JP6020307B2 - Multi-cylinder engine controller - Google Patents

Description

  The present invention relates to a control device for a multi-cylinder engine such as an automobile, and particularly belongs to the technical field of hydraulic control of a multi-cylinder engine.

  Conventionally, in a multi-cylinder engine for automobiles or the like, a hydraulically operated valve characteristic control device that controls opening and closing timings of intake valves and exhaust valves by oil pressure according to the operating state of the engine for improving fuel efficiency, and low engine load A hydraulically operated valve stop device that stops the opening / closing operation of the intake valve and the exhaust valve by hydraulic pressure is provided in an operating state or the like, and the valve characteristics are changed or the cylinder reduction operation is performed by operating these devices.

  For example, Patent Document 1 discloses a technique in which both a hydraulically operated valve characteristic control device and a valve stop device are provided, and the valve characteristics are changed or the reduced cylinder operation is performed by operating these devices.

  In the device described in Patent Document 1, the operating state (rotational speed, load) of the engine is used for the valve characteristic control device and the valve stop device in the reduced cylinder operation region where the valve stop device is operated. When it is determined that the operating pressure is less than or equal to the specified hydraulic pressure, the valve characteristic control device is prohibited from operating. The valve stop device is prevented from malfunctioning due to a lack of hydraulic pressure supplied to the device, thereby preventing the engine from becoming unstable.

Japanese Patent No. 4792478

  However, the fuel efficiency of the engine can be further improved by performing valve characteristic control according to the operating state of the engine in the cylinder that is operating even during the reduced cylinder operation. Since the operation of the characteristic control device is prohibited and there is no technical idea of performing the valve characteristic control during the reduced cylinder operation, the effect of improving the fuel consumption of the engine obtained by the conventional technique is also low.

  Therefore, the present invention controls the appropriate valve characteristics by operating the hydraulically operated valve characteristic control device while preventing the malfunction of the valve stopping device even during the reduced cylinder operation by the operation of the hydraulically operated valve stop device. To improve the fuel efficiency of the engine.

  In order to solve the above-described problems, a control device for a multi-cylinder engine according to the present invention is configured as follows.

First, the invention according to claim 1 of the present application is
An oil pump, a hydraulically operated valve characteristic control device that changes the characteristics of at least one of the intake valve and the exhaust valve by the oil pressure of the oil supplied from the oil pump, and of the intake valve and the exhaust valve by the oil pressure A control device for a multi-cylinder engine comprising a hydraulically operated valve stop device for reducing the cylinder by stopping at least one of the valves,
A first branch oil passage that branches off from a main oil passage through which oil is discharged from the oil pump and supplies oil to the hydraulically operated valve characteristic control device via a first direction switching valve;
A second branch oil passage that branches from the main oil passage and supplies oil to the hydraulically operated valve stop device via a second direction switching valve;
The hydraulically operated valve characteristic control device operates in the second branch oil passage between the branch point from the main oil passage to the second branch oil passage and the second direction switching valve during the reduced cylinder operation. During this period, a shielding valve that shields the second branch oil passage is provided.

The invention according to claim 2 of the present application is the invention according to claim 1,
The shielding valve is energized by a spring so as to open when the hydraulic pressure in the main oil passage exceeds the required hydraulic pressure of the hydraulically operated valve stop device, and allows oil flow only in one direction from the upstream side to the downstream side. It is a check valve to be regulated.

The invention according to claim 3 of the present application is the invention according to any one of claim 1 or claim 2,
The hydraulically operated valve characteristic control device includes an advance hydraulic chamber and a retard hydraulic chamber, and is configured to change the relative rotation phase of the camshaft with respect to the crankshaft by controlling the hydraulic pressure supplied to each hydraulic chamber. It is characterized by.

  With the above configuration, according to the invention according to each claim of the present application, the following effects can be obtained.

According to the first aspect of the present invention, while the hydraulically operated valve characteristic control device operates during the reduced cylinder operation, it is between the branch point from the main oil passage to the second branch oil passage and the second direction switching valve. In this embodiment, since the oil flow between the main oil passage and the hydraulically operated valve stop device is shielded by the shielding valve provided in the second branch oil passage, the main oil passage is activated by the operation of the hydraulically operated valve characteristic control device. It is possible to prevent the hydraulic pressure of the engine from temporarily decreasing. As a result, it is possible to prevent the oil pressure supplied to the hydraulically operated valve stop device from being lowered, causing the valve stop device to malfunction, and preventing the reduced cylinder operation that keeps the intake valve and the exhaust valve in the stopped state. Therefore, it is possible to further improve the fuel efficiency of the engine by changing the valve characteristics during the reduced cylinder operation.

According to the second aspect of the present invention, when the oil pressure in the main oil passage is equal to or higher than the required oil pressure of the valve stop device, the check valve is opened, so that the oil pressure in the second branch oil passage is It becomes the same and can supply the oil pressure more than the required oil pressure to the valve stop device. On the other hand, when the oil pressure in the main oil passage is less than the required oil pressure of the valve stop device, the check valve closes, so the oil pressure in the second branch oil passage is not affected by the oil pressure in the main oil passage,
The required oil pressure of the valve stop device is maintained. Therefore, the malfunction of the hydraulically operated valve stop device can be prevented only by adding a simple configuration in which a spring-biased check valve is provided in the second branch oil passage without performing special control.

  According to the third aspect of the invention, the hydraulically operated valve characteristic control device is a variable valve timing mechanism having an advance hydraulic chamber and a retard hydraulic chamber, and the hydraulic operation on the intake side and the exhaust side during the reduced cylinder operation. Even when the valve-type valve characteristic control device is operated at the same time, it is possible to reliably prevent the hydraulically operated valve stop device from malfunctioning due to insufficient hydraulic pressure, as in the first aspect of the invention.

It is a figure showing a schematic structure of an engine which is one embodiment of the present invention. It is sectional drawing which shows schematic structure of a hydraulically operated valve stop apparatus. It is a side view showing a schematic structure of a hydraulically operated valve characteristic control device. It is a figure which shows schematic structure of an oil supply apparatus. It is a figure which shows the characteristic of a variable displacement type oil pump. It is a conceptual diagram which shows the reduced cylinder operating area | region of an engine. It is a figure explaining the setting of the target oil pressure of a pump. It is a hydraulic control map which shows the target oil pressure with respect to the operating state of an engine. It is a duty ratio map which shows the duty ratio with respect to the driving | running state of an engine. It is a flowchart which shows the flow control method of a pump. It is a flowchart which shows the cylinder number control method of an engine. It is a time chart which shows the driving | running state of the engine at the time of switching to a reduced cylinder driving | operation. It is an enlarged view which shows the downstream structure of the oil supply apparatus of FIG.

