JPH11148380A - Hydraulic valve timing control system - Google Patents

Abstract

PROBLEM TO BE SOLVED: To prevent oil leakage from the system of an actuator even after stopping of an engine so that oil-pressure of the actuator may be prevented from dropping rapidly by constituting a fluid control means so as to conduct switching-control between normal control at the time of inputting engine power supply, and holding control at the time of the engine power supply shutdown. SOLUTION: Whether power supply is shutdown or not is detected by a power supply shutdown means M14 at the time of engine stop, a fluid supply means M10 is operated for a fixed time from a power supply shutdown time obtained by the detected results, and the fluid supply means M10 is controlled at a holding position by a holding control means M13 so as to hold the oil- pressure of an actuator 9, and the fluid supply means M10 is switching-controlled between normal control at the time of inputting engine power supply, and holding control at the time of engine power supply shutdown by a switching means M15.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a hydraulic valve timing adjusting system for changing the opening / closing timing of one or both of an intake valve and an exhaust valve according to the operating conditions of an engine.

[0002]

2. Description of the Related Art As a conventional hydraulic valve timing adjusting system, when a camshaft is driven by a timing pulley or a chain sprocket that rotates synchronously with an engine crankshaft, a vane type valve is provided between the timing pulley and the camshaft. By providing a timing mechanism and supplying hydraulic oil from an oil pump via an oil control valve (hereinafter referred to as OCV) to an actuator that drives the valve timing mechanism with hydraulic oil, the camshaft is moved relative to the crankshaft. By rotating the camshaft with respect to the crankshaft, the opening and closing timing of the intake and exhaust valves is shifted with respect to the engine rotation to reduce exhaust gas and improve fuel efficiency. What is intended is, for example, Japanese Unexamined Patent Publication No. 7-1. 9319, JP-A No. 7-13
No. 9320, JP-A-8-28219, JP-A-8-121122, JP-A-9-60507, JP-A-9-60508 and the like.

[0003]

Since the conventional hydraulic valve timing adjusting system is configured as described above, the OCV between the actuator and the oil pump remains open immediately after the engine is stopped. As a result, the hydraulic oil accumulated inside the actuator may leak from the OCV when the engine is stopped. As a result, if the hydraulic pressure of the actuator system rapidly decreases immediately after the engine is stopped, the rotational inertia of the crankshaft and the camshaft Due to the spring reaction force, the rotor of the actuator moves to the advanced side, and the rotor cannot be held at the most retarded position.At the time of subsequent engine startup, it takes time for hydraulic oil to be supplied from the oil pump to the actuator. As a result, cam angle control at the time of engine start The rotor there is a problem, such as issuing a noise causing hunting.

SUMMARY OF THE INVENTION The present invention has been made to solve the above-mentioned problem, and the hydraulic oil of the actuator system is prevented from falling out of the OCV immediately after the engine is stopped, and the hydraulic pressure is applied to the actuator even after the engine is stopped. It is an object of the present invention to obtain a hydraulic valve timing adjustment system capable of maintaining a state. Another object of the present invention is to provide a hydraulic valve timing adjustment system in which the supply pressure of hydraulic oil supplied from an oil pump to an actuator system does not decrease.

[0005]

As shown in FIG. 1 which is a block diagram of the basic concept, the hydraulic valve timing adjusting system according to the present invention has a predetermined timing synchronized with the rotation of the engine M1. Driven, the combustion chamber M2
An intake valve M5 and an exhaust valve M6 for opening and closing an intake passage M3 and an exhaust passage M4 respectively, an operating state detecting means M7 for detecting an operating state of the engine M1, and a target valve timing for the operating state of the engine. A target valve timing calculating means M8 for calculating based on a detection result of the operating state detecting means, a valve timing variable actuator M9 for changing at least one of the intake valve M5 and the exhaust valve M6 valve opening / closing timing, Fluid supply means M10 capable of supplying hydraulic oil to the actuator M9 and controlling the flow rate thereof to drive the actuator M9, and the intake valve M5 provided with the actuator M9.
And actual valve timing detecting means M11 for detecting the actual valve timing of at least one of the exhaust valves M6.
And an actual valve timing control means M12 for controlling the actuator M9 by controlling the fluid supply means M10 in order to change the actual valve timing to the target valve timing. In the control system, a power cutoff detecting means M14 for detecting a power cutoff time when the engine M1 is stopped, and the fluid supply means M10 for a predetermined time from the power cutoff time based on the detection result of the power cutoff detecting means M14
Is operated to hold the hydraulic pressure of the actuator M9 so as to control the fluid supply means M10 to a holding position, and to control the fluid supply means M10 to normal control when the engine power is turned on and to shut off the engine power. And switching means M15 for switching to the time keeping control.

In the hydraulic valve timing adjusting system according to the present invention, when the engine speed after the engine power is cut off is equal to or less than the predetermined speed, the holding control means M13 controls the actuator M9 until the engine speed reaches the predetermined speed.
So that the fluid supply means M10 is maintained at the most retarded position.
Is to operate.

In the hydraulic valve timing adjusting system according to the present invention, the holding control means M13 controls the fluid supply means M10 to the holding position when the engine speed after the engine power is cut off is equal to or higher than a predetermined speed. .

[0008] In the hydraulic valve timing adjusting system according to the present invention, the power cutoff detecting means M14 cuts off the main relay of the engine power supply system after a predetermined time has elapsed after the engine power cutoff.

[0009]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS One embodiment of the present invention will be described below. Embodiment 1 FIG. FIG. 2 is a schematic sectional view showing a gasoline engine system provided with a hydraulic valve timing adjusting system according to Embodiment 1 of the present invention. In the drawing, reference numeral 1 denotes an engine (M1 in FIG. 1), which is composed of a plurality of cylinders and one cylinder of which is shown. Reference numeral 2 denotes a cylinder block which forms a plurality of cylinders of the engine 1. Reference numeral 3 denotes an upper portion of the cylinder block 2. 4 is a piston that reciprocates up and down in each cylinder of the cylinder block 2, and 5 is a crankshaft connected to the lower end of the piston 4.
It is driven to rotate by the up and down movement of.

