JP6013223B2 - Hydraulic control device for engine - Google Patents

Description

  The present invention relates to a hydraulic control apparatus for an engine, and more particularly to control of hydraulic pressure when an oil jet that injects oil directly onto a piston is provided.

  Conventionally, an engine is generally provided with an oil passage for supplying engine oil to a lubricated portion such as a piston or a crank journal, and includes an oil jet that directly injects oil to the piston for cooling. In some cases. The oil pressure in such an oil passage is preferably controlled to be as low as possible within a range where necessary oil can be supplied to each part.

  For example, in the engine described in Patent Document 1, the engine state such as the temperature state and the load factor is monitored, and the hydraulic pressure is controlled so as to be as low as possible. In this way, for example, in a high-load high-rotation operation state where the amount of heat received by the piston is large, oil supply sufficient for cooling can be provided, while in a low-load or low-rotation operation state where the heat reception amount is small, the hydraulic pressure is reduced. Thus, the driving load of the oil pump can be reduced.

JP 2007-146839 A

  However, the state of the engine does not represent the state of the piston as it is, and particularly during a transition in which the operating state of the engine changes, a time lag occurs between the operating state and the temperature state of the piston. If only the state of the engine is taken into consideration as in the conventional example, the amount of oil injected to the piston by the oil jet may become excessive or conversely insufficient.

  For example, even if the engine speed or load decreases after high-load operation, the temperature of the piston does not decrease immediately, and it is necessary to inject a certain amount of oil to the piston by an oil jet for a while. is there. At this time, if the hydraulic pressure is reduced in accordance with the decrease in engine load as in the conventional example, the amount of oil injected into the piston is insufficient, the piston temperature becomes too high, and damage is caused. There is also a risk of giving.

  In view of this point, an object of the present invention is to improve the fuel consumption by reducing the driving load of the oil pump as much as possible while sufficiently securing the cooling performance of the piston by the oil jet.

  In order to achieve the above object, the present invention is characterized in that the control target value of the hydraulic pressure is set so as to be a suitable value corresponding to the temperature state of the piston estimated from the state of the engine. That is, the present invention is directed to an engine hydraulic control device having an oil jet that injects oil directly onto a piston. Based on the operating state of the engine, the higher the temperature state of the piston, the higher the value. Thus, a target oil pressure setting unit for setting the control target value of the oil pressure is provided.

  The target hydraulic pressure setting unit calculates a piston time constant representing a response delay of the temperature change of the piston with respect to a change in the operating state of the engine, and corrects the control target value of the hydraulic pressure using the piston time constant. Configured.

  According to the above-mentioned invention specific matter, during operation of the engine, basically, the hydraulic pressure is controlled to be higher as the temperature state of the piston is higher based on the operating state. That is, in a high-load high-rotation operation state where the amount of heat received by the piston is large, the hydraulic pressure becomes high, and sufficient oil for cooling the piston can be injected into the piston. On the other hand, in a low load or low rotation operation state in which the amount of heat received by the piston is small, the hydraulic pressure is low, and the driving load of the oil pump is reduced, thereby reducing fuel consumption.

  In addition, for example, when the engine operating state changes, the hydraulic pressure control is performed using the piston time constant so as to compensate for this delay in consideration of the delay in the change in piston temperature with respect to the change in the operating state. The target value is corrected. Thereby, even at the time of transition, an appropriate oil pressure corresponding to the actual temperature state of the piston can be set as the control target value.

  Since the piston time constant mainly depends on the heat capacity of the piston, a value adapted in advance through experiments and simulations may be used. However, the amount of heat received and radiated by the piston may vary depending on the operating state of the engine. It is preferable to reflect the value of the piston time constant.

  For example, the target hydraulic pressure setting unit may calculate the piston time constant based on the intake amount of the engine so that the smaller the intake amount, the larger the value. This is because when the amount of intake air is small, the amount of heat generated in the combustion chamber of the engine is also small, and the response delay of the piston temperature rise is relatively large. Further, when the intake air amount is small, the piston cooling action by the intake air at the time of fuel cut is weakened, so that the response delay of the temperature drop of the piston becomes relatively large.