  Hereinafter, an embodiment of an oil supply apparatus 1 for an engine according to the present invention will be described with reference to FIGS. 1 to 13.

  First, an engine 2 to which an oil supply apparatus 1 according to an embodiment of the present invention is applied will be described with reference to FIG. As shown in the figure, the engine 2 is an in-line four-cylinder gasoline engine in which the first cylinder to the fourth cylinder are arranged in series, and includes a cam cap 3, a cylinder head 4, a cylinder block 5, and a crankcase (not shown). And an oil pan 6 (see FIG. 4) are vertically connected to each other, a piston 8 slidable in each of four cylinder bores 7 formed in the cylinder block 5, and a crankshaft 9 rotatably supported by a crankcase, Are connected by a connecting rod 10, and a combustion chamber 11 is formed for each cylinder by a cylinder bore 7 and a piston 8 in the upper part of the cylinder block 5.

  The cylinder head 4 is provided with an intake port 12 and an exhaust port 13 that open to the combustion chamber 11, and an intake valve 14 and an exhaust valve 15 that open and close the intake port 12 and the exhaust port 13 are provided in the ports 12 and 13. Yes. These intake and exhaust valves 14 and 15 are urged in the closing direction (upward in FIG. 1) by return springs 16 and 17, and swing arms are provided by cam portions 18a and 19a provided on the outer circumferences of the rotating camshafts 18 and 19, respectively. The cam followers 20a and 21a, which are rotatably provided at substantially central portions of the swing arms 20 and 21, are pushed downward, and the swing arm 20 is supported with the top of the pivot mechanism 25a provided on one end side of the swing arms 20 and 21 as a fulcrum. , 21 is configured such that the intake / exhaust valves 14 and 15 are pushed downward at the other end of the swing arms 20 and 21 against the urging force of the return springs 16 and 17 to open. Yes.

  As the pivot mechanism 25a of the swing arms 20 and 21 of the second and third cylinders in the center of the engine, a known hydraulic lash adjuster 24 that automatically adjusts the valve clearance to zero by hydraulic pressure (hereinafter abbreviated as "Hydraulic Lash Adjuster" is used. (Referred to as “HLA”).

  Further, HLA 25 (see FIG. 1) with a valve stop mechanism, which will be described later, provided with a pivot mechanism 25a is provided for the swing arms 20 and 21 of the first and fourth cylinders at both ends of the engine. The HLA 25 with a valve stop mechanism is configured so that the valve clearance can be automatically adjusted to zero similarly to the HLA 24, but the intake and exhaust valves 14 and 15 of the first and fourth cylinders according to the operating state of the engine 2. It is also possible to switch between a reduced cylinder operation for stopping the opening / closing operation of the cylinder and an all cylinder operation for opening / closing the intake / exhaust valves 14 and 15 of all the cylinders.

  The cylinder head 4 is provided with mounting holes 26 and 27 for inserting and mounting the lower ends of the HLA 24 and the HLA 25 with a valve stop mechanism. The cylinder head 4 has two oil passages 61, 62, 63, 64 communicating with the mounting holes 26, 27 for the HLA 25 with a valve stop mechanism, and the HLA 25 with a valve stop mechanism is mounted in the mounting hole 26. 27, the oil passages 61 and 62 supply hydraulic pressure (operating pressure) for operating the valve stop mechanism 25b of the HLA 25 with a valve stop mechanism, and the oil passages 63 and 64 are HLA 25 with a valve stop mechanism. The pivot mechanism 25a is configured to automatically supply hydraulic pressure for adjusting the valve clearance to zero.

  The cylinder block 5 is provided with a main gallery 54 that extends in the cylinder row direction in the side wall on the exhaust side of the cylinder bore 7. An oil jet 28 for cooling the piston communicating with the main gallery 54 is provided for each piston 8 in the vicinity of the lower side of the main gallery 54. The oil jet 28 has a nozzle portion 28 a disposed on the lower side of the piston 8, and engine oil (hereinafter simply referred to as “oil”) from the nozzle portion 28 a toward the back surface of the top portion of the piston 8. Is configured to inject fuel.

  Above each camshaft 18, 19, oil showers 29, 30 formed of pipes are provided, and the cam portions of the camshafts 18, 19 below which lubricate oil from the oil showers 29, 30. It is comprised so that it may be dripped at the contact part with the cam followers 20a and 21a of 18a and 19a, and the swing arms 20 and 21 further down.

  Next, the valve stop mechanism 25b which is one of the hydraulic actuators will be described with reference to FIG. The valve stop mechanism 25b has a reduced-cylinder operation that stops the opening and closing operations of the intake and exhaust valves 14 and 15 of the first and fourth cylinders according to the operating state of the engine 2, and all the HLAs 24 are normally operated to This is a mechanism for switching to full cylinder operation that causes the intake and exhaust valves 14 and 15 to open and close.

  As shown in the drawing, the HLA 25 with a valve stop mechanism includes a pivot mechanism 25a and a valve stop mechanism 25b. The pivot mechanism 25a is configured by an HLA that automatically adjusts the valve clearance to zero by hydraulic pressure, and since it has substantially the same configuration as the known HLA 24 used for the second and third cylinders, a description thereof is omitted. . The valve stop mechanism 25b can enter and exit a bottomed outer cylinder 251 that accommodates the pivot mechanism 25a so as to be slidable in the axial direction, and two through holes 251a that are provided to face the side peripheral surface of the outer cylinder 251. A pair of lock pins 252 that can be switched to a locked state or an unlocked state, and a lock spring 253 that biases these lock pins 252 radially outward. The lost motion spring 254 is provided between the inner bottom portion of the outer cylinder 251 and the bottom portion of the pivot mechanism 25a and presses the pivot mechanism 25a upwardly of the outer cylinder 251 to urge it.

  As shown in FIG. 2A, when the lock pin 252 is fitted in the through hole 251a of the outer cylinder 251 and the pivot mechanism 25a protrudes upward and is fixed, as shown in FIG. Since the top portion of the pivot mechanism 25a in the locked state serves as a fulcrum for swinging the swing arms 20, 21, when the cam portions 18a, 19a push the cam followers 20a, 21a downward by the rotation of the cam shafts 18, 19, the intake / exhaust valves 14 and 15 are pushed downward against the urging force of the return springs 16 and 17 to open the valves. Therefore, all cylinder operation can be performed by setting the valve stop mechanism 25b to the locked state for the first and fourth cylinders.