Reference numeral 6 denotes a crank angle sensor disposed in the vicinity of the crankshaft 5, which detects the rotational speed NE of the engine 1 and that the crankshaft 5 is at a predetermined crank angle. Reference numeral 7 denotes a signal rotor connected to the crankshaft 5. Two teeth are formed on the outer periphery of the signal rotor 7 at every 180 °, and each time these teeth pass in front of the crank angle sensor 6. , Crank angle sensor 6
Generates a pulse-like crank angle detection signal.

Reference numeral 8 denotes a combustion chamber (M2 in FIG. 1) for burning the air-fuel mixture, which is defined by the inner walls of the cylinder block 2 and the cylinder head 3 and the top of the piston 4. Reference numeral 9 denotes a spark plug for igniting an air-fuel mixture in the combustion chamber 8, which is disposed at the top of the cylinder head 3 and protrudes into the combustion chamber 8. 10 is the cylinder head 3
A distributor 11 connected to an exhaust camshaft 20 described later is an igniter for generating a high voltage. Here, each spark plug 9 is connected to a distributor 10 via a high-pressure cord (not shown), and the high voltage output from the igniter 11 is synchronized with the rotation of the crankshaft 5 by the distributor 10 so that each ignition plug 9 It is distributed to the plug 9.

Reference numeral 12 denotes a water temperature sensor disposed in the cylinder block 2. The water temperature sensor 12 detects the temperature (cooling water temperature) THW of the cooling water flowing through the cooling water passage. 13
Is an intake port provided in the cylinder head 3, 14 is an exhaust port provided in the cylinder head 3, 15 is an intake passage (M3 in FIG. 1) connected to the intake port 13, and 16 is connected to the exhaust port 14. Exhaust passage (M4 in FIG. 1),
Reference numeral 17 denotes an intake valve (M5 in FIG. 1) provided on the cylinder head 3 for opening and closing the intake port 13; and 18, an exhaust valve (M6 in FIG. 1) provided on the cylinder head 3 for opening and closing the exhaust port 14.

Reference numeral 19 denotes an intake side camshaft arranged above the intake valve 17, and 19a denotes an intake side camshaft 19
An intake cam, which is rotatably mounted on the intake cam 17 and opens and closes the intake valve 17, an exhaust camshaft 20 disposed above the exhaust valve 18, and a synchronous motor 20 a rotatably mounted on the exhaust camshaft 20. An exhaust cam for opening and closing the exhaust valve 18, an intake timing pulley 21 mounted on one end of the intake camshaft 19, an exhaust timing pulley 22 mounted on one end of the exhaust camshaft 20, and 23 each This is a timing belt that links the timing pulleys 21 and 22 with the crankshaft 5.

Therefore, when the engine 1 is running, the camshafts 19, 20 are transmitted from the crankshaft 5 via the timing belt 23 and the timing pulleys 21, 22.
The rotation driving force is transmitted to the camshafts 19 and 20, and the cams 19a and 20a rotate integrally with the respective camshafts 19 and 20 to open and close the intake valve 17 and the exhaust valve 18, respectively. Driven at a predetermined opening / closing timing in synchronization with the rotation of the shaft 5 and the vertical movement of the piston 4, that is, in synchronization with a series of four steps of the engine 1 including an intake step, a compression step, an explosion / expansion step, and an exhaust step. You.

Reference numeral 24 denotes a cam angle sensor arranged near the intake side camshaft 19, which detects the opening / closing timing of the intake valve 17 (so-called valve timing). 2
Reference numeral 5 denotes a signal rotor connected to the intake-side camshaft 19. Four teeth are formed on the outer periphery of the signal rotor 25 every 90 °, and these teeth pass in front of the cam angle sensor 24. Each time, the cam angle sensor 24 generates a pulse-like cam angle detection signal.

Reference numeral 26 denotes a throttle valve arranged in the middle of the intake passage 15. The throttle valve 26 is opened and closed in conjunction with an accelerator pedal (not shown) to adjust the amount of intake air. 27 is a throttle sensor connected to the throttle valve 26,
The throttle opening TVO is detected. Reference numeral 28 denotes an intake air amount sensor disposed on the upstream side of the throttle valve 26, and an air flow amount (intake air amount) Q to be taken into the engine 1.
A is detected. Reference numeral 29 denotes a surge tank formed on the downstream side of the throttle valve 26 for suppressing intake pulsation.
Reference numeral 0 denotes an injector that is disposed near the intake port 13 of each cylinder and supplies fuel to the combustion chamber 8.

Here, the injector 30 is constituted by an electromagnetic valve which is opened by energization, and is supplied with fuel under pressure from a fuel pump (not shown). Therefore, when the engine 1 is operating, air is taken into the intake passage 15 and at the same time, fuel is injected from each injector 30 toward the intake port 13. As a result, an air-fuel mixture is generated at the intake port 13, and the air-fuel mixture is sucked into the combustion chamber 8 with the opening of the intake valve 17 that is opened in the intake process.

Reference numeral 40 denotes a valve timing variable actuator (M9 in FIG. 1) which is connected to the intake side camshaft 19 and disposed. The actuator 40 is driven by the lubricating oil of the engine 1 as a working oil, thereby changing the displacement angle of the intake camshaft 19 with respect to the intake timing pulley 21 to change the opening / closing timing (valve timing) of the intake valve 17. It is changed continuously, and its detailed configuration will be described later.

Reference numeral 80 denotes an OCV as fluid supply means for supplying hydraulic oil to the actuator 40 and adjusting the oil amount.
(M10 in FIG. 1), and the detailed configuration will be described later.

Reference numeral 100 denotes an electronic control unit (hereinafter referred to as an ECU).
The ECU 100 mainly includes an intake air amount sensor 28, a throttle sensor 27, a water temperature sensor 12,
Based on signals from the crank angle sensor 6 and the cam angle sensor 24, the injector 30, the igniter 11, the OCV 80
To control the fuel injection amount, the ignition timing, and the valve timing, and also control the valve closing timing of the OCV 80 after the IG switch, which will be described later, is turned off. The detailed configuration thereof will be described later.

FIG. 3 is a sectional view showing a hydraulic valve timing system according to Embodiment 1 of the present invention. In the figure, reference numeral 40 denotes an actuator for adjusting the valve timing of the intake valve 17, and the configuration of the actuator 40 will be described below. The same parts as those in FIG. 2 are denoted by the same reference numerals, and redundant description will be omitted. FIG.
, Reference numeral 41 denotes a bearing of the intake-side camshaft 19;
Is a housing of the actuator 40, and the housing 42 is rotatably attached to the intake camshaft 19. 43 is a case fixed to the housing 42, 44 is a bolt 45 on the intake side camshaft 19
A vane-type rotor that is connected and fixed in the case 43 and accommodated in the case 43, and the rotor 44 is rotatable relative to the case 43. Reference numeral 46 denotes a chip seal interposed between the case 43 and the rotor 44. The chip seal 46 prevents oil from moving between hydraulic chambers partitioned by the case 43 and the rotor 44.

