JP2003322034A - Internal-combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide an internal-combustion engine equipped with a plurality of banks (bunch of cylinders) capable of lessening a deviation of the responsiveness in the valve timing control of the banks generated by the loads of auxiliary units, etc., driven by a cam shaft. <P>SOLUTION: The internal-combustion engine 1 of V-form includes A-bank and B-bank whereon high-pressure fuel pumps 20 (auxiliary units) driven by the suction side cam shaft 4 are installed. The high-pressure fuel pumps 20 on different banks should have the same capacity so that the pump loads are applied uniformly to the suction side cam shafts 4 of different banks. Thereby the load torques of the cam shafts 4 are made approximately equal, and the responsiveness in the valve timing control of each bank can be made the same, and it is possible to solve the problem such as torque fluctuation, deterioration of the drivability, etc., upon transient in driving resulting from a deviation of the responsiveness. <P>COPYRIGHT: (C)2004,JPO

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to an internal combustion engine having a variable valve timing mechanism for controlling the valve timing of each cylinder group of an internal combustion engine having a plurality of cylinder groups (a plurality of banks).

[0002]

2. Description of the Related Art The types of cylinder arrangements of internal combustion engines (hereinafter referred to as "engines") for vehicles currently in practical use are classified into in-line type, V type, and horizontally opposed type (H type). Since an in-line engine in which all cylinders are arranged in a row has a long overall engine length, in a multi-cylinder engine with 6 or more cylinders, all cylinders are left and right 2 in order to shorten the overall engine length and improve the mountability on a vehicle. In many cases, a V-type engine that is divided into one bank (cylinder group) and arranged in a V-shape is adopted. The V-type and horizontally-opposed DOHC engines have a configuration in which the intake / exhaust cam shafts are provided in the left and right banks, so that when controlling the valve timing of the intake valve, for example, it is variable for each bank. A valve timing mechanism is mounted on the intake side camshaft to control the valve timing of the intake valve for each bank.

In such an engine with a variable valve timing mechanism, the valve timing is advanced / retarded by advancing / retarding the rotational phase of the cam shaft by the variable valve timing mechanism. The camshaft, whose rotation phase is adjusted by the variable valve timing mechanism, drives the intake valve (or exhaust valve) and, in the case of a cylinder injection engine, pumps the fuel in the fuel tank at high pressure to the fuel injection valve side. The high-pressure fuel pump is also driven.

[0004]

As described above, in the V-type or horizontally-opposed in-cylinder injection type engine with a variable valve timing mechanism, only one bank is used to load auxiliary machinery such as a high-pressure fuel pump. Is applied to the cam shaft, and the cam shaft of the other bank is not loaded with such auxiliary machinery, so that the load torque of the cam shafts of the left and right banks is unbalanced by the load of the auxiliary machinery. In particular, an in-cylinder injection type engine requires a fuel injection pressure of 30 times or more that of an intake port injection type engine, so a large camshaft torque is required to drive a high-pressure fuel pump, and as a result,
There will be a large difference in the load torque of the camshafts on the left and right banks.

Generally, the responsiveness of the change in the actual valve timing (actual camshaft displacement angle) to the change in the target valve timing (target camshaft displacement angle), that is, the responsiveness of the valve timing control, is that the load torque of the camshaft is large. If the camshaft load torques on the left and right banks differ significantly, the responsiveness of the valve timing control will differ significantly between the left and right banks, and as a result, during transient operation in which the target valve timing changes suddenly, There is a problem that the internal EGR amount and the charging efficiency of fresh air become unbalanced, resulting in torque fluctuations and deterioration of drivability.

The present invention has been made in view of the above circumstances, and an object thereof is to provide an auxiliary machine driven by a cam shaft in an internal combustion engine having a plurality of cylinder groups (a plurality of banks). It is possible to reduce the deviation of the responsiveness of valve timing control between each cylinder group caused by the load,
An object of the present invention is to provide an internal combustion engine capable of solving problems such as torque fluctuations during transient operation and deterioration of drivability due to the deviation of the responsiveness.

[0007]

In order to achieve the above object, an internal combustion engine according to a first aspect of the present invention has, in each of a plurality of cylinder groups, an auxiliary machine driven by a cam shaft of each cylinder group. The configuration is such that the auxiliary machines are installed in each cylinder group so that the loads thereof are substantially equal. In this configuration, since the loads of the auxiliary machinery are applied to the cam shafts of the cylinder groups almost uniformly, the load torques of the cam shafts of the cylinder groups can be made substantially uniform,
The responsiveness of the valve timing control of each cylinder group can be made uniform, and problems such as torque fluctuation during transient operation and deterioration of drivability due to the difference in responsiveness can be solved.