  As shown in FIG. 2 (b), when the outer end surfaces of both lock pins 252 are pressed by the operating oil pressure, the inner diameter of the outer cylinder 251 approaches the lock pins 252 against each other against the tensile force of the lock spring 253. Retreats in the direction and does not engage with the through hole 251a of the outer cylinder 251, and the pivot mechanism 25a located above is in an unlocked state in which it can move in the axial direction.

  In this unlocked state, when the pivot mechanism 25a is pressed downward against the urging force of the lost motion spring 254, the valve is stopped as shown in FIG. That is, the return springs 16 and 17 that urge the intake and exhaust valves 14 and 15 upward are configured to have a stronger urging force than the lost motion spring 254 that urges the pivot mechanism 25a upward. When the cam portions 18a and 19a push the cam followers 20a and 21a downward by the rotation of the camshafts 18 and 19, the top portions of the intake and exhaust valves 14 and 15 become fulcrums of the swing arms 20 and 21, and the intake and exhaust valves 14 15 and 15 are closed, the pivot mechanism 25a is pushed downward against the urging force of the lost motion spring 254. Therefore, a reduced cylinder operation can be performed by setting the valve stop mechanism 25b to the unlocked state.

  Next, variable valve timing mechanisms 32 and 33 (hereinafter simply referred to as “VVT”), which is one of the hydraulic actuators, will be described with reference to FIG. Referring to FIG. 3A, the VVTs 32 and 33 have a substantially annular housing 331 and a rotor 332 accommodated in the housing 331, and the housing 331 rotates in synchronization with the crankshaft 9. The cam pulley 333 and the rotor 332 are coupled to the camshafts 18 and 19 for opening and closing the intake and exhaust valves 14 and 15 so as to be integrally rotatable. A plurality of retarded hydraulic chambers 335 and advanced hydraulic chambers 336 defined by an inner peripheral surface of the housing 331 and a vane 334 provided on the rotor 332 are formed inside the housing 331. The hydraulic chambers 335 and 336 are connected to a pump 36 that supplies oil via first direction switching valves 34 and 35. When the oil is guided to the retarded hydraulic chamber 335 by the control of the first direction switching valves 34 and 35, the camshafts 18 and 19 are moved in the direction opposite to the rotation direction by the hydraulic pressure. On the other hand, when oil is guided to the advance hydraulic chamber 336, the camshafts 18 and 19 move in the rotational direction due to the hydraulic pressure, so that the opening timing of the intake and exhaust valves 14 and 15 is advanced.

  FIG. 3B shows the valve opening phase of the intake valve 14 and the exhaust valve 15. As can be seen from the figure, the valve opening phase of the intake valve 14 is advanced in the advance direction (see arrows) by the VVTs 32 and 33. If changed, the valve opening period of the exhaust valve 15 and the valve opening period of the intake valve 14 (see the alternate long and short dash line) overlap. By overlapping the opening periods of the intake valve 14 and the exhaust valve 15 in this way, the amount of internal EGR during engine combustion can be increased, and pumping loss can be reduced to improve fuel efficiency. Further, since the combustion temperature can be suppressed, NOx generation can be suppressed and exhaust purification can be achieved. On the other hand, if the valve opening phase of the intake valve 14 is changed to the retarded direction by the VVTs 32 and 33, the exhaust valve 15, the valve opening period, and the valve opening period of the intake valve 14 (see the solid line) do not overlap. Stable combustion can be secured and engine output can be improved during high-speed operation.

  Next, the oil supply apparatus 1 according to the embodiment of the present invention will be described in detail with reference to FIG. As shown in the figure, an oil supply device 1 of the present embodiment is a device for supplying oil to the engine 2 described above, and is a variable displacement oil pump 36 (hereinafter referred to as “drive”) driven by rotation of a crankshaft 9 (not shown). And an oil supply passage 50 which is connected to the pump 36 and guides the pressurized oil to each part of the engine.

  The oil supply passage 50 is composed of a pipe, a passage formed in the cylinder block 5, the cylinder head 4, and the like. The oil supply passage 50 communicates with the pump 36, a first communication path 51 extending from the oil pan 6 to the branch point 54 a in the cylinder block 5, a main gallery 54 extending in the cylinder row direction in the cylinder block 5, and the main gallery 54. A second communication path 52 extending from the upper branching point 54 b to the cylinder head 4, a third communication path 53 extending in the horizontal direction between the intake side and the exhaust side in the cylinder head 4, and a second communication path in the cylinder head 4. A plurality of oil passages 61 to 68 branched from the three-way passage 53 are provided.

  The pump 36 of the present embodiment is a known variable displacement oil pump, which is formed so that one end side is open, and has a pump body having a U-shaped cross section having a pump housing chamber formed of a cylindrical space inside. A housing 361 comprising a cover member that once closes the opening of the pump body, and a drive shaft 362 that is rotatably supported by the housing 361 and is driven to rotate by the crankshaft 9 through substantially the center of the pump housing chamber. And a rotor 363 rotatably accommodated in the pump housing chamber and having a central portion coupled to the drive shaft, and a plurality of slits formed in the outer peripheral portion of the rotor 363 in a radially cutout manner. A pump element composed of a rotary element 364 and disposed on the outer peripheral side of the pump element so as to be eccentric with respect to the rotation center of the rotor 363. A cam ring 366 that defines a pump chamber 365 that is a plurality of hydraulic oil chambers with adjacent vanes 364, and a cam ring 366 that is accommodated in the pump body and increases in the amount of eccentricity of the cam ring 366 with respect to the rotation center of the rotor 363. And a pair of ring members 368 having a diameter smaller than that of the rotor 363 slidably disposed on both side portions on the inner peripheral side of the rotor 363. The housing 361 includes a suction port 361 a that supplies oil to the internal pump chamber 365 and a discharge port 361 b that discharges oil from the pump chamber 365. A pressure chamber 369 defined by the inner peripheral surface of the housing 361 and the outer peripheral surface of the cam ring 366 is formed in the housing 361, and an introduction hole 369 a that opens to the pressure chamber 369 is provided. The pump 36 introduces oil into the pressure chamber 369 from the introduction hole 369 a, so that the cam ring 366 swings with respect to the fulcrum 361 c, and the rotor 363 is eccentric relative to the cam ring 366. It is configured to increase capacity.