Reference numeral 47 denotes a back spring disposed between the case 43 and the tip seal 46. The back spring 47 comprises a leaf spring for pressing the tip seal 46 against the rotor 44. 48 is a cover fixed to the case 43, 49 is a bolt for fastening and fixing the housing 42, the case 43 and the cover 48 together, 50 is an O-ring for preventing oil from leaking outside through a gap between the bolt 49 and the bolt hole, Reference numeral 51 denotes a plate fixed to the cover 48 with screws 52. Here, the housing 42 and the case 43
And the cover 48 constitute a housing member of the actuator 40.

Reference numeral 53 denotes an O-ring for preventing oil leakage interposed between the housing 42 and the case 43, and reference numeral 54 denotes a case 4
An O-ring 55 for preventing oil leakage interposed between the cover 3 and the cover 48 is a cylindrical holder provided on the rotor 44. The holder 55 is a concave portion 55a for engaging a plunger 56 described later. At one axial end. 5
Reference numeral 6 denotes a plunger slidably provided in the housing 42.
has a convex portion 56a for engaging with a. 57 is a spring for urging the plunger 56 toward the holder 55,
Numeral 58 denotes a plunger oil passage for introducing hydraulic oil into the holder 55. The hydraulic oil introduced into the concave portion 55a of the holder 55 from the plunger oil passage 58 moves the plunger 56 against the spring 57, thereby holding the holder. The lock of the plunger 56 with respect to 55 is released.

Reference numeral 59 denotes an air hole provided in the housing 42 for constantly setting the spring 57 side of the plunger 56 to atmospheric pressure. Reference numeral 60 denotes a shaft bolt for connecting and fixing the intake side camshaft 19 and the rotor 44 at the shaft center. The shaft bolt 60 is rotatable with respect to the cover 48. Reference numeral 61 denotes an air hole provided on the shaft bolt 60 and the intake side camshaft 19 to make the inside of the plate 51 equal to the atmospheric pressure.

Reference numeral 62 denotes a first oil passage provided in the intake camshaft 19 and the rotor 44. The first oil passage 62 communicates with a later-described retard hydraulic chamber 73 for moving the rotor 44 in the retard direction. doing. A second oil passage 63 is provided in the intake camshaft 19 and the rotor 44. The second oil passage 63 communicates with a later-described advance hydraulic chamber 74 for moving the rotor 44 in the advance direction. ing.

Next, the configuration of the OCV 80 for controlling the pressure of the hydraulic oil supplied to the actuator 40 configured as described above in FIG. 3 will be described. 81 is OCV80
82 (hereinafter referred to as a valve housing)
Is a spool that slides in the valve housing 81, 83 is a spring that urges the spool 82 in one direction, 84 is a linear solenoid that operates the spool 82 against the urging force of the spring 83, and 85 is a valve housing 8
1, a supply port (primary port) formed at 1; 86, an A port (secondary port) formed at the valve housing 81; 87, a B port (secondary port) formed at the valve housing 81; 88a and 88b are drain ports formed in the valve housing 81, 88 is a common drain pipe connected to the drain ports 88a and 88b, and 89 is a first pipe connecting the first oil passage 62 and the A port 86. , 9
Reference numeral 0 denotes a second pipe connecting the second oil passage 63 and the B port 87, 91 denotes an oil pan, 92 denotes an oil pump, and 93 denotes an oil filter.

Here, the oil pump 92 has a suction side communicating with the oil pan 91 and a discharge side connected to the supply port 85 via an oil filter 93. Also,
A drain pipe 88 communicates with the oil pan 91.

In the above, the oil pan 91, the oil pump 92 and the oil filter 93 constitute a lubricating device for lubricating various parts of the engine 1, and constitute a hydraulic oil supply device to the actuator 40 together with the OCV 80. I have.

FIG. 5 is a sectional view taken along the line X--X in FIG. 3, FIG. 6 is a partial sectional view showing the moving state of the slide plate in FIG. 5, and FIG. 7 is a line Y--Y in FIG. FIG. 8 is a sectional view taken along line ZZ of FIG. 3.
In these figures, 64 to 67 are first to fourth vanes protruding from the rotor 44 in the radial direction, and the tips of the vanes 64 to 67 are in sliding contact with the inner peripheral surface of the case 43. A tip seal 68 is provided at the leading end sliding contact portion.

Reference numeral 71 denotes a plurality of (four in the figure) equally spaced shoes protruding from the inner peripheral surface of the case 43, and reference numeral 72 denotes a bolt hole provided in the shoe 71. The inner bolt 49 is inserted. The protruding end of the shoe 71 comes into sliding contact with the vane support 69, which is the central portion of the rotor 44, and the tip sliding contact portion is provided with the tip seal 46 described in FIG.

Reference numeral 73 denotes a retard hydraulic chamber for rotating the first to fourth vanes 64 to 67 in the retard direction, and 74 denotes a retard hydraulic chamber for rotating the first to fourth vanes 64 to 67 in the advance direction. The retard hydraulic chamber 73 and the advance hydraulic chamber 74 are provided between the case 43 and the rotor 44.
3 is formed between the first shoe 71 and each of the vanes 64 to 67 of the rotor 44 in a fan shape.

A communication oil passage 75 is provided in the first vane 64 and communicates with the retard hydraulic chamber 73 and the advance hydraulic chamber 74 on both sides of the vane 64. A recess 76 is provided in the middle of the communication oil passage 75. The plunger oil passage 58 communicates in the middle of the moving groove 76. Reference numeral 77 denotes a slide plate that moves in the moving oil passage 76, and the communication oil passage 75 is divided by the slide plate 77, and a retard hydraulic chamber 73 is provided.
Oil leakage does not occur between the hydraulic fluid and the advance hydraulic chamber 74. Further, the slide plate 77 is provided with the retard hydraulic chamber 73.
When the hydraulic pressure of the lead hydraulic chamber 74 is high, it moves to the advance hydraulic chamber 74 side as shown in FIG. 6, and when the hydraulic pressure of the advance hydraulic chamber 74 is high, it moves to the retard hydraulic chamber 73 side as shown in FIG. is there. 78
Is provided in each of the vanes 64-67, and case 4
This is a tip seal that seals between 3 and each of the vanes 64-67 to prevent oil leakage. The arrows in FIGS. 5, 7, and 8 indicate the rotation direction of the entire actuator 40.