In this case, as in claim 2, high pressure fuel pumps having the same capacity may be installed as auxiliary machines in each of the plurality of cylinder groups. In this case, the load (discharge amount) of the high-pressure fuel pump of each cylinder group fluctuates according to the target fuel discharge amount during operation of the internal combustion engine, but the high-pressure fuel pump of each cylinder group is independently controlled for each cylinder group. It is not necessary to do so, and the high pressure fuel pump of each cylinder group may be controlled in the same manner in all operating regions. Therefore, the auxiliary machinery of each cylinder group (high pressure fuel pump)
It is possible to maintain the balanced state of the load in the entire operating range, and it is possible to reliably reduce the deviation of the responsiveness of the valve timing control between the cylinder groups in the entire operating range.

Alternatively, as in claim 3, a high-pressure fuel pump is installed in at least one cylinder group, and an auxiliary machine other than the high-pressure fuel pump, which has substantially the same load as this high-pressure fuel pump, is installed in the other cylinder groups. You may make it install a kind. By doing so, the loads of the auxiliary machinery are applied to the cam shafts of the respective cylinder groups substantially evenly, so that the load torques of the cam shafts of the respective cylinder groups can be substantially equalized, and the valves between the cylinder groups can be It is possible to reduce the deviation of the responsiveness of the timing control.

Further, as in claim 4, the control amount of the valve timing adjusting means may be corrected by the correcting means for each of the plurality of cylinder groups. That is, the load torque of the camshaft of each cylinder group fluctuates due to variations in manufacturing of auxiliary machinery and valve timing adjusting means, their changes over time, operating conditions, etc., and the responsiveness of valve timing control of each cylinder group is increased. Although there is a possibility of a slight deviation, if a function for correcting the control amount of the valve timing adjusting means is provided for each of a plurality of cylinder groups as in claim 4, variations in manufacturing of auxiliary machinery and valve timing adjusting means and aging may occur. It is possible to reduce the deviation of the responsiveness of the valve timing control between the cylinder groups due to changes, operating conditions, etc. by correcting the control amount of the valve timing adjusting means.

[0011]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS An embodiment in which the present invention is applied to a cylinder injection type V DOHC engine will be described below with reference to the drawings.

First, a schematic configuration of the entire system will be described with reference to FIG. V-type engine 1 has all cylinders 2
The engine is divided into one bank (cylinder group) and arranged in a V-shape. In the following description, one bank is referred to as “A bank” and the other bank is referred to as “B bank”. An intake-side camshaft 4 and an exhaust-side camshaft 5 are provided at the top of each bank, and the power from the crankshaft 2 of the engine 1 is supplied to the sprocket 1 of each bank by the timing chain 3a.
3a is transmitted to the intake-side camshaft 4 of each bank, and further, the rotation of the intake-side camshaft 4 of each bank is transmitted to the exhaust-side camshaft 5 of each bank by the timing chain 3b. There is.

Further, in the present embodiment, the high-pressure fuel pumps 20 (auxiliaries) driven by the intake-side camshaft 4 are installed in the banks A and B, respectively. The high-pressure fuel pumps 20 of each bank are high-pressure fuel pumps of the same capacity, and the load of the high-pressure fuel pumps 20 is evenly applied to the intake side camshaft 4 of each bank. The high-pressure fuel pump 20 of each bank raises the pressure of fuel pumped up from a fuel tank (not shown) during engine operation and discharges it into the fuel pipe 19, and the high-pressure fuel is supplied through this fuel pipe 19 to the fuel injection valve of each bank. 21 pressure feed.

A phase difference adjusting device 40, which is a valve timing adjusting means, is provided on each intake side camshaft 4 of each bank.
(A hatched portion in FIG. 1) is provided. The hydraulic oil (engine oil) in the oil pan 28 is the oil pump 2
9 supplies the hydraulic pressure control valve 30 of each bank, and the hydraulic pressure discharged from the hydraulic control valve 30 of each bank is supplied to the phase difference adjusting device 40 of each bank. Then, by controlling the hydraulic pressure supplied to the phase difference adjusting device 40 of each bank by the hydraulic control valve 30, the actual valve timing (actual camshaft displacement angle) of the intake valve (not shown) of each bank is controlled. . In this case, a hydraulic actuator that drives the phase difference adjusting device 40 is constituted by the hydraulic control valve 30, the oil pump 29, and the like.

On the other hand, a cam shaft position sensor 44 is attached to the intake side cam shaft 4 of each bank, and a crank position sensor 42 is attached to the crank shaft 2. This crank position sensor 42 is the crankshaft 2
When N detection pulse signals are generated with one rotation of
The camshaft position sensor 44 of each bank is 1 of the intake side camshaft 4.
With the rotation, 2N detection pulse signals are generated. Further, when the maximum value of the timing conversion angle of the intake side camshaft 4 of each bank is θmax crank angle, N
The number N of detection pulse signals so that <360 degrees / θmax
Is set. As a result, the relative rotation angle θ between the detection pulse signal from the crank position sensor 42 and the detection pulse signal from the cam shaft position sensor 44 of each bank that occurs subsequent to this detection pulse signal causes the intake valve of each bank to rotate. The actual valve timing (not shown) (actual camshaft displacement angle) is calculated.