  On the other hand, since the amount of heat generated in the combustion chamber increases when the intake air amount is large, the response delay of the temperature rise of the piston becomes relatively small, and the piston cooling action by the intake air at the time of fuel cut becomes strong. Therefore, in this case, the piston time constant may be set to a relatively small value.

  The target hydraulic pressure setting unit may calculate the piston time constant so that when the intake air amount is increased, the value is smaller than when the intake air amount is decreasing. In this way, the actual piston temperature changes during the acceleration transient when the piston heat received increases as the intake air volume increases and during the deceleration transient when the piston heat sink decreases as the intake air volume decreases. The hydraulic control target value can be suitably set so as to correspond to the above.

  In other words, the amount of heat generated in the combustion chamber rapidly increases during acceleration transients, so that the temperature of the piston receiving this rises relatively quickly and the response delay becomes relatively small. On the other hand, although the amount of heat received by the piston decreases rapidly during the deceleration transition, the increase in the amount of heat released is moderate compared to the increase in the amount of heat received during the acceleration transition. This is because the response delay is relatively large.

  Further, the target hydraulic pressure setting unit is configured to estimate a hydraulic control target value before correction using the piston time constant calculated as described above based on at least the engine load factor and the rotational speed of the piston. The higher the value, the higher the pressure value may be set. In this way, it is possible to suitably set the hydraulic control target value so as to correspond to the actual piston temperature in accordance with the amount of heat received.

  Further, the target hydraulic pressure setting unit sets the control target value of the hydraulic pressure so that the higher the estimated piston temperature state, the higher the value, taking into account at least one of the engine water temperature and the oil temperature. May be. In this way, it is possible to suitably set the hydraulic control target value so as to correspond to the actual piston temperature in accordance with the amount of heat radiation.

  The target hydraulic pressure setting unit sets a control target value of the hydraulic pressure so that the higher the estimated piston temperature state, the higher the pressure side value, taking into account the air-fuel ratio of the combustion chamber of the engine. Alternatively, the hydraulic control target value is set so that the higher the estimated piston temperature state, the higher the value, taking into account at least one of engine ignition timing, cylinder internal pressure, and exhaust temperature. It may be set.

  In addition, the target oil pressure setting unit may be configured to set the oil pressure control target value such that the higher the oil temperature of the engine, the higher the value on the high pressure side. This is because the higher the engine oil temperature, the lower the cooling performance of the piston by the oil jet.

  By the way, in order to control the hydraulic pressure of the engine so that it becomes the control target value set by the target hydraulic pressure setting unit as described above, it is preferable to equip the engine with a variable capacity oil pump and its capacity. Is changed based on the control target value of the oil pressure, and the discharge pressure of the oil pump is controlled. However, it is also possible to use a flow rate control valve or the like that can adjust the opening continuously without providing a variable capacity oil pump.

  According to the hydraulic control apparatus for an engine according to the present invention, when oil is directly injected into a piston by an oil jet, the higher the estimated piston temperature state, the higher the value on the basis of the engine operating state. In this way, by setting the control target value of the hydraulic pressure and correcting this control target value using the piston time constant, the hydraulic pressure is controlled appropriately so as to correspond to the actual temperature state of the piston even during a transition. The oil pump drive load can be reduced as much as possible and the fuel consumption can be improved while sufficiently ensuring the cooling performance of the piston.