In the above, the retard hydraulic chamber 73 and the advance hydraulic chamber 74 are formed by the housing 42, the case 43, and the rotor 4.
4 and the cover 48, the retard hydraulic chamber 73 communicates with the first oil passage 62, and hydraulic oil is supplied from the first hydraulic passage 62, and the advance hydraulic chamber 74 is connected to the second hydraulic passage 74. Hydraulic oil is supplied from the second oil passage 63, and
The rotor 44 moves relative to the housing 42 in accordance with the amount of hydraulic oil supplied to the retard hydraulic chamber 73 and the advance hydraulic chamber 74, and each of the retard hydraulic chamber 73 and the advance hydraulic chamber 74 The volume changes.

Next, the actuator 40 and the OCV 80
Will be described. First, the rotor 44 in a state where the engine 1 is stopped is in a maximum retard position as shown in FIG. 5, that is, a position in which the rotor 44 is rotated to the maximum in the advance direction with respect to the housing 42. In the stop state, the hydraulic oil is not supplied to the first oil passage 62 and the second oil passage 63, and the hydraulic oil is not supplied to the plunger oil passage 58. ing. Therefore, the plunger 56 is pressed against the holder 55 by the urging force of the spring 57, and the plunger 56 and the holder 55 are engaged with each other, so that the housing 42 and the rotor 44 are locked.

When the engine 1 is started from this state, the oil pump 92 is operated, and the pressure of the working oil supplied to the OCV 80 is increased, so that the first pipe 89 and the first oil passage 89 are connected from the A port 86 of the OCV 80. Hydraulic oil is supplied to the retard hydraulic chamber 73 in the actuator 40 via 62. At this time, the slide plate 77 moves toward the advance hydraulic chamber 74 due to the hydraulic pressure in the retard hydraulic chamber 73, and the retard hydraulic chamber 73 and the plunger oil passage 58 communicate with each other.
6 is pushed to move to the housing 42 side, and the engagement between the plunger 56 and the rotor 44 is released.

However, since hydraulic oil is supplied to the retard hydraulic chamber 73, the vanes 64 to 6 of the rotor 44
7 is in a state of pressing against the shoe 71 in the retard direction,
Therefore, even if the engagement of the rotor 44 by the plunger 56 is released, the housing 42 and the rotor 44 are pressed against each other by the hydraulic pressure of the retard hydraulic chamber 73, and vibration and impact can be reduced or eliminated. As described above, the plunger 56 can be moved by the hydraulic pressure of the retard hydraulic chamber 73. Therefore, when the engine 1 is started and a predetermined hydraulic pressure (a hydraulic pressure at which the slide plate 77 and the plunger 56 can move) is generated, the plunger 56 is moved. When the engagement between the rotor 56 and the rotor 44 is released, when it is necessary to advance the rotor 44, it is possible to immediately respond.

Next, in order to advance the rotor 44, O
When the B port 87 of the CV 80 is opened, the second pipeline 90
Hydraulic fluid is supplied to the advance hydraulic chamber 74 through the second oil passage 63, and the hydraulic pressure is transmitted from the advance hydraulic chamber 74 to the communication oil passage 75 to press the slide plate 77,
The slide plate 77 moves to the retard hydraulic chamber 73 side.
By the movement of the slide plate 77, the plunger oil passage 58 communicates with the advance hydraulic chamber 74 side of the communication oil passage 75,
Hydraulic pressure is transmitted from the advance hydraulic chamber 74 to the plunger oil passage 58, and the plunger 56 is moved by the spring 57 by this hydraulic pressure.
Is moved toward the housing 42 side against the urging force, and the engagement between the plunger 56 and the holder 55 is released.

In the disengaged state, by adjusting the oil supply by opening and closing the A port 86 and the B port 87 of the OCV 80, the oil amounts of the retard hydraulic chamber 73 and the advance hydraulic chamber 74 are adjusted. The rotation of the rotor 44 can be advanced or retarded with respect to the rotation of 42. For example, when the rotor 44 is advanced to the maximum, as shown in FIG. 6, each vane 64 to 67 of the rotor 44 rotates in a state of contacting the shoe 71 on the retard hydraulic chamber 73 side. When the hydraulic pressure in the retard hydraulic chamber 73 is set higher than the hydraulic pressure in the advance hydraulic chamber 74, the rotor 44 rotates in the retard direction with respect to the housing 42. In this way, the hydraulic pressure supplied to the retard hydraulic chamber 73 and the advance hydraulic chamber 74 can be adjusted to adjust the retard / advance angle of the rotor 44 with respect to the housing 42.

Here, the supply hydraulic pressure of the OCV 80 is determined by a signal from a position sensor for detecting the relative rotation angle of the rotor 44 with respect to the housing 42 and a crank angle sensor for determining the amount of pressurization by the oil pump 92. And feedback control is performed.

FIGS. 9A to 9C are explanatory diagrams showing typical operating states of the OCV 80. FIG. FIG. 9A shows the ECU 1
This is an example when the control current value from 00 is 0.1 A, and the spool 82 is
1 shows a state where communication is performed between the supply port 85 and the A port 86 and between the B port 87 and the drain port 88b. In this state, the retard hydraulic chamber 73
Is supplied from the advance hydraulic chamber 74, while the hydraulic oil is discharged from the advance hydraulic chamber 74. Therefore, the rotor 44 shown in FIG. 9A rotates counterclockwise with respect to the housing 42 in FIG. The phase of the intake-side camshaft 19 with respect to the intake-side timing pulley 21 is delayed, and the retard control is performed.

FIG. 9B shows an example in which the control current from the ECU 100 is 0.5 A. When the opposing linear solenoid 84 and spring 83 are balanced, the spool 82
, The state in which both the A port 86 and the B port 87 are closed, and the supply and discharge of hydraulic oil from the retard hydraulic chamber 73 and the advance hydraulic chamber 74 are not performed. In this state, when there is no leakage of hydraulic oil from the retard hydraulic chamber 73 and the advance hydraulic chamber 74, the rotor 44 is maintained at the current position, and the phases of the intake-side timing pulley 21 and the intake-side camshaft 19 remain unchanged. Will be maintained.