  An oil strainer 39 facing the oil pan 6 is connected to the suction port 361 a of the pump 36. An oil filter 37 and an oil cooler 38 are disposed in order from the upstream side to the downstream side in the first communication path 51 communicating with the discharge port 361b of the pump 36, and the oil stored in the oil pan 6 is stored in the oil strainer. The pump is pumped up by a pump 36 through 39, filtered by an oil filter 37, cooled by an oil cooler 38, and then introduced into a main gallery 54 in the cylinder block 5.

  The main gallery 54 includes an oil jet 28 for injecting cooling oil to the back side of the four pistons 8, and a metal bearing oil supply unit disposed on five main journals that rotatably support the crankshaft 9. 41 and an oil supply part 42 of a metal bearing disposed on a crank pin of the crankshaft 9 that rotatably connects the four connecting rods, and oil is constantly supplied to the main gallery 54.

  In the downstream of the branch point 54 c on the main gallery 54, the oil is supplied from an oil supply portion 43 that supplies oil to the hydraulic chain tensioner and an introduction hole 369 a to the pressure chamber 369 of the pump 36 via the linear solenoid valve 49 in order. Is connected to the oil passage 40.

  The oil passage 68 branched from the branch point 53a of the third communication passage 53 is delayed from the advance hydraulic chamber 336 of the exhaust side VVT 33 for changing the opening / closing timing of the exhaust valve 15 via the exhaust side first direction switching valve 35. It is connected to the angular hydraulic chamber 335, and is configured such that oil is supplied by operating the first direction switching valve 35. The oil passage 66 that branches from the branch point 64 a of the oil passage 64 is connected to an oil shower 30 that supplies lubricating oil to the exhaust-side swing arm 21, and oil is constantly supplied to the oil passage 66. The oil passage 64 includes a metal bearing oil supply unit 45 (see a white triangle △ in FIG. 4) disposed in the cam journal of the camshaft 19 on the exhaust side, an HLA 24 (see a black triangle ▲ in FIG. 4), It is connected to an HLA 25 with a valve stop mechanism (see a white ellipse in FIG. 4), and oil is always supplied to the oil passage 64.

  Similarly, on the intake side, the oil passage 67 branched from the branch point 53c of the third communication passage 53 is advanced by the VVT 32 for changing the opening / closing timing of the intake valve 14 via the intake side first direction switching valve 34. The angular hydraulic chamber 326 and the retard hydraulic chamber 325 are connected. An oil passage 65 that branches from a branch point 63a of the oil passage 63 is connected to an oil shower 29 that supplies lubricating oil to the swing arm 20 on the intake side. The oil passage 63 branched from the branch point 53d of the third communication passage 53 includes a metal bearing oil supply portion 44 (see a white triangle Δ in FIG. 4) arranged in the cam journal of the camshaft 18 on the intake side, and the HLA 24. (Refer to the black triangle ▲ in FIG. 4) and the HLA 25 with a valve stop mechanism (see the white ellipse in FIG. 4).

  The oil passage 69 branched from the branch point 53c of the third communication passage 53 branches at the branch point 69a via a check valve 48 that restricts the direction of oil flow in only one direction from the upstream side to the downstream side. Are connected to the valve stop mechanism 25b of the HLA 25 with the valve stop mechanism on the exhaust side and the intake side via the second direction switch valves 46, 47 on the exhaust side and the intake side, respectively. Is configured so that oil is supplied to each valve stop mechanism 25b. Further, a hydraulic pressure sensor 70 for detecting the hydraulic pressure is connected between the check valve 48 on the oil passage 69 and the branch point 53c.

  Lubricating and cooling oil supplied to the metal bearing, the oil jet 28, the oil shower 29, 30 and the like that rotatably support the crankshaft 9 and the camshafts 18 and 19 are not shown after cooling and lubrication. It drops and circulates in the oil pan 6 through the drain oil passage.

  The operating state of the engine is detected by various sensors. For example, the rotation angle of the crankshaft 9 is detected by the crank position sensor 71, and the engine rotation speed is calculated based on the detection signal. The throttle position sensor 72 detects the opening of the throttle valve, and the engine load is calculated based on this. The oil temperature sensor 73 and the hydraulic pressure sensor 70 detect the temperature and pressure of the engine oil, respectively. The rotational angle of the camshafts 18 and 19 is detected by a cam angle sensor 74 provided in the vicinity of the camshafts 18 and 19, and the operating angles of the VVTs 32 and 33 are detected based on the cam angles. Further, the water temperature sensor 75 detects the water temperature of the cooling water that cools the engine 2.

  The controller 100 includes a microcomputer and the like, and includes a signal input unit for inputting detection signals from various sensors (crank position sensor 71, throttle position sensor 72, oil temperature sensor 73, oil pressure sensor 70, etc.), and arithmetic processing related to control. Stores a calculation unit to be performed, a signal output unit that outputs a control signal to a device to be controlled (such as a linear solenoid valve 49 described later), and a program and data (such as a hydraulic control map and a duty ratio map described later) necessary for control. A storage unit is provided.

  The linear solenoid valve 49 is a pump control device for controlling the discharge amount from the pump 36 in accordance with the operating state of the engine. Although oil is supplied to the pressure chamber 369 of the pump 36 when the linear solenoid valve 49 is opened, the description of the configuration of the linear solenoid valve 49 itself is omitted because it is well known.

  The linear solenoid valve 49 controls the hydraulic pressure supplied to the pressure chamber 369 of the pump 36 in accordance with the duty ratio control signal sent from the controller 100 based on the operating state of the engine 2. The discharge amount (flow rate) of the pump 36 is controlled by controlling the amount of eccentricity of the cam ring 366 and the amount of change in the internal volume of the pump chamber 365 by the hydraulic pressure of the pressure chamber 369. That is, the capacity of the pump 36 is controlled by the duty ratio. Here, since the pump 36 is driven by the crankshaft 9 of the engine 2, as shown in FIG. 5, the flow rate (discharge amount) of the pump 36 is proportional to the engine rotation speed. When the duty ratio represents the ratio of the energization time to the linear solenoid valve with respect to the time of one cycle, as shown in the figure, the hydraulic pressure to the pressure chamber 369 of the pump 36 increases as the duty ratio increases. The slope of the flow rate of the pump 36 is reduced.

  Next, the reduced cylinder operation of the engine will be described with reference to FIG. The reduced-cylinder operation or all-cylinder operation of the engine is switched according to the operating state of the engine. That is, the reduced cylinder operation is executed when the engine operating state ascertained from the engine rotation speed, the engine load, and the coolant temperature of the engine is within the reduced cylinder operation region shown in the figure. In addition, as shown in the figure, a reduced cylinder operation preparation area is provided adjacent to the reduced cylinder operation area, and when the engine is in the reduced cylinder operation preparation area, the reduced cylinder operation is executed. As a preparation for this, the hydraulic pressure is increased in advance toward the required hydraulic pressure of the valve stop mechanism. When the engine operating state is outside these reduced-cylinder operation region and reduced-cylinder operation preparation region, all-cylinder operation is executed.