It is a figure which shows the example of schematic structure of the engine which concerns on embodiment of this invention. It is sectional drawing which shows the structure of the combustion chamber of an engine, and its peripheral part. It is explanatory drawing which shows the outline of the oil supply system of an engine. It is sectional drawing which shows the structure of the oil pump of an engine, Comprising: A pump capacity | capacitance shows the state of the maximum. FIG. 5 is a view corresponding to FIG. 4 illustrating a state where the capacity of the oil pump is minimum. It is a graph which shows the relationship between the electric current command value to OCV, and a pump discharge pressure. It is a characteristic view which shows the relationship between the rotation speed and load factor of an engine, and piston temperature. It is a graph which shows an example of the rise delay of piston temperature at the time of the transition of acceleration. (a) is a flowchart showing an overall control operation of engine oil pressure control, and (b) is a flowchart showing setting of a target oil pressure. It is a figure which shows an example of the map which set the piston time constant corresponding to the engine intake air amount. FIG. 9B is a diagram corresponding to FIG. 9B, in which (a) shows the case where the oil temperature is taken into account, and (b) shows the case where the air-fuel ratio is further taken into account. FIG. 9 (b) is a view corresponding to FIG. 9 (b) according to another embodiment in which the cylinder internal pressure, the exhaust temperature or the ignition timing is further added to the flow of FIG. b) shows the case where the exhaust temperature is taken into account, and (c) shows the case where the ignition timing is taken into account.

  Hereinafter, embodiments of the present invention will be described with reference to the drawings. In this embodiment, the case where the present invention is applied to a diesel engine for automobiles will be described as an example, but the present invention is not limited to this. The description of this embodiment is merely an example, and does not limit the configuration or use of the present invention.

-Engine-
First, a schematic configuration of a diesel engine 1 (hereinafter also simply referred to as an engine 1) will be described. As schematically shown in FIG. 1, the engine 1 is an in-line four-cylinder engine as an example, and includes four cylinders 12 ( An intake passage 2 for supplying air (intake) to each of the cylinders) is connected to one side (upper side in the figure) of the cylinder head 1a. Further, an exhaust passage 3 for discharging burned gas (exhaust gas) from each of the four cylinders 12 is connected to the opposite side (lower side in the figure) of the cylinder head 1a.

  That is, as will be described later with reference to FIG. 2, the cylinder block 1b of the engine 1 is formed with four cylindrical cylinders 12 (only one is shown in the figure), and the pistons 13 are accommodated therein. Has been. A combustion chamber 11 is defined between the top surface 13a of the piston 13 and the lower surface of the cylinder head 1a, and communicates with the intake passage 2 and the exhaust passage 3 through an intake port 15 and an exhaust port 16, respectively. Yes.

  As shown in FIG. 1, the downstream side of the intake passage 2 (downstream side of the intake flow) is an intake manifold 2 a that distributes intake air to each cylinder 12. On the upstream side of the intake passage 2, an air cleaner 6 that filters air, a compressor impeller 24 of a turbocharger 20 that will be described later, an intercooler 7 that cools hot air compressed thereby, and an intake throttle valve (diesel throttle) 8 Etc. are arranged.

  Further, the upstream side of the exhaust passage 3 (upstream side of the exhaust flow) is an exhaust manifold 3a where the exhaust flows from the cylinders 12 merge, and on the downstream side thereof, the turbine wheel 22 of the turbocharger 20, the exhaust purification device. 9 etc. are arranged. As an example, the exhaust purification device 9 includes a NOx storage catalyst (NSR catalyst: NOx Storage Reduction catalyst) 9a and a DPNR catalyst (Diesel Particulate-NOx Reduction catalyst) 9b.

  The turbocharger 20 supercharges intake air by the force of the exhaust flow, and the turbine wheel 22 accommodated in the turbine housing 21 of the exhaust passage 3 is rotated by receiving the exhaust flow. On the other hand, a compressor impeller 24 is accommodated in the compressor housing 23 of the intake passage 2 and is connected to the turbine wheel 22 by a turbine shaft 25. As a result, the compressor impeller 24 rotates integrally with the turbine wheel 22 and sends out the intake air while compressing it (supercharging).

  The turbocharger 20 in the present embodiment is a variable nozzle type turbocharger, and the exhaust flow velocity can be changed by a plurality of nozzle vanes 26 arranged so as to surround the turbine wheel 22. The nozzle vane 26 is operated by an electric actuator 27 via a link mechanism (not shown) and the angle thereof is changed.