FIG. 9C shows an example in which the control current from the ECU 100 is 1.0 A. The spool 82 is driven by the linear solenoid 84 to the right end of the housing 42.
Between the supply port 85 and the B port 87, and the A port 8
6 shows a state in which communication is established between the drain port 88a and the drain port 88a. In this state, the hydraulic oil is supplied to the advance hydraulic chamber 74 through the second oil passage 63, while the hydraulic oil is discharged from the retard hydraulic chamber 73. Accordingly, the rotor 44 shown in FIG. 9C rotates clockwise in FIG. 9 with respect to the housing 42, and the phase of the intake side camshaft 19 with respect to the intake side timing pulley 21 is advanced to perform the advance angle control.

9 (a), 9 (b) and 9 (c), the supply port 85 and the A port 86 (or the B port 8)
7) and the degree of communication between drain port 88b (or drain port 88a) and B port 87 (or A port 86)
Is controlled by the spool 82.
Further, the position of the spool 82 and the current value of the linear solenoid 84 are in a proportional relationship.

FIG. 10 is a characteristic diagram showing the relationship between the current value of the linear solenoid 84 (hereinafter referred to as the solenoid current) under a certain operating condition of the engine 1 and the actual valve timing change speed. In the figure, the positive region of the actual valve timing change speed corresponds to the region moving in the advance direction,
On the other hand, the negative region of the actual valve timing change speed corresponds to the region moving in the retard direction.

Symbols (a), (b) and (c) in FIG.
9 shows solenoid currents corresponding to the respective positions of the spool 82 in FIGS. 9A, 9B, and 9C, respectively. here,
The solenoid current whose actual valve timing does not change as shown by the symbol (b) is supplied to each of the hydraulic chambers 73 and 74 and the hydraulic piping 8.
9, 90, there is only one point where the amount of hydraulic oil leaking from each part of the spool 82 and the amount of hydraulic oil pumped from the oil pump 92 are balanced.

Further, this point always fluctuates because the characteristic changes as shown in FIG. 11 due to the hydraulic oil discharge pressure due to a change in the engine speed or temperature. In addition, spool 8
This point and the way of changing the characteristics differ for each product due to the product variation such as the size of 2. The solenoid current at a point where the actual valve timing does not change is hereinafter referred to as a holding current, and is represented by a symbol HLD. When it is desired to advance the valve timing with reference to the holding current HLD, the solenoid current can be increased, and when it is desired to retard the valve timing, the solenoid current can be decreased.

Next, a method of detecting the valve timing is shown in FIG.
2 will be described. FIG. 12A shows a crank angle signal, FIG. 12B shows a cam angle signal at the time of retarding, and FIG.
(C) is a timing chart showing the cam angle signal at the time of advance, respectively. The ECU 100 measures the phase difference time TVT between the crank angle signal period T and the cam angle signal, and calculates the phase difference time TV when the valve timing is in the most retarded state.
From T0 and the crank angle signal period T, the most retarded valve timing VTR is obtained by the equation 1 and stored in advance. (Equation 1) VTR = TVT0 / T × 180 ゜ CA The actual valve timing VTA is obtained from the phase difference time TVT, the crank angle signal cycle T, and the most retarded valve timing VTR by the equation (2). (Equation 2) VTR = TVT / T × 180 ゜ CA-VTR The ECU 100 feeds back the solenoid current based on the difference between the actual valve timing VTA and the target valve timing VTT, thereby changing the actual valve timing VTA to the target valve timing VTT. To converge.

Next, the internal configuration of the ECU 100 will be described. FIG. 13 is a block diagram showing the internal configuration of ECU 100. In the figure, reference numeral 101 denotes a microcomputer, which is a central processing unit (CP) for performing various calculations and determinations.
U) 102, a read-only memory (ROM) 103 for storing a predetermined control program in advance, a random access memory (RAM) 104 for temporarily storing the calculation results and the like by the CPU 102, and an A / A for converting an analog voltage into a digital value. A D converter 105, a counter 106 for measuring the period of the input signal, a timer 107 for measuring the drive time of the output signal, etc., and an output port 10 for outputting the output signal
8, 109 and an input port 110 for inputting an input signal
And a common bus 111 for connecting these.

Reference numeral 112 denotes a first input circuit, which shapes the waveforms of the input signals from the crank angle 6 and the cam angle sensor 24, and converts them into an interrupt command signal (INT).
Enter 01. The CPU 102 reads the value of the counter 106 every time the
To memorize. Here, the first input circuit 112 shapes the waveform of the signal from the crank angle sensor 6, and outputs the interrupt signal (IN
T) is input to the microcomputer 101, so that each time the interrupt is issued, the CPU 102
The value of the counter 106 is read and stored in the RAM 104, and the crank angle signal cycle T is calculated from the difference from the counter value when the signal was previously input from the crank angle sensor 6,
Further, an engine speed NE is calculated based on the crank angle signal period T. Further, the phase difference time TVT is calculated from the difference from the counter value when the signal is input from the cam angle sensor 24 stored in the RAM 104.

Reference numeral 113 denotes a second input circuit, which is a water temperature sensor 12, a throttle sensor 27, and an intake air amount sensor 28.
Is input to the A / D converter 105 after noise components are removed or amplified, and the cooling water temperature THW and the throttle opening T
It is converted into digital data of VO and intake air amount QA.

A driving circuit 114 drives the injector 30 and a driving circuit 115 drives the igniter 11. Here, the CPU 102 calculates the injector drive time and the ignition timing based on the various input signals described above, and based on the time measurement result of the timer 107,
The injector 30 and the igniter 11 are driven via the output port 108 and each of the drive circuits 112 and 113 to perform fuel injection amount control and ignition timing control.

Reference numeral 116 denotes a current control circuit for controlling the solenoid current of the OCV 80. Here, the CPU 10
2 calculates the solenoid current CNT of the OCV 80 based on the various input signals, and outputs a duty signal corresponding to the solenoid current CNT of the OCV 80 to the output port 108 based on the time measurement result of the timer 107. The current control circuit 116 generates an O based on the duty signal.
Valve timing control is performed by controlling the current flowing through the solenoid 84 of the CV 80 to be the solenoid current CNT.