  Referring to FIG. 6A, for example, when the engine speed is increased by accelerating at a predetermined engine load, all cylinder operation is performed if the engine speed is less than V1, and the engine speed is V1 or more, V2 When the engine speed is less than V2, preparation for the reduced cylinder operation is started, and when the engine speed becomes V2 or more, the reduced cylinder operation is performed. Further, for example, when the engine speed is decreased by decelerating with a predetermined engine load, the entire cylinder operation is performed when the engine speed is V4 or more, and the reduced cylinder operation is performed when the engine speed is V3 or more and less than V4. When the engine speed becomes V3 or less, the reduced cylinder operation is performed.

  Referring to FIG. 6B, for example, when the vehicle runs at a predetermined engine speed and a predetermined engine load and the engine warms up and the temperature of the cooling water rises, all cylinder operation is performed when the water temperature is lower than T0. When the water temperature is equal to or higher than T0 and lower than T1, preparation for reduced cylinder operation is performed, and when the water temperature is equal to or higher than T1, reduced cylinder operation is performed.

  If this reduced-cylinder operation preparation area is not provided, when switching from all-cylinder operation to reduced-cylinder operation, the hydraulic pressure is increased to the required hydraulic pressure of the valve stop mechanism after the engine operating state enters the reduced-cylinder operation area. However, since the time for performing the reduced cylinder operation is shortened by the time until the hydraulic pressure reaches the required oil pressure, the fuel efficiency of the engine is reduced by the amount of time for performing the reduced cylinder operation.

  Therefore, in this embodiment, in order to maximize the engine fuel efficiency, a reduced cylinder operation preparation region is provided adjacent to the reduced cylinder operation region, and the hydraulic pressure is increased in advance in the reduced cylinder operation preparation region. A target hydraulic pressure map (see FIG. 7A) is set so as to eliminate a loss for the time required to reach the required hydraulic pressure.

  In addition, as shown to Fig.6 (a), it is good also considering the area | region shown with the dashed-dotted line adjacent to the high engine load side of a reduced cylinder operation area | region as a reduced cylinder operation preparation area | region. Thereby, for example, when the engine load decreases at a predetermined engine speed, all cylinder operation is performed when the engine load is L1 (> L0) or more, and when the engine load becomes L0 or more and less than L1, the reduced cylinder operation is performed. When the engine load becomes L0 or less, the reduced cylinder operation may be performed.

  Next, the required hydraulic pressure of each hydraulic actuator and the target hydraulic pressure of the pump 36 will be described with reference to FIG. The oil supply device 1 in this embodiment supplies oil to a plurality of hydraulic actuators by a single pump 36, and the required hydraulic pressure required by each hydraulic actuator changes depending on the operating state of the engine. Therefore, in order to obtain a hydraulic pressure that is required for all hydraulic operating devices in all engine operating states, the pump 36 has a hydraulic pressure higher than the highest required hydraulic pressure among the required hydraulic pressures of each hydraulic operating device for each engine operating state. It is necessary to set the hydraulic pressure at the target hydraulic pressure according to the operating state of the engine. For this purpose, in the present embodiment, the required hydraulic pressures of the valve stop mechanism 25b, the oil jet 28, the journal of the crankshaft, and the VVTs 32, 33, which are relatively high among all the hydraulic actuators, are satisfied. The target hydraulic pressure may be set as follows. This is because, if the target oil pressure is set in this way, other hydraulic actuators having a relatively low required oil pressure naturally satisfy the required oil pressure.

  Referring to FIG. 7A, when the engine is operating at a low load, the hydraulic actuators having a relatively high required oil pressure are VVTs 32 and 33, metal bearings, and a valve stop mechanism 25b. The required oil pressure of each of these hydraulic actuators changes according to the operating state of the engine. For example, the VVT required oil pressure is substantially constant above a predetermined engine speed (V0). The required oil pressure of the metal bearing (shown as “metal required oil pressure”) increases as the engine speed increases. The valve stop required hydraulic pressure is substantially constant at a predetermined range of engine speed (V2 to V3). When these required oil pressures are compared for each engine rotation speed, there is only a metal required oil pressure when the engine rotation speed is V0 or less, and when the engine rotation speed is between V0 and V2, the VVT required oil pressure is the highest and the engine rotation speed is From V2 to V3, the valve stop required oil pressure is the highest, when the engine speed is V3 to V6, the VVT required oil pressure is the highest, and when the engine speed is V6 or higher, the metal required oil pressure is the highest. Therefore, it is necessary to set the above-described highest required oil pressure as the reference target oil pressure as the target oil pressure of the pump 36 for each engine speed.

  Here, at the engine rotational speeds (V1 to V2, V3 to V4) before and after the engine rotational speed (V2 to V3) at which the reduced cylinder operation is performed, the target hydraulic pressure is the valve stop request hydraulic pressure in preparation for the reduced cylinder operation. Is corrected and set from the reference target hydraulic pressure so as to increase the pressure in advance. According to this, as described with reference to FIG. 6, when the engine rotational speed reaches the engine rotational speed at which the reduced cylinder operation is performed, the loss of time until the hydraulic pressure reaches the valve stop required hydraulic pressure is eliminated, and the fuel consumption of the engine is reduced. Efficiency can be improved. An example of the oil pump target oil pressure set by this correction is indicated by thick lines (V1 to V2, V3 to V4) in FIG.

  Further, in consideration of a delay in response of the pump 36, an overload of the pump 36, etc., an engine in which the required oil pressure changes abruptly with respect to the engine speed with respect to the reference target oil pressure after the above-described reduction cylinder reduction preparation is corrected. The target oil pressure is set by correcting so that the change in the oil pressure at the rotation speed (for example, V0, V1, V4) becomes small so that the oil pressure is higher than the required oil pressure and gradually increases or decreases according to the engine rotation speed. Good. An example of the oil pump target oil pressure set by performing this correction is shown by a thick line (V0 or less, V0 to V1, V4 to V5) in FIG.