Specifically, each detection pulse signal from the crank position sensor 42 and the cam shaft position sensor 44 is input to the microcomputer 48 of the engine control device 46, and based on this, the actual valve timing (actual cam shaft displacement). Angle) is calculated. Although not shown, various detection signals output from various sensors that detect the engine operating state, such as an intake air amount sensor, a water temperature sensor, and a throttle opening sensor, are also input to the microcomputer 48, and these various sensor data are output. Based on this, the target valve timing (target camshaft displacement angle) of the intake valve is calculated.

Further, the microcomputer 48 performs each routine so that the actual valve timing (actual camshaft displacement angle) of the intake valve of each bank matches the target valve timing (target camshaft displacement angle). The current (control current) flowing through the linear solenoid 64 that is the drive source of the hydraulic control valve 30 of the phase difference adjusting device 40 of the bank is calculated by PD control described later, and the signal of the control current is output circuit in the engine control device 46. By outputting from 49 to the linear solenoid 64 of each bank, the hydraulic control valve 30 of each bank
Of the intake valves (not shown) of each bank to control the actual valve timing (actual camshaft displacement angle) to match the target valve timing (target camshaft displacement angle).

Next, the structure of the phase difference adjusting device 40 of each bank will be described in detail with reference to FIG. The phase difference adjusting device 40 of each bank has exactly the same configuration.

The phase difference adjusting device 40 is attached to the cylinder head 25 of the engine 1. The phase difference adjusting device 40 includes a cam shaft sleeve 11 having a substantially cylindrical shape. The cam shaft sleeve 11 has a large-diameter cylindrical portion coaxial with the left end portion of the intake side cam shaft 4 in FIG. It is fitted. The hollow partition wall 11c of the cam shaft sleeve 11 is connected to the end portion of the intake side cam shaft 4 by press fitting the pin 12 and tightening the bolt 10. As a result, the cam shaft sleeve 11 rotates integrally with the intake cam shaft 4. An external tooth helical spline 11 a is formed on the outer peripheral surface of the large diameter cylindrical portion of the camshaft sleeve 11.

Further, the camshaft sleeve 11 is provided with a small-diameter cylindrical portion 11b, and the small-diameter cylindrical portion 11b extends coaxially in the substantially cylindrical hollow portion of the housing 23. The housing 23 is attached to the cylinder head 25 by tightening bolts 24 at its flange portion 23a.

The sprocket 13a is rotatable relative to the intake-side camshaft 4 while being sandwiched between the annular rib 4a of the intake-side camshaft 4 and the open end of the large-diameter cylindrical portion of the camshaft sleeve 11. Is coaxially supported. On the left side surface of the sprocket 13a in FIG. 2, a sprocket sleeve 15 having a substantially cylindrical shape is provided with a pin 14 through each flange portion.
It is mounted coaxially by press-fitting and bolt 16 fastening. This causes the sprocket sleeve 15 to rotate integrally with the sprocket 13a. The sprocket sleeve 15 has a cylindrical portion 15b, and the cylindrical portion 15b extends coaxially into the hollow portion of the housing 23 so as to surround the camshaft sleeve 11.

At the intermediate portion of the inner peripheral surface of the cylindrical portion 15b,
An internal tooth helical spline 15a is formed, and the internal tooth helical spline 15a is formed on the camshaft sleeve 11
The external tooth helical spline 11a is formed to have a twist in the opposite direction. Either one of the external tooth helical spline 11a and the internal tooth helical spline 15a may be formed by a spline having linear teeth parallel to the axial direction with a twist angle of zero.

An annular space 90 having a substantially uniform cross section in the axial direction is formed between the small-diameter cylindrical portion 11b of the camshaft sleeve 11 and the cylindrical portion 15b of the sprocket sleeve 15 described above. Within 90, a substantially cylindrical hydraulic piston 17 is coaxially supported by the camshaft sleeve 11 so as to be slidable axially and liquid-tightly.

On the right side of the inner peripheral surface of the hydraulic piston 17, the external tooth helical spline 11 of the camshaft sleeve 11 is provided.
An internal tooth helical spline 17a that meshes with a is formed. On the other hand, on the right side of the outer peripheral surface of the hydraulic piston 17,
Internal tooth helical spline 1 of sprocket sleeve 15
An external tooth helical spline 17b that meshes with 5a is formed. As a result, each of these splines 17a,
The timing chain 3a
The rotation of the crankshaft 2 transmitted to the sprocket 13a via (see FIG. 1) is transmitted to the intake side camshaft 4 via the sprocket sleeve 15, the hydraulic piston 17 and the camshaft sleeve 11.