Reference numeral 117 denotes a third input circuit for inputting a power signal to the input port 110; 118, a power supply circuit;
120 is a diode, 121 is a transistor, 122 is a battery (power supply), 123 is a key switch (IG switch), and 124 is a main relay. Here, when the IG switch 123 is turned on, the transistor 121 and the main relay 124 are sequentially turned on via the diode 119, and the power supply voltage from the battery 122 is changed to the power supply circuit 118.
The microcomputer 101 operates by being provided to the microcomputer 101 via the. During operation of the microcomputer 101,
When the CPU 102 sets the output port 109 to high output, the high output voltage keeps turning on the transistor 121 via the diode 120.
Even when the state becomes F, the microcomputer 120 maintains the operating state. Then, the CPU 102 sets the output port 109
Is low output (zero voltage), the transistor is off,
The main relay 124 is turned off, and the microcomputer 101 is turned off.

The battery 122 is connected to the various sensors 6, 24, 28, 27, 12 and the ECU 100 and the OC.
Although the power source is a common power source such as the V80, the crank angle sensor 6 and the OCV 80 are powered after the main relay 124. Therefore, the microcomputer 101
During the operation, the crank angle 6 and the OCV 80 also operate.

Next, the operation of the CPU 102 will be described with reference to FIG.
This will be described with reference to FIGS. FIG. 14 is an operation timing chart in the case where the actual holding current HLD matches the reference value 0.5 A in the control device without the integration control means,
FIG. 15 is an operation timing diagram when the actual holding current HLD is shifted to a higher current side than the reference value 0.5 A in the control device without the integration control means, and FIG. 16 is the control device with the integration control means. FIG. 9 is an operation timing chart when an actual holding current HLD is shifted to a higher current side than a reference value of 0.5 A.

The OCV 80 can adjust the amount of hydraulic oil supplied per unit time, and the displacement angle is determined according to the integrated amount of hydraulic oil supplied to the actuator 40. In this sense, since the actuator 40 to be controlled includes an integral element, if the actual holding current HLD of the OCV 80 is equal to the standard value 0.5 A, the control means sets a value of 0.
By performing proportional control according to the deviation ER between the target valve timing VTT and the actual valve timing VTA based on 5A, the actual valve timing can be made to converge on the target valve timing. At this time, the solenoid current CNT of the OCV 80 is given by Equation (3). (Equation 3) CNT = KP × ER + 0.5A In the equation (3), ER is the target valve timing VTT.
And the actual valve timing VTA deviation ER is calculated by the equation (4). (Equation 4) ER = VTT-VTA In the equation of Equation 3, KP is a gain corresponding to the proportional operation. FIG. 14 shows the movements of the target valve timing VTT, actual valve timing VTA, and solenoid current CNT at this time.

However, the actual holding current HL of the OCV 80
D does not always agree with the standard value 0.5A. For example, when the actual holding current HLD is shifted to a higher current side than the standard value 0.5 A, if the control based on the equation (3) is performed, as shown in FIG.
Does not converge to the target valve timing VTT and the steady-state error E
R1 remains. This steady-state error ER1 is expressed by the equation (5). (Equation 5) ER1 = (HLD-0.5A) / KP

Therefore, in the conventional control device,
The integral control is further added to the proportional control of the formula (1), and the control expressed by the formula (6) is performed so that the above-mentioned steady-state error does not remain. (Equation 6) CNT = KP × ER + ΣKI + 0.5A In the equation (6), ΣKI is the target valve timing VT.
An integrated correction value obtained by integrating the integral increase / decrease value calculated based on the deviation ER between T and the actual valve timing VTA.
Is calculated as in the following equation. (Equation 7) ΣKI = ΣKI (i−1) + KI × ER In the equation (7), ΣKI (i−1) is an integration correction value before the current integration, and KI is a gain corresponding to the integration operation. Also, KI × ER corresponds to the integral increase / decrease value, and this KI is the transient deviation E generated at the time of step response or the like.
Due to the increase of R, the integral correction value ΣKI fluctuates greatly,
As a result, it is set to a very small value so that the control does not become unstable. The target valve timing VTT and the actual valve timing VTT in a state where no steady-state deviation remains between the target valve timing VTT and the actual valve timing VTA, that is, in a state where the integral correction value ΣKI satisfies the equation 8 as a result of the integral control. FIG. 16 shows the movement of the valve timing VTA and the solenoid current CNT. (Equation 8) HLD@KI+0.5A

Next, the operation of the hydraulic valve timing system for an engine according to the first embodiment of the present invention will be described. FIG. 17 is an operation timing chart showing a control state in a state where the actual valve timing VTA converges near the target valve timing VTT. When the absolute value of the deviation ER between the target valve timing VTT and the actual valve timing VTA is less than a predetermined value E1 (for example, 1 CA), it is determined that the actual valve timing has converged to the target valve timing, and the integral increase / decrease is determined. The value is set to a value KI1 (for example, 0.1 mA) that is small enough to cope with a smile change in the holding current of the OCV 80. However, when the actual valve timing VTA is slightly moving toward the target valve timing VTT, it is not necessary to further increase or decrease the integral correction value ΣKI. 0.01 CA / 25
When moving toward the target valve timing at a speed of at least (ms) or more, the integration of the integral increase / decrease value is stopped.

FIG. 18 is an operation timing chart showing a control operation in a state where a steady deviation occurs between the actual valve timing VTA and the target valve timing VTT. Target valve timing VTT and actual valve timing VTA
Is greater than or equal to a predetermined value E1 and the actual valve timing VTA is moving at a speed equal to or higher than a predetermined integration stop determination speed toward the target valve timing VTT,
Target valve timing VTT and actual valve timing VT
In order to judge that a steady-state error has occurred between T and T, and to rapidly control in a direction to eliminate the steady-state error, the integral increase / decrease value is a value KI2 (for example, 1 mA) larger than the above-mentioned KI1.
Set to. Here, if the predetermined integration stop determination speed remains at V1, no matter how much the integration increase / decrease value is increased, the integration increase / decrease value stops when the change speed of the actual valve timing VTA becomes V1 or more. Therefore, the changing speed of the actual valve timing VTA does not become equal to or higher than V1. Therefore, when the absolute value of the deviation ER is equal to or greater than the predetermined value E1, the integration stop determination speed is set to a value V2 (for example, 0.1 CA / 25 ms) larger than V1 in accordance with the increase of the integration increase / decrease value. , The actual valve timing VTA is the speed V
At the speed of 2, the target valve timing can be approached.