  Referring to FIG. 7B, when the engine is operating at a high load, the hydraulic actuators having a relatively high required oil pressure are the VVTs 32 and 33, the metal bearing, and the oil jet 28. As in the case of low load operation, the required hydraulic pressure of each of these hydraulic actuators changes according to the operating state of the engine. For example, the VVT required hydraulic pressure is substantially constant at a predetermined engine speed (V0 ′) or higher. The metal required hydraulic pressure increases as the engine speed increases. The oil jet 28 increases in accordance with the engine rotational speed up to a predetermined engine rotational speed, and is constant above the predetermined engine rotational speed.

  In the case of high load operation, as in the case of low load operation, the target target oil pressure is corrected by correcting the reference target oil pressure at the engine rotation speed (for example, V0 ′, V2 ′) at which the required oil pressure changes rapidly with respect to the engine rotation speed. An example of the oil pump target oil pressure set by appropriately correcting is shown in FIG. 7B by a thick line (in particular, V0 ′ or less, V1 ′ to V2 ′).

  Note that the oil pump target hydraulic pressure shown in the figure changes in a polygonal line shape, but may change smoothly in a curved line shape. In the present embodiment, the target hydraulic pressure is set based on the required hydraulic pressures of the valve stop mechanism 25b, the oil jet 28, the metal bearing, and the VVTs 32, 33 having a relatively high required hydraulic pressure. However, the hydraulic actuator is not limited to these. What is necessary is just to set the target hydraulic pressure in consideration of the required hydraulic pressure, whatever the hydraulic actuator having a relatively high required hydraulic pressure.

  Next, the hydraulic control map will be described with reference to FIG. The oil pump target oil pressure shown in FIG. 7 uses the engine rotational speed as a parameter. Further, the oil pump target oil pressure is shown in a three-dimensional graph using the engine load and oil temperature as parameters, as shown in FIG. This is a hydraulic control map. That is, this hydraulic pressure control map corresponds to the operating state of the engine based on the highest required hydraulic pressure among the required hydraulic pressures of each hydraulic actuator for each engine operating state (engine speed, engine load and oil temperature). The target oil pressure is set in advance.

  FIGS. 8A, 8B and 8C show hydraulic control maps when the engine (oil temperature) is hot, warm and cold, respectively. The controller 100 uses these hydraulic control maps properly according to the oil temperature. That is, when the engine is started and the engine is in a cold state (oil temperature is lower than T1), the controller 100 determines whether the engine is in an operational state (see FIG. 8C) based on the cold hydraulic control map. Read the target oil pressure according to the engine speed and engine load. When the engine warms up and the oil reaches a predetermined oil temperature T1 or more, the target oil pressure is read based on the warm oil pressure control map shown in FIG. When the oil temperature becomes equal to or higher than the oil temperature T2 (> T1), the target oil pressure is read based on the oil pressure control map at the time of high temperature shown in FIG.

  In this embodiment, the target oil pressure is read using a hydraulic control map set in advance for each temperature range by dividing the oil temperature into three temperature ranges of high temperature, warm time, and cold time. More hydraulic control maps may be prepared by dividing the temperature range more finely. In addition, the oil pressure t within the temperature range (T1 ≦ t <T2) targeted by one oil pressure control map (for example, the oil pressure control map at the time of warming) has read the target oil pressure P1 having the same value. In consideration of the target oil pressure (P2) within the temperature range before and after (T2 ≦ t), the target oil pressure p is proportionally converted according to the oil temperature t (p = (t−T1) × (P2−P1) / ( It may be calculated according to T2-T1)). As described above, the target hydraulic pressure corresponding to the temperature can be read and calculated with higher accuracy, so that the pump displacement can be controlled with higher accuracy.

  Next, the duty ratio map will be described with reference to FIG. The duty ratio map here refers to the target oil pressure for each engine operating state (engine speed, engine load, oil temperature) read from the aforementioned oil pressure control map, and the flow path resistance of the oil path based on the read target oil pressure. The target operating amount of oil supplied from the pump 36 is set in consideration of the above, and the operating state of the engine calculated in consideration of the engine rotation speed (oil pump speed) and the like based on the set target discharging amount The target duty ratio corresponding to is set in advance.

  FIGS. 9A, 9B, and 9C show duty ratio maps when the engine (oil temperature) is hot, warm, and cold, respectively. The controller 100 uses these duty ratio maps depending on the oil temperature. That is, since the engine is in a cold state when the engine is started, the controller 100 sets the engine operating state (engine speed, engine load) based on the cold duty ratio map shown in FIG. 9C. Read the corresponding duty ratio. When the engine warms up and the oil reaches a predetermined oil temperature T1 or higher, the target duty ratio is read based on the duty ratio map during warming shown in FIG. When the oil temperature becomes equal to or higher than a predetermined oil temperature T2 (> T1), the target duty ratio is read based on the high-temperature duty ratio map shown in FIG.

  In this embodiment, the oil temperature is divided into three temperature ranges of high temperature, warm time, and cold time, and the duty ratio is read using a duty ratio map set in advance for each temperature range. Similarly to the hydraulic control map described above, more duty ratio maps may be prepared by dividing the temperature range more finely, or the target duty ratio may be calculated by proportional conversion according to the oil temperature. This makes it possible to control the pump capacity with higher accuracy.

  Next, a method for controlling the flow rate (discharge amount) of the pump 36 by the controller 100 will be described with reference to the flowchart of FIG.

  First, the engine 2 is started, and in order to grasp the operating state of the engine 2, the engine load, the engine speed, and the oil temperature are read from various sensors (step S1).

  Next, a duty ratio map stored in advance in the controller 100 is read, and a target duty ratio corresponding to the engine load, engine speed, and oil temperature read in step S1 is read (step S2).

  The target duty ratio read in step S2 is compared with the current duty ratio (step S3).

  If it is determined in step S3 that the current duty ratio has reached the target duty ratio, the process proceeds to the next step S5.

  If it is determined in step S3 that the current duty ratio has not reached the target duty ratio, the target duty ratio is output to the linear solenoid valve 49, and then the process proceeds to the next step S5 (step S4).

  Next, the current hydraulic pressure is read from the hydraulic pressure sensor 70 (step S5).

  Next, a pre-stored hydraulic control map is read, and a target hydraulic pressure corresponding to the current engine operating state is read from the hydraulic control map (step S6).

  The target hydraulic pressure read in step S6 is compared with the current hydraulic pressure (step S7).

  If it is determined in step S7 that the current hydraulic pressure has not reached the target hydraulic pressure, an output signal obtained by changing the target duty ratio to the linear solenoid valve 49 by a predetermined ratio is output, and the process returns to step S5 (step S8).