An oil seal 70 is provided on the outer peripheral edge of the annular flange portion formed on the left end portion of the hydraulic piston 17 in the annular space 90.
It is mounted so as to make liquid-tight contact with the inner peripheral surface of the.
The annular leg portion 17c formed so as to extend in an L-shaped cross section inside the left side portion of the inner surface of the hydraulic piston 17 is a central step portion of the camshaft sleeve 11 (hereinafter referred to as "right side stopper").
And the rightward movement to the hydraulic piston 17 is stopped.

By providing the hydraulic piston 17 in the annular space 90 as described above, the annular space 90 is
Is divided into two chambers. As a result, the advance side hydraulic chamber 2
2 is formed on the left side of the hydraulic piston 17, while the retard side hydraulic chamber 32 is formed on the right side of the flange portion of the hydraulic piston 17. The seal between the hydraulic chambers 22 and 32 is secured by the oil seal 70 described above.

An end plate 50 is coaxially attached to the left end opening of the sprocket sleeve 15. The end plate 50 is formed in a reverse L-shape in cross section by a cylindrical portion and an annular flange portion, and the annular flange portion of the end plate 50 is coaxially fixed to the left end opening of the sprocket sleeve 15. An annular groove is formed on the outer peripheral surface of the cylindrical portion of the end plate 50, and an oil seal 71 is mounted in the annular groove. The annular flange portion of the end plate 50 abuts on the annular flange portion of the hydraulic piston 17 to stop the hydraulic piston 17 from moving to the left (hereinafter referred to as “left side stopper”).
It also plays a role as.

End plate 50 and camshaft sleeve 1
On the left side of FIG. 1, a ring plate 51 formed in an annular shape having a U-shaped cross section is coaxially attached to the camshaft sleeve 11 on the inner surface of the annular left side wall of the housing 23 by press fitting the knock pin 53. The cylindrical portion of the end plate 50 and the small diameter cylindrical portion 11b of the camshaft sleeve 11 are rotatably supported in the U-shaped right side surface of the ring plate 51.

An oil seal 72 is mounted in an annular groove formed on the outer peripheral surface of the small diameter side cylindrical portion of the ring plate 51. The oil seal 72 is provided between the ring plate 51 and the camshaft sleeve 11. To secure the sealing property of. On the other hand, the oil seal 71 described above ensures the sealing performance between the end plate 50 and the ring plate 51. As a result, the sealing property in the advance side hydraulic chamber 22 is secured.

A bolt 52 is coaxially fitted into the small-diameter cylindrical portion of the ring plate 51 and the hollow portion of the annular left side wall of the housing 23. The bolt 52 has a small diameter of the camshaft sleeve 11 at its right end surface. A cylindrical space 91 is formed between the inner peripheral surface of the cylindrical portion and the hollow partition wall 11c.
Further, inside the bolt 52, the hydraulic passage 61b has a section T
The hydraulic passage 61b is formed in a V shape and communicates with the inside of the cylindrical space 91 at its axial passage portion. The hydraulic passage 61b communicates with the annular groove formed on the outer peripheral surface of the bolt 52 at both ends of the radial passage portion.

A hydraulic passage 61a is formed in the left wall portion of the housing 23, and the hydraulic passage 61a is formed.
Is communicated with the inside of the cylindrical space 91 through the annular groove of the bolt 52 and the hydraulic passage 61b, and further passes through the hydraulic passage 61c formed in the camshaft sleeve 11 so as to open in the cylindrical space 91 and is retarded. It communicates with the side hydraulic chamber 32. A hydraulic passage 60 communicating with the advance-side hydraulic chamber 22 is formed in the left wall portion of the housing 23. These hydraulic passages 6
1a and 60 are formed in a left wall portion of the housing 23 and open in a cylindrical hollow portion 95 that accommodates a hydraulic control valve 30 described later. Further, in the cylindrical hollow portion 95,
The hydraulic pressure supply passage 65 is open at the tip thereof, and the hydraulic pressure supply passage 65 supplies the working oil pressure-fed by the oil pump 29 from the oil pan 28 of the engine 1 to the cylindrical hollow portion 9.
Supply within 5. The oil pressure release passage 66 is connected to the oil pan 2
8 is opened to return the hydraulic oil into the oil pan 28.

Next, the structure of the hydraulic control valve 30 will be described with reference to FIG. The hydraulic control valve 30 is a spool including a cylinder 30a formed by the inner wall of the cylindrical hollow portion 95, and a spool 31 having a pair of left and right lands fitted in the cylinder 30a so as to be slidable in the axial direction. It is a valve. The cylinder 30a includes a hydraulic port 30b communicating with the hydraulic passage 61a and a hydraulic port 60.
And a hydraulic port 30c communicating with.