FIG. 19 is an operation timing chart showing a control operation at the time of a step response in a state where the integral correction value is stable. As shown in this figure, the target valve timing VTT
May change, the actual valve timing VTA may not move for a while or may move slowly due to a delay in transmission of hydraulic oil. In this state, the target valve timing V
Since the absolute value of the deviation ER between TT and the actual valve timing VTA is equal to or greater than the predetermined value E1, if the above control is performed, it is erroneously determined that a steady-state deviation has occurred, and the integral increase / decrease value is set to KI2. Since the integration stop determination speed is set to VI2, the integration correction value ΣKI, which does not need to be increased or decreased, is increased or decreased. Therefore, the absolute value of the deviation ER is equal to the predetermined value E.
A predetermined time TD after the value has become equal to or more than the predetermined value EI from less than I
During (for example, 0.2 sec), the integral increase / decrease value is set to KI1 in the same manner as when the absolute value of the deviation ER is less than the predetermined value E1.
Then, the integration stop determination speed is set to V1.

Thus, the target valve timing VTT
Is changed, the increase / decrease of the integral correction value ΣKI in a state where the actual valve timing VTA has not yet moved can be suppressed slightly, and even if the actual valve timing VTA is slight, the movement toward the target valve timing VTT can be started. For example, since the calculation of the integral increase / decrease value is a low apex, the integral correction value ΣKI hardly changes. And a predetermined time TD
Has elapsed, the actual valve timing VTA is already V
(2) Moving toward the target valve timing VTT at the change speed of the abnormality, the integration of the integral increase / decrease value is stopped also at this time. Further, when the actual valve timing VTA approaches the target valve timing VTT, the change speed is changed. However, if the deviation ER from the target valve timing VTT falls within a region less than the predetermined value E1, the integration stop determination speed becomes V1.
In addition, since the integral increase / decrease value is a small value such as KI1, the integral correction value ΣKI does not unnecessarily increase / decrease at this time, and the actual valve timing VTA converges stably to the target valve timing VTT. .

The above operation will be described below with reference to the operation flowchart of FIG. FIG. 20 is an operation flowchart of the control program stored in the ROM 103. This flowchart is processed in the CPU 102 of the input 100 at every predetermined time, for example, every 25 ms. In FIG. 20, first, in step ST1, the crank angle sensor 6, the cam angle sensor 2
4. From each of the intake air amount sensor 28, the throttle sensor 27, and the water temperature sensor 12, the crank angle signal cycle T,
An engine operating state signal such as an engine speed NE, a phase difference time TVT, an intake air amount sensor QA, a throttle opening TVO, and a cooling water temperature THW is input.

Next, in step ST2, the crankshaft 5 is determined from the crank angle signal period T and the phase difference time TVT.
, The displacement angle (actual valve timing) VTA of the intake-side camshaft 19 with respect to is calculated by the equation (2). In step ST3, the engine speed NE, the intake air amount QA,
The target valve timing VTT is calculated based on each of the throttle opening TVO and the cooling water temperature THW.

In step ST4, the IG switch 123
Of the throttle valve 26 is determined to be ON, and if the IG switch 123 is ON, it is determined in step ST5 whether or not the throttle valve 26 is fully closed.
At 6, the deviation ER between the target valve timing VTT and the actual valve timing VTA is calculated according to the equation (4).

In step ST7, the integral correction value ΣKI is obtained according to the equation (7). In this case, ΣKI (i-1) in Equation 7 indicates ΣKI 25 ms before. The integral correction value ΣKI is initialized to 0 immediately after power is applied to the input 100. Step ST8
Then, the solenoid current CNT of the OCV 80 is obtained according to the equation (6).

In step ST9, when the result of determination in step ST5 is that the throttle valve 26 is fully closed, the OCV is held so that the actuator 40 is maintained at the most retarded position.
Energize (0.1 A → CNT) to the solenoid 84 of 80.

In step ST10, if it is determined in step ST4 that the IG switch 123 is OFF, the injector 30 and the igniter 11 are turned OFF, respectively.
In step ST11, it is determined whether or not the engine speed NE after the IG switch 123 is turned off is 500 rpm or more. If NE ≧ 500 rpm, the process proceeds to step ST9 where the solenoid 84 of the OCV 80 is set to (0.1A → CNT).
If NE ≧ 500 rpm is not satisfied, the solenoid current CNT of the OCV 80 is set to Σ in step ST12.
It is assumed that the holding current is KI + 0.4A. However, ΣKI + 0.
If 5A is shifted to the advanced side from the actual holding current, the actuator 40 shifts to the most retarded position. Therefore, with a slight shift, the solenoid current CNT of the OCV 80 is controlled so that the actuator 40 does not advance to the advanced side. To KI + 0.
5A-α (α: predetermined value, set to a value small enough to prevent oil leakage from the OCV when the engine is stopped).

Next, in step ST13, the timer 10
7, the OCV is output to the output port 108.
A duty signal corresponding to the solenoid current CNT of 80 is output. This duty signal is supplied to the current control circuit 11
6 and is controlled so that the current flowing through the solenoid 83 of the OCV 80 becomes the solenoid current CNT. As a result, control is performed so that the actual valve timing VTA becomes the target valve timing VTT.

In step ST14, the IG switch 12
It is determined whether or not a predetermined time (for example, 7 seconds) has elapsed after the OFF of the switch 3, and if the predetermined time has not elapsed, the main relay 124 is turned on in step ST15. In ST16, the main relay 124 is turned off.

According to the first embodiment described above, the holding current (0.1 V) is applied to OCV 80 for a predetermined time after the engine is stopped.
A → CNT), the OCV is maintained to maintain the actuator 40 at the most retarded position after the engine stops.
80 can be set to the holding position, thereby preventing oil leakage from the OCV 80. Therefore, even after the engine is stopped, the state where the hydraulic pressure is applied to the actuator 40 can be maintained, and the cam angle control at the time of starting the engine thereafter can be smoothly performed. This prevents the rotor from hunting and generating noise when the engine is started.