  If it is determined in step S7 that the current hydraulic pressure has reached the target hydraulic pressure, the engine load, engine speed, and oil temperature are read (step S9).

  Finally, it is determined whether the engine load, the engine speed, and the oil temperature have changed (step S10). If it is determined that the engine load has changed, the process returns to step S2, and if it has not been changed, the process returns to step S5. The above control is continued until the engine 2 is stopped.

  The flow rate control of the pump 36 described above is a combination of the feedforward control of the duty ratio and the feedback control of the hydraulic pressure. According to this flow rate control, the responsiveness is improved by the feedforward control and the accuracy is improved by the feedback control. Realized.

  Next, a method for controlling the number of cylinders by the controller 100 will be described below in accordance with the flowchart of FIG.

  First, the engine 2 is started, and the engine load, engine speed, and water temperature are read from various sensors in order to grasp the operating state of the engine (step S11).

  Next, based on the read engine load, engine speed, and water temperature, it is determined whether the current engine operating condition satisfies the valve stop operating condition (is in the reduced cylinder operating range) (step S12).

  If it is determined in step S12 that the valve stop operation condition is not satisfied (not in the reduced cylinder operation region), the four cylinder operation is performed (step S13).

  If it is determined in step S12 that the valve stop operation condition is satisfied, the first direction switching valves 34 and 35 connected to the VVTs 32 and 33 are operated (step S14).

  Next, the current cam angle is read from the cam angle sensor 74 (step S15).

  Next, the current operating angle of the VVTs 32 and 33 is calculated based on the read current cam angle, and it is determined whether the current operating angle is the target operating angle (step S16).

If it is determined in step S16 that the current operating angle of the VVTs 32 and 33 is not the target operating angle (θ ₁ ), the process returns to step S15 (step S14). That is, the operation of the second direction switching valves 46 and 47 is prohibited until the target operating angle is reached.

  If it is determined in S16 that the target operating angle has been reached, the second direction switching valves 46 and 47 are operated to perform a two-cylinder operation (step S17).

  Next, referring to FIG. 12, when the VVTs 32 and 33 are operating at the time of the reduced cylinder operation request when the engine operation state falls within the reduced cylinder operation region, the cylinder number control method shown in FIG. 11 is executed. A specific example will be described.

  At time t1, the first direction switching valves 34 and 35 of the VVTs 32 and 33 are operated. As a result, the supply of oil to the advance hydraulic chambers 326 and 336 of the VVTs 32 and 33 is started, and the operating angles of the VVTs 32 and 33 change (θ2 to θ1). As a result, the hydraulic pressure is lower than the valve stop request hydraulic pressure P1.

  Here, when the current engine operating state enters the reduced cylinder operating region and satisfies the valve stop operating condition, the operation of the VVT 32, 33 is continued until the operating angle of the VVT 32, 33 reaches the target operating angle θ1. That is, the valve stop mechanism 25b is not operated while the oil pressure is lower than the valve stop request oil pressure P1.

  At time t2, the operating angle of the VVT 32, 33 becomes the target operating angle θ1, and when the operation of the VVT 32, 33 is completed, the supply of oil to the advance hydraulic chambers 326, 336 of the VVT 32, 33 is terminated. Returns to the valve stop request hydraulic pressure P1.

  At time t3 after time t2 when the hydraulic pressure returns to the valve stop requesting hydraulic pressure P1, the second direction switching valves 46 and 47 are operated to supply hydraulic pressure to the valve stop mechanism 25b, and the engine is switched from the four-cylinder operation to the two-cylinder operation. Switch. As described above, after the advance angle control of the VVTs 32 and 33, the shift to the reduced cylinder (two cylinders) operation is performed. Therefore, the intake charge amount is increased by the advance angle control of the intake and exhaust valves 14 and 15, and the load is applied to the two cylinders. As a result, engine rotation fluctuations can be suppressed.

  FIG. 13 is an enlarged view of the configuration on the downstream side of the oil supply device 1 in FIG. 4, in which the intake side and the exhaust side are simplified and shown. As shown in the drawing, oil passages 67, 68, and 69 branch from a third communication passage 53 that leads to a main gallery 54 from which oil is discharged from the pump 36. Advance oil pressure chambers 326 and 336 and retard oil pressure chambers 325 and 335 are connected to the oil passages 67 and 68 via first direction switching valves 34 and 35, respectively. The oil passage 69 is connected to a valve stop mechanism 25b of the HLA 25 with a valve stop mechanism via a check valve 48 and second direction switching valves 46 and 47.

  The check valve 48 is energized by a spring so as to open when the hydraulic pressure in the third communication passage 53 exceeds the required hydraulic pressure of the valve stop mechanism 25b, and the oil flows only in one direction from the upstream side to the downstream side. To regulate. The check valve 48 opens with a hydraulic pressure larger than the required hydraulic pressure of the VVTs 32 and 33.

  Here, when the VVTs 32 and 33 are operated during the reduced cylinder operation for operating the valve stop mechanism 25b, the hydraulic pressure in the third communication passage 53 is reduced. However, the check valve 48 provided in the oil passage 69 reduces the valve stop mechanism. Since the flow of oil from 25 b to the third communication passage 53 upstream of the check valve 48 on the oil passage 69 is blocked, the valve stop mechanism 25 b on the oil passage 69 downstream of the check valve 48 is used. The required hydraulic pressure is ensured.

  According to the above embodiment, the operating state of the engine based on the highest required hydraulic pressure among the required hydraulic pressures of the hydraulic operating devices such as the VVTs 32 and 33, the valve stop mechanism 25b, and the oil jet 28 for each operating state of the engine. Since the current target hydraulic pressure is set from the hydraulic control map in which the target hydraulic pressure corresponding to the hydraulic pressure is set in advance, the hydraulic pressure of each hydraulic actuator and the oil injection pressure are set by matching the hydraulic pressure of the oil passage with this target hydraulic pressure. The required oil pressure can be secured. Further, since the oil pressure in the oil passage is feedback-controlled based on the detected value so as to realize the target oil pressure, the capacity of the pump 36 can be controlled with high accuracy. Therefore, further improvement in fuel consumption of the engine can be realized.

  Further, in the hydraulic control map, a correction hydraulic pressure that is higher than the highest required hydraulic pressure is set in a region adjacent to the operating region of the engine in which the valve stop mechanism 25b operates. Therefore, the pump 36 is controlled based on this hydraulic control map. As a result, the operation responsiveness of the valve stop mechanism 25b can be improved, the shift to the reduced cylinder operation can be promoted, and the fuel consumption reduction effect can be enhanced.