Further, the cylinder 30a has a suction port 30d communicating with the hydraulic pressure supply passage 65 and a hydraulic pressure release passage 66.
Both discharge ports 30e, 30f communicating with the. The switching of the communication between the hydraulic ports 30b and 30c, the suction port 30d, and the discharge ports 30e and 30f and the control of the communication degree (opening of the hydraulic control valve 30) are performed by sliding the spool 31 in the cylinder 30a. Done. A coil spring 31a is interposed in the right end portion of the cylinder 30a in FIG. 3 on the right end side of the spool 31.
Always urges the spool 31 to the left in the drawing.

On the other hand, a linear solenoid 64 is provided on the left end side of the spool 31 in the figure, inside the left end portion of the cylinder 30a in the figure. When an electromagnetic force is induced in the linear solenoid 64 according to the value of the current flowing through the linear solenoid 64, the electromagnetic force causes the spool 31 to slide to the right against the biasing force of the coil spring 31a. Has become.

Hereinafter, the hydraulic control valve 30 configured as above will be described.
The switching of communication of each hydraulic passage and the opening degree control by sliding the spool 31 will be described.

As shown in FIG. 3A, the spool 31
However, when an electromagnetic force is received from the linear solenoid 64 and slides to the right against the urging force of the coil spring 31a, the suction port 30d and the hydraulic port 30c communicate with each other by the right movement of the right side land of the spool 31 and the hydraulic pressure. Supply path 65
Communicates with the hydraulic passage 60. Therefore, the oil pressure from the oil pump 29 is pumped into the advance side hydraulic chamber 22.
At the same time, the discharge port 30e and the hydraulic port 30b are communicated by the rightward movement of the left land of the spool 31 to communicate the hydraulic passage 61a and the hydraulic release passage 66. Therefore, the hydraulic pressure in the retard side hydraulic chamber 32 is released.
As a result, the hydraulic piston 17 is pushed to the right in the annular space 90 (see FIG. 2), so that the intake camshaft 4 rotates and advances relative to the sprocket 13a and then the crankshaft 2.

As shown in FIG. 3 (b), when the spool 31 is located at the center, the communication of the hydraulic port 30b with the discharge port 30e and the communication of the hydraulic port 30c with the suction port 30d are carried out by the spool 31. It is blocked by the lands on both sides. Therefore, when there is no leakage of hydraulic oil from the hydraulic chambers 22 and 32 on the advance side and the retard side,
The position of the hydraulic piston 17 is maintained as it is. Therefore, the rotational phase difference between the sprocket 13a and the intake camshaft 4, that is, the actual valve timing does not change.

On the other hand, as shown in FIG. 3C, when the spool 31 is biased by the coil spring 31a and slides to the left while the generation of the electromagnetic force from the linear solenoid 64 is stopped, the suction port 30b is moved. The hydraulic port 30d communicates with the left land of the spool 31 in the leftward direction to communicate the hydraulic pressure supply passage 65 with the hydraulic passage 61a. Therefore, the oil pressure from the oil pump 29 is supplied to the retard side hydraulic chamber 32. On the other hand, the discharge port 30f and the hydraulic port 30c communicate with each other by moving the right side land of the spool 31 to the left, and the hydraulic passage 60 and the hydraulic release passage 66 are connected.
And communicate with. Therefore, the hydraulic pressure in the advance side hydraulic chamber 22 is released. As a result, the hydraulic piston 17 is pushed to the left in the annular space 90, so that the intake side camshaft 4 rotates in the opposite direction to the above, and the retard angle is relatively retarded with respect to the sprocket 13a and then the crankshaft 2. To do.

3A, 3B and 3C, the port 30b and the port 30e (or the port 30) are used.
d) the degree of communication with port 30c and port 30d
(Or port 30f), the degree of communication with the spool 31
Is controlled by the opening degree of each land on both the left and right sides with respect to each port 30b and 30c accompanying the rightward movement (or leftward movement).

FIG. 4 is a characteristic diagram showing the relationship between the position of the spool 31 in the hydraulic control valve 30 (hereinafter referred to as "spool position") and the actual valve timing change speed under certain operating conditions of the engine 1. In this characteristic diagram, the area where the actual valve timing change rate is positive (+) corresponds to the area moving in the advance direction, while the area where the actual valve timing change rate is negative (-) is It corresponds to the area moving in the retard direction. The spool position on the horizontal axis in this characteristic diagram is proportional to the linear solenoid current.