  Further, in the present embodiment, an EGR passage 10 is provided that branches from the vicinity of the connection portion of the exhaust passage 3 to the exhaust manifold 3 a and extends to the intake passage 2 to recirculate part of the exhaust to the intake passage 2. The EGR passage 10 is provided with an EGR valve 10a for adjusting the exhaust gas recirculation amount and an EGR cooler 10b for cooling the EGR gas.

  Here, the structure of the combustion chamber 11 of the engine 1 and its peripheral part will be described with reference to FIG. As shown in the figure, a piston 13 is accommodated in the cylinder 12 so as to reciprocate in the vertical direction in the drawing along the center line P thereof. The piston 13 is connected to a crankshaft (not shown) by a connecting rod 14 extending downward in the figure, and the reciprocating motion and the rotational motion of the crankshaft are mutually converted.

  On the other hand, a cavity (concave portion) 13 b is formed in a substantially central portion of the top surface 13 a of the piston 13 to constitute a part of the combustion chamber 11 partitioned in the cylinder 12. Further, an intake port 15 and an exhaust port 16 are opened on the lower surface of the cylinder head 1a so as to face the top surface 13a of the piston 13, and these openings are respectively formed by an intake valve 17 and an exhaust valve 18. It is designed to be opened and closed.

  The cylinder head 1 a is provided with an injector 19 so as to inject fuel directly into the combustion chamber 11. The injector 19 is, for example, a piezo injector, and is disposed in a substantially central upper portion of the combustion chamber 11 in a standing posture along the cylinder center line P. As shown in the figure, the piston 13 is positioned near the top dead center. Then, fuel is injected into the cavity 13b.

  Furthermore, in this embodiment, an oil jet 30 is provided so as to inject engine oil directly toward the piston 13. That is, in FIG. 2, an oil jet gallery 31 is formed at the lower part of the exhaust-side side wall of the cylinder block 1b shown on the left side so as to extend in the cylinder row direction (direction orthogonal to the paper surface in the figure), and communicates with this. Each cylinder 12 is provided with a piston jet nozzle 32 (hereinafter also simply referred to as a nozzle 32).

  In the example shown in the drawing, the nozzle 32 extends substantially horizontally from the side wall of the cylinder block 1b toward the inside of the crankcase, then curves upward in an L shape, and has an injection hole opened at its upper end. Yes. Engine oil supplied from an oil supply system 4 to be described later through an oil jet gallery 31 is injected from the injection hole at the tip (upper end) of the nozzle 32 toward the back side of the upper piston 13 (arrow in FIG. 2). O).

  However, although not shown, the nozzle 32 of the present embodiment has a built-in check valve. If the hydraulic pressure acting from the oil jet gallery 31 is less than a predetermined pressure, the above-described oil injection is performed. Absent. When the oil pressure from the oil jet gallery 31 becomes equal to or higher than a predetermined pressure, the check valve is opened and the inflowed oil is injected from the upper end of the nozzle 32 as described above.

  Further, as shown in FIG. 2, a water jacket w is formed on the side wall of the cylinder block 1b so as to surround the four cylinders 12, and the water temperature sensor 105 faces the water jacket w on the intake side (right side in the figure). It is arranged. Although only shown in FIG. 1, the engine 1 includes a crank angle sensor 101 that detects the rotation angle of the crankshaft to calculate the engine speed Ne, and a flow rate of intake air (intake amount Ga) flowing through the intake passage 2. An air flow meter 102 for detection is also provided.

  Signals output from various sensors and switches such as the crank angle sensor 101 and the air flow meter 102 are input to an ECU (Electronic Control Unit) 100 as indicated by a broken line in FIG. Although detailed description is omitted, the ECU 100 has a general configuration including a CPU, a ROM, a RAM, a backup RAM, and the like. For example, the ECU 100 outputs a command signal to the actuator of the diesel throttle 8 to control the intake air amount. A command signal is also output to an injector 19 not shown in FIG. 1 to control the fuel injection amount and the like.