[0072]

As described above, according to the present invention, whether or not the power is cut off when the engine is stopped is detected by the power cut-off means, and the fluid supply means is operated for a predetermined time from the time of the power cut-off based on the detection result. The holding control means controls the fluid supply means to the holding position so that the hydraulic pressure of the actuator is held, and switches the fluid control means between normal control when the engine power is turned on and holding control when the engine power is turned off. Since the switching control is performed by means, even after the engine is stopped, oil leakage from the actuator system can be prevented, and the hydraulic pressure of the actuator does not suddenly decrease. Oil can be prevented from flowing out of the system due to inertia, so that cam angle control at the time of subsequent engine start is performed smoothly and engine start There is an effect, such as there is no such thing as emit a noise at the time of the rotor is causing hunting.

Further, according to the present invention, when the engine speed after the engine power supply is cut off is equal to or lower than the predetermined speed, the actuator is operated by operating the fluid supply means until the engine speed reaches the predetermined speed. In this case, the fluid supply means is controlled to the holding position in the operating state so that the actuator can be controlled to the most retarded position.

Further, according to the present invention, even when the engine speed after the engine power is cut off is equal to or higher than the predetermined speed, the fluid supply means is controlled to the operating position until the engine speed reaches the predetermined speed. Configuration,
There is an effect that the actuator after the engine is stopped can be controlled to the most retarded position.

Further, according to the present invention, the main relay of the engine power supply system is shut off after a lapse of a predetermined time after the engine power supply is shut off, so that oil leakage from the actuator system can be prevented even after the engine is stopped. hand,
The hydraulic pressure of the actuator does not suddenly drop, so that the cam angle control at the time of the subsequent engine start is performed smoothly, and the rotor does not hunt and generate noise at the time of engine start. The same effect can be obtained.

[Brief description of the drawings]

FIG. 1 is a block diagram showing a basic conceptual configuration of the present invention.

FIG. 2 is a schematic cross-sectional view showing a gasoline engine system including a hydraulic valve timing adjustment system according to Embodiment 1 of the present invention.

FIG. 3 is a sectional view showing a hydraulic valve timing adjustment system according to Embodiment 1 of the present invention;

FIG. 4 is a sectional view showing a state in which hydraulic pressure is applied to the plunger in FIG.

5 is a sectional view taken along the line XX of FIG. 3;

FIG. 6 is a partial cross-sectional view showing a moving state of the slide plate in FIG.

FIG. 7 is a sectional view taken along the line YY in FIG. 3;

8 is a sectional view taken along the line ZZ in FIG. 3;

FIG. 9 is an explanatory sectional view showing a typical operation state of the oil control valve.

FIG. 10 is a characteristic diagram showing a relationship between a linear solenoid current and an actual valve timing change speed.

FIG. 11 is a characteristic diagram showing a relationship between a linear solenoid current and an actual valve timing change speed.

FIG. 12 is an operation timing chart showing a phase relationship between a crank angle signal and a cam angle signal and a method for calculating an actual valve timing.

FIG. 13 is a block diagram showing an internal configuration of the electronic control unit.

FIG. 14 is an operation timing chart for explaining an operation in the case where the actual holding current HLD matches the standard value of 0.5 A in the control device having no integral control means.

FIG. 15 is an operation timing chart for explaining an operation in a case where an actual holding current HLD is shifted to a higher current side than a standard value of 0.5 A in a control device having no integration control means.

FIG. 16 is an operation timing chart for explaining an operation in a case where an actual holding current HLD is shifted to a higher current side than a standard value of 0.5 A in a device having integration control means.

FIG. 17 is an operation timing chart for explaining an operation in a state where the actual valve timing converges near the target valve timing in the embodiment of the present invention.

FIG. 18 is an operation timing chart for explaining an operation in a state where a steady deviation occurs between the actual valve timing and the target valve timing patent in the embodiment of the present invention.

FIG. 19 is an operation timing chart for explaining an operation at the time of a step response in a state where the integral correction value is stable in the embodiment of the present invention.

FIG. 20 is an operation flowchart showing the operation of the embodiment in the present invention.

[Explanation of symbols]

M1 engine, M2 combustion chamber, M3 intake passage, M4
Exhaust passage, M5 intake valve, M6 exhaust valve, M
7 Operating state detection means, M8 target valve timing calculation means, M9 actuator, M10 fluid supply means, M11 actual valve timing detection means, M12 actual valve timing control means, M13 holding control means, M
14 power cutoff detecting means, M15 switching means.

Claims (4)

[Claims] An intake valve and an exhaust valve which are driven at a predetermined timing in synchronization with rotation of an engine to open and close an intake passage and an exhaust passage leading to a combustion chamber, respectively, and for detecting an operation state of the engine. Operating state detecting means, target valve timing calculating means for calculating a target valve timing for the operating state of the engine based on the detection result of the operating state detecting means, and valve opening / closing timing of at least one of the intake valve and the exhaust valve Actuator for changing valve timing, fluid supply means for supplying hydraulic oil to the actuator to drive the actuator and adjusting the flow rate thereof, and the intake valve and the exhaust valve provided with the actuator At least one of the actual valve timings An actual valve timing detecting means for outputting the actual valve timing, and an actual valve timing control means for controlling the actuator by controlling the fluid supply means in order to change the actual valve timing to the target valve timing. In a hydraulic valve timing adjusting system, a power cutoff detecting means for detecting a power cutoff time when the engine is stopped, and the fluid supply means is operated for a fixed time from the power cutoff time based on a detection result of the power cutoff detecting means, so that the actuator is operated. Holding control means for controlling the fluid supply means to a holding position so that the hydraulic pressure is maintained, and switching for switching the fluid supply means between normal control when the engine power is turned on and hold control when the engine power is turned off. Means for adjusting the valve timing of a hydraulic valve. Beam. 2. The holding control means according to claim 1, wherein, when the engine speed after the engine power is cut off is equal to or lower than a predetermined speed, the fluid is controlled so that the actuator is held at the most retarded position until the engine speed reaches the predetermined speed. 2. The hydraulic valve timing adjustment system according to claim 1, wherein the supply means is operated. 3. The holding control unit according to claim 1, wherein the holding control unit controls the fluid supply unit to the holding position when the engine speed after the engine power is turned off is equal to or higher than a predetermined speed. Hydraulic valve timing adjustment system. 4. The hydraulic valve timing adjustment system according to claim 1, wherein the power cutoff detecting means cuts off a main relay of the engine power supply system after a predetermined time has elapsed after the engine power cutoff. .