In this characteristic diagram, each symbol (a),
3B and 3C respectively show spool positions corresponding to the positions of the spool 31 shown in FIGS. 3A, 3B, and 3C. The linear solenoid current at the point where the actual valve timing does not change as shown by the symbol (b) is called "holding current". The phase difference adjusting device 40 can be controlled by increasing the linear solenoid current when advancing the valve timing based on this holding current and decreasing the linear solenoid current when retarding the valve timing. it can.
The holding current is learned during engine operation.

By the way, in the conventional V-type engine, the load of the high-pressure fuel pump is applied to the intake side camshaft in only one bank (for example, B bank), and the intake side camshaft in the other bank (for example, A bank). Since no load is applied to such a high-pressure fuel pump, the load torques of the intake-side camshafts of the two banks are unbalanced by the load of the high-pressure fuel pump. In particular, an in-cylinder injection type engine requires a fuel injection pressure of 30 times or more as compared with an intake port injection type engine, so a large camshaft torque is required to drive a high pressure fuel pump, resulting in a 2 There will be a large difference in the load torque of the intake camshafts of the two banks.

Generally, the responsiveness of the change of the actual valve timing (actual camshaft displacement angle) to the change of the target valve timing (target camshaft displacement angle), that is, the responsiveness of the valve timing control is the load torque of the intake side camshaft. As the load torque of the intake side camshafts of the two banks greatly differs, the response of valve timing control (camshaft displacement) between the two banks increases as shown in FIGS. 4 and 5A. Velocity) is greatly different, and as a result, the internal EGR amount and the charging efficiency of fresh air become unbalanced between the two banks during transient operation in which the target valve timing suddenly changes, resulting in torque fluctuation and deterioration of drivability. There's a problem.

Therefore, in the present embodiment, the high-pressure fuel pumps 20 having the same capacity are installed in the banks A and B, respectively, so that the loads of the high-pressure fuel pumps 20 on the intake side camshafts 4 of the banks are equalized. It is configured to join. As a result, the load torque of the intake side camshaft 4 of each bank can be made substantially equal, and as shown in FIG.
The responsiveness of the valve timing control of each bank can be made uniform, and problems such as torque fluctuations and drivability deterioration during transient operation due to deviations in the responsiveness can be solved.

However, the load torque of the intake side camshaft 4 of each bank varies due to variations in the manufacture of the high-pressure fuel pump 20, variations in the manufacture of the phase difference adjusting device 40, their changes over time, operating conditions, and the like. Since the responsiveness of the valve timing control of the banks may be slightly shifted, in the present embodiment, as shown in FIG. 6, a function (correction means) for independently correcting the valve timing control gain for each bank.
By providing the above, it is possible to reduce the deviation of the responsiveness of the valve timing control of each bank caused by manufacturing variations, changes over time, operating conditions, and the like.

FIG. 6 is a block diagram showing the functions realized by the routines, which will be described later, executed by the microcomputer 48. The outline of the valve timing control will be described below with reference to FIG.

The target displacement angle (target valve timing) is
It is calculated based on signals output from various sensors such as an intake air amount sensor, a water temperature sensor, and a throttle opening sensor that detect engine operating conditions, and the same target displacement angle is used for both bank A and bank B. The valve timing control (current control of the linear solenoid 64) of each bank is executed by PD control. The proportional term (P) gain of the PD control is calculated by a map or the like based on, for example, the engine rotation speed Ne and the oil temperature of the hydraulic oil (engine oil), and similarly, the differential term (D) gain is also calculated, for example, at the engine rotation speed. It is calculated by a map or the like based on the speed Ne and the oil temperature.

For each bank, the P term correction amount obtained by multiplying the deviation between the target displacement angle and the actual displacement angle by the proportional term (P) gain, and the camshaft displacement speed (de / dt) differential term (D). The feedback correction amount is obtained by adding the D term correction amount obtained by multiplying the gain, and this is converted into a current value to obtain the correction current. Then, add this correction current to the holding current,
The control current of the hydraulic control valve 30 of each bank is calculated.

Next, the processing contents of each routine for executing the above-mentioned valve timing control of the two banks will be described.

[Main Routine] The main routine shown in FIG. 7 is executed by the microcomputer 48 at every predetermined crank angle (for example, 1 for a 6-cylinder engine) during engine operation.
Repeatedly every 20 ° C). When this routine is started, first, in step 100, A in FIG.
After the bank control current calculation routine is executed to calculate the control current Id of the A bank, the routine proceeds to step 200, where the B bank control current calculation routine (not shown) is executed to calculate the control current of the B bank. The B bank control current calculation routine has the same processing contents as the A bank control current calculation routine of FIG. 8 described below.

[A Bank Control Current Calculation Routine] When the A bank control current calculation routine of FIG. 8 is started in step 100 of the main routine shown in FIG. 7, first, in steps 110 and 111, the target displacement angle VTT and the actual value are calculated. The displacement angle VT is read, and in the next step 112, the deviation PT between them is calculated. PT = VTT-VT

After that, the routine proceeds to step 113, where the actual displacement angle change amount D per predetermined crank angle (sampling interval)
T (information proportional to the camshaft displacement speed) is calculated. DT = VT-VT (i-1) where VT (i-1) is the actual displacement angle sampled last time.