  Further, when the VVTs 32 and 33 are operated, particularly when the engine 2 is rotating at a low speed and the oil discharge amount from the pump 36 is small, the VVTs 32 and 33 on the intake side and the exhaust side are operated simultaneously. Although the hydraulic pressure of the third communication passage 53 that is communicated decreases, according to the present embodiment, while the VVTs 32 and 33 are operating during the reduced cylinder operation, the check valve 48 provided in the oil passage causes the third communication passage 53 to operate. Since the oil flow between the passage 53 and the valve stop mechanism 25b is shielded, the oil pressure in the oil passage is prevented from temporarily decreasing due to the operation of the VVTs 32 and 33. Accordingly, it is possible to prevent the oil pressure supplied to the valve stop mechanism 25b from being lowered and the valve stop mechanism 25b from malfunctioning, so that the reduced-cylinder operation that holds the intake valve 14 and the exhaust valve 15 in the stopped state cannot be performed. . Therefore, it is possible to further improve the fuel efficiency of the engine by changing the valve characteristics during the reduced cylinder operation.

  Further, when the hydraulic pressure of the third communication path 53 is equal to or higher than the required hydraulic pressure of the valve stop mechanism 25b, the check valve 48 opens, so that the hydraulic pressure of the oil passage 69 becomes the same as the hydraulic pressure of the third communication path 53, A hydraulic pressure higher than the required hydraulic pressure can be supplied to the valve stop mechanism 25b. On the other hand, when the hydraulic pressure of the third communication path 53 is less than the required hydraulic pressure of the valve stop mechanism 25b, the check valve 48 is closed, so that the hydraulic pressure of the oil path 69 is influenced by the hydraulic pressure of the third communication path 53. Without being received, the required oil pressure of the valve stop mechanism 25b is maintained. Therefore, the malfunction of the valve stop mechanism 25b can be prevented only by adding a simple configuration in which a spring biased check valve 48 is provided in the oil passage 69 without performing special control.

  Further, according to the present embodiment, when the VVT 32, 33 is operating when the reduced cylinder operation is requested, the valve stop mechanism 25b is operated after the operation of the VVT 32, 33 is completed. The valve stop mechanism 25b operates after the lowered hydraulic pressure rises again, and it is possible to prevent the valve stop mechanism 25b from malfunctioning due to insufficient hydraulic pressure. Therefore, the VVTs 32 and 33 and the valve stop mechanism 25b can be reliably operated simultaneously.

  Note that the present invention is not limited to the illustrated embodiments, and it goes without saying that various improvements and design changes can be made without departing from the scope of the present invention.

  For example, although the present embodiment is applied to an in-line four-cylinder gasoline engine, the number of cylinders of the present invention may be any number, and may be applied to a diesel engine. In this embodiment, the linear solenoid valve is used to control the pump 36. However, the present invention is not limited to this, and an electromagnetic control valve may be used.

  In the present embodiment, a check valve 48 is provided in the oil passage of the valve stop mechanism 25b, and the check valve 48 opens more than the required oil pressure of the valve stop mechanism 25b, and the required oil pressure of the VVTs 32 and 33. Although a valve that opens at a higher hydraulic pressure is used, only the prevention of malfunction of the valve stop mechanism 25b when there is a request to reduce the cylinder and the valve characteristic control request such that the operation periods of the valve stop mechanism 25b and the VVT 32, 33 overlap. If the purpose is to use a valve that opens at a hydraulic pressure higher than the required hydraulic pressure of the VVTs 32 and 33 as the check valve 48, this objective can be achieved. In place of such a check valve 48, a known electromagnetic control valve that can control opening and closing at a desired timing based on the operating angle of the VVTs 32 and 33 may be used.

  As described above, according to the present invention, in an engine for an automobile or the like, the hydraulically operated valve characteristics can be prevented while the malfunction of the valve stopping device is prevented even during the reduced cylinder operation due to the operation of the hydraulically operated valve stopping device. Since it is possible to control the appropriate valve characteristics by operating the control device and further improve the fuel efficiency of the engine, it is preferably used in the manufacturing industry of this type of engine.

DESCRIPTION OF SYMBOLS 1 Oil supply apparatus 2 Multi-cylinder engine 9 Crankshaft 14 Intake valve 15 Exhaust valve 18, 19 Camshaft 25b Valve stop mechanism (hydraulic actuated valve stop device)
28 Oil jet (oil injection valve)
32, 33 Variable valve timing mechanism (hydraulic actuated valve characteristic control device)
34, 35 First direction switching valve 36 Variable displacement oil pump 46, 47 Second direction switching valve 48 Check valve 49 Linear solenoid valve (pump control device)
53 Main Gallery (Main Oilway)
67, 68 Oil passage (first branch oil passage)
69 Oil passage (second branch oil passage)
70 Oil pressure sensor (oil pressure detection means)
100 controller (control means)
325, 335 Retarded hydraulic chamber 326, 336 Advanced hydraulic chamber

Claims (3)

An oil pump, a hydraulically operated valve characteristic control device that changes a characteristic of at least one of the intake valve and the exhaust valve by the oil pressure of the oil supplied from the pump, and of the intake valve and the exhaust valve by the oil pressure In a multi-cylinder engine control device comprising a hydraulically operated valve stop device that stops at least one of the valves and performs a reduced cylinder operation,
A first branch oil passage that branches off from a main oil passage through which oil is discharged from the oil pump and supplies oil to the hydraulically operated valve characteristic control device via a first direction switching valve;
A second branch oil passage that branches from the main oil passage and supplies oil to the hydraulically operated valve stop device via a second direction switching valve;
The hydraulically operated valve characteristic control device operates in the second branch oil passage between the branch point from the main oil passage to the second branch oil passage and the second direction switching valve during the reduced cylinder operation. A control apparatus for a multi-cylinder engine, comprising: a shielding valve that shields the second branch oil passage during the operation. The shielding valve is energized by a spring so as to open when the hydraulic pressure in the main oil passage exceeds the required hydraulic pressure of the hydraulically operated valve stop device, and allows oil flow only in one direction from the upstream side to the downstream side. 2. The control device for a multi-cylinder engine according to claim 1, wherein the control device is a check valve for regulating the multi-cylinder engine. The hydraulically operated valve characteristic control device includes an advance hydraulic chamber and a retard hydraulic chamber, and is configured to change the relative rotation phase of the camshaft with respect to the crankshaft by controlling the hydraulic pressure supplied to each hydraulic chamber. The control apparatus for a multi-cylinder engine according to any one of claims 1 and 2.

Images