amount is the maximum lift amount, and does not perform the determination when the current lift amount is the minimum lift amount. The variable valve operating apparatus for an internal combustion engine according to claim 2, wherein the control command value is corrected to increase.   The variable valve operating apparatus for an internal combustion engine according to claim 1, wherein the correction unit calculates the correction value for each variable lift characteristic.   The variable valve operating apparatus for an internal combustion engine according to any one of claims 1 to 4, wherein the correction means calculates the correction value for each engine speed. The control command value calculating means includes
A feedforward compensator that calculates the feedforward compensation output so that the final lift characteristic set according to the operating state of the engine is input and the actual lift characteristic follows with a desired response;
A target lift characteristic setting unit configured to set a target lift characteristic for realizing the desired response to the ultimate lift characteristic when the ultimate lift characteristic is input;
A feedback compensator for calculating a feedback compensation output based on a deviation between the target lift characteristic and the actual lift characteristic;
6. The variable valve operating apparatus for an internal combustion engine according to claim 1, wherein the control command value is calculated based on the feedforward compensation output and the feedback compensation output.   7. The variable valve operating apparatus for an internal combustion engine according to claim 6, wherein the correction means determines a change direction of the lift characteristic based on the feedforward compensation output, the target lift characteristic or the control command value. . The variable valve operating device is:
A drive shaft that rotates in conjunction with the rotation of the engine,
A swing cam that is fitted on the outer periphery of the drive shaft so as to be relatively rotatable, and that drives the intake valve or the exhaust valve to open and close;
An eccentric cam fixed eccentrically on the outer periphery of the drive shaft;
A ring-shaped link that is fitted on the outer periphery of the eccentric cam so as to be relatively rotatable;
A control shaft extending substantially parallel to the drive shaft;
A control cam eccentrically fixed to the outer periphery of the control shaft;
A rocker arm that is fitted on the outer periphery of the control cam so as to be relatively rotatable, and is linked to the ring-shaped link at one end thereof;
A rod-shaped link that links the other end of the rocker arm and the swing cam;
8. The internal combustion engine according to claim 1, wherein when the actuator rotates the control shaft, a rocking center position of the rocker arm is changed to change the lift characteristics. 9. Variable valve gear.

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