Then, in the next step 114, the P gain calculation subroutine of FIG.
After calculating the P gain PG, the process proceeds to step 115, and the D gain DG is calculated by executing the D gain calculation subroutine of FIG. 11 described later.

After that, the routine proceeds to step 116, where the P term correction amount (PT
× PG) and the D term correction amount (DT × DG) obtained by multiplying the actual displacement angle change amount DT by the D gain DG to obtain the feedback correction amount FD, which is calculated as the feedback correction amount − The correction current If is obtained by converting the current value using the current value conversion table. FD = PT × PG + DT × DG If ← FD (current conversion)

The feedback correction amount-current value conversion table of FIG. 9 is set from the data obtained by previously measuring the relationship between the current value of the linear solenoid 64 and the actual valve timing change speed (actual camshaft displacement speed). ing.

After that, the routine proceeds to step 117, where the correction current If is added to the holding current Ih to obtain the control current I of the A bank.
Find d. Id = If + Ih

[P gain calculation subroutine] When the P gain calculation subroutine of FIG. 10 is started in step 114 of FIG. 8, first, in step 120, the base P gain using the displacement angle deviation PT and the engine rotation speed Ne as parameters. Search the map of PBAS to find the current displacement angle deviation PT
And base P gain PBA according to the engine speed Ne
Calculate S. The map of the base P gain PBAS is created in advance by experiments or simulations and is stored in the ROM of the microcomputer 48.

After that, the routine proceeds to step 121, where the table of the oil temperature correction coefficient PCMP having the oil temperature as a parameter is searched to calculate the oil temperature correction coefficient PCMP corresponding to the current oil temperature. The reason why the oil temperature correction coefficient PCMP is used is that the driving force of the phase difference adjusting device 40 is hydraulic pressure, the viscosity of the hydraulic oil changes depending on the oil temperature, and the responsiveness of the phase difference adjusting device 40 itself changes. This is because the leak amount changes, and the friction loss changes depending on the temperature of the phase difference adjusting device 40 that changes according to the oil temperature. The table of the oil temperature correction coefficient PCMP is also created in advance by experiments, simulations, etc.
8 ROMs.

Then, in the next step 122, the base P
The final P gain PG is obtained by multiplying the gain PBAS by the oil temperature correction coefficient PCMP. PG = PBAS × PCMP

[D Gain Calculation Subroutine] When the D gain calculation subroutine of FIG. 11 is started in step 115 of FIG. 8, first, in step 130, the actual displacement angle change amount DT.
D with the engine and engine speed Ne as parameters
The map of the gain DBAS is searched to calculate the base D gain DBAS corresponding to the current actual displacement angle change amount DT and the engine rotation speed Ne. The map of the base D gain DBAS is created in advance by experiments or simulations and is stored in the ROM of the microcomputer 48.

After that, the routine proceeds to step 131, where the table of the oil temperature correction coefficient DCMP having the oil temperature as a parameter is searched to calculate the oil temperature correction coefficient DCMP corresponding to the current oil temperature. This oil temperature correction coefficient DCMP table is also created in advance by experiments or simulations and is stored in the ROM of the microcomputer 48.

Then, in the next step 132, the base D
The final D gain DG is obtained by multiplying the gain DBAS by the oil temperature correction coefficient DCMP. PG = DBAS × DCMP Incidentally, the control current of the B bank is also calculated by the same method as the control current of the A bank.

In the present embodiment described above, the high-pressure fuel pumps 20 of the same capacity are respectively provided in the banks A and B.
Since the load of the high-pressure fuel pump 20 is evenly applied to the intake side camshafts 4 of each bank, the load torque of the intake side camshafts 4 of each bank can be made substantially equal. As shown in FIG. 5B, the responsiveness of valve timing control of each bank can be made uniform, and problems such as torque fluctuation during transient operation and deterioration of drivability due to deviation of the responsiveness can be solved. be able to.

In this case, the high-pressure fuel pump 20 of each bank
Load (discharging amount) changes according to the target fuel discharging amount during engine operation, but it is not necessary to control the high-pressure fuel pump 20 of each bank independently for each bank, and the high-pressure fuel pump of each bank does not need to be controlled independently. Since it suffices to control the fuel pump 20 in the same manner, it is possible to maintain the load balance state of the auxiliary machinery (high-pressure fuel pump 20) of each bank in the entire operation range, and to control the valve timing between the banks. It is possible to surely reduce the deviation of responsiveness in the entire operation region.

However, the present invention is not limited to the configuration in which the high-pressure fuel pump 20 having the same capacity is installed in each bank, and the high-pressure fuel pump is installed in one bank and the high-pressure fuel pump is installed in the other bank. Auxiliary machinery other than the high-pressure fuel pump, which has the same load, may be installed, or auxiliary machinery other than the high-pressure fuel pump, which has substantially the same load, may be installed in each bank. good. As auxiliary machinery other than the high-pressure fuel pump, for example, a vacuum pump for securing a negative pressure of the brake booster, a water pump,
There is an oil pump for power steering, etc., and these accessories may be driven by a cam shaft.

The point is that the accessories driven by the camshafts of each bank may be installed so that the loads on the accessories in each bank are approximately equal. At this time, a plurality of auxiliaries are installed in one bank, and the total load of the plurality of auxiliaries is almost the same as the load of one auxiliaries in the other bank (or the total load of the plurality of auxiliaries). You may make it install evenly. In this way, the load of the auxiliary machinery is applied to the camshafts of each bank almost evenly, so that the load torque of the camshafts of each bank can be made almost equal, and the valve timing control of each bank can be controlled. It is possible to reduce the shift in responsiveness.

Although the above-described embodiment applies the present invention to the valve timing control of the intake valve, it goes without saying that the present invention can also be applied to the valve timing control of the exhaust valve. In the above embodiment, the hydraulic actuator is used as the power source of the phase difference adjusting device 40, but it goes without saying that an electric actuator may be used.

In addition, the present invention is not limited to the V-type engine, but can be applied to various engines including a plurality of banks (a plurality of cylinder groups) such as a horizontally opposed engine, and the fuel injection system can be a cylinder. The injection method is not limited to the internal injection method, and may be the intake port injection method.

[Brief description of drawings]

FIG. 1 is a schematic configuration diagram of an entire system showing an embodiment of the present invention.

FIG. 2 is a sectional view showing the configuration of a phase difference adjusting device.

FIG. 3 is a sectional view for explaining the operation of the hydraulic control valve.

FIG. 4 is a characteristic diagram showing a relationship between an actual valve timing change speed and a spool position.

5A is a diagram illustrating a difference in responsiveness between a conventional A bank and a B bank, and FIG. 5B is a diagram illustrating responsiveness between the A bank and the B bank according to the present embodiment.

FIG. 6 is a functional block diagram of a control system of this embodiment.

FIG. 7 is a flowchart showing a processing flow of a main routine.

FIG. 8 is a flowchart showing a processing flow of an A bank control current calculation routine.

FIG. 9 is a diagram conceptually showing an example of a table for converting a feedback correction amount FD into a correction current If.

FIG. 10 is a flowchart showing a processing flow of a P gain calculation subroutine.

FIG. 11 is a flowchart showing a processing flow of a D gain calculation subroutine.

[Explanation of symbols]

DESCRIPTION OF SYMBOLS 1 ... Engine (internal combustion engine), 2 ... Crank shaft, 4 ... Intake side cam shaft, 5 ... Exhaust side cam shaft, 17 ... Hydraulic piston, 2
0 ... High-pressure fuel pump (auxiliaries), 21 ... Fuel injection valve, 2
2 ... Advance side hydraulic chamber, 28 ... Oil pan, 29 ... Oil pump, 30 ... Hydraulic control valve, 32 ... Delay side hydraulic chamber, 40 ...
Phase difference adjusting device (valve timing adjusting means), 42 ...
Crank position sensor, 44 ... Cam shaft position sensor, 46 ...
Engine control device, 48 ... Microcomputer (correction means).

Claims (4)

[Claims] 1. A rotation phase of an intake-side and exhaust-side camshaft provided for each cylinder group of an internal combustion engine including a plurality of cylinder groups, and at least one of intake-side and exhaust-side camshafts of each cylinder group. In an internal combustion engine provided with valve timing adjusting means for advancing or retarding the valve timing of at least one of the intake side and the exhaust side of each cylinder group by advancing or retarding with respect to the rotational phase of the crankshaft. An internal combustion engine in which an auxiliary machine driven by a cam shaft of each cylinder group is installed in each of the plurality of cylinder groups such that loads of the auxiliary machines are substantially equalized in each cylinder group. . 2. The internal combustion engine according to claim 1, wherein the accessories installed in each of the plurality of cylinder groups are high-pressure fuel pumps having the same capacity. 3. An auxiliary machine installed in at least one cylinder group is a high-pressure fuel pump for boosting and discharging fuel, and an auxiliary machine installed in another cylinder group is the high-pressure fuel pump. The internal combustion engine according to claim 1, which is an auxiliary machine other than a high-pressure fuel pump that has substantially the same load. 4. The internal combustion engine according to claim 1, further comprising a correcting unit that corrects a control amount of the valve timing adjusting unit for each of the plurality of cylinder groups.