  The ECU 100 also outputs a command signal to the electric actuator 27 of the turbocharger 20 and controls the oil pump 5 control valve (OCV 60: FIG. 3) to control the oil pressure of the oil supply system 4 as described below. The command signal is also output.

-Engine oil supply system-
FIG. 3 shows an outline of the oil supply system 4 of the engine 1. In the figure, the outline of the main body portion of the engine 1 composed of the cylinder head 1a and the cylinder block 1b is indicated by phantom lines. As shown in the figure, an oil pan 1c is attached to the lower part of the cylinder block 1b, and each cover of the engine 1 such as a piston 13 and a crank journal, or a cam journal of a valve system for driving the intake valve 17 and the exhaust valve 18 is used. The oil recirculated from the lubrication part is stored.

  An oil strainer 41 is disposed so as to be immersed in the stored oil, and the suction pipe 41 a is connected to the suction port 50 d of the oil pump 5. As will be described in detail later, the oil pump 5 is an internal gear pump including an external gear drive rotor 51 and an internal gear driven rotor 52 that mesh with each other, and an input shaft 5 a that penetrates through the center of the drive rotor 51. Driven by the rotation of the crankshaft, the oil in the oil pan 1c is sucked up through the oil strainer 41.

  On the other hand, the upstream end of the first oil passage 42 formed in the cylinder block 1 b communicates with the discharge port 50 e of the oil pump 5, and the downstream end of the first oil passage 42 is connected to the oil cooler 43. . The oil cooler 43 cools the oil by exchanging heat with the engine cooling water. The cooled oil flows through the second oil passage 44 formed in the cylinder block 1b and is sent to the oil filter 45.

  The oil filter 45 filters foreign matters and impurities in the oil by the filter element, and the oil thus filtered flows through the third oil passage 46 and is sent to the main gallery 47. The main gallery 47 is formed in the cylinder block 1b so as to extend, for example, in the cylinder row direction, and maintains a predetermined pressure of the oil sent as described above, and a plurality of oil passages branching therefrom. To distribute to the parts to be lubricated.

  For example, although not shown, oil is supplied to the crank journal by branch oil passages that branch at equal intervals in the longitudinal direction of the main gallery 47 and extend downward. Further, oil is supplied to the valve operating system of the cylinder head 1 a through a branch oil passage extending upward from the main gallery 47. Further, oil is supplied to the oil jet gallery 31 through a branch oil passage 48 that extends substantially horizontally from the main gallery 47.

  The hydraulic pressure of the main gallery 47 is maintained in a predetermined state in order to make the flow rate and hydraulic pressure of the oil distributed to the lubricated portion of the engine 1 appropriate. That is, the oil pressure sensor 103 is disposed in the main gallery 47, and a signal output from the oil pressure sensor 103 is input to the ECU 100, and the capacity of the oil pump 5 is variably controlled as described below. The oil pan 1c is provided with an oil temperature sensor 104, and a signal output from the oil temperature sensor 104 is also input to the ECU 100.

-Oil pump-
Next, the structure of the oil pump 5 will be described in detail with reference to FIG. In the illustrated example, the oil pump 5 holds an external gear drive rotor 51 rotated by an input shaft 5a, an internal gear driven rotor 52 rotated by meshing with the drive rotor 51, and the driven rotor 52 rotatably held from the outer periphery. The adjustment ring 53 is accommodated in the housing 50. The adjustment ring 53 changes the pump capacity by displacing the drive rotor 51 and the driven rotor 52 as will be described later.

  The housing 50 is a thick plate as a whole, and as shown in FIG. 4, the housing 50 has an elliptical shape that is long to the left and right in a plan view when viewed from the rear of the engine. Moreover, the protrusion part 50b is formed toward the downward direction from the lower left part of a figure, respectively. In addition, a recess 50c that is open toward the rear, that is, the inner side of the engine 1 (the front side in the drawing) is formed in the entire housing 50.

  The recess 50c accommodates the drive rotor 51, the driven rotor 52, the adjustment ring 53, and the like (hereinafter referred to as an accommodation recess 50c), and is closed by a cover (not shown) superimposed on the housing 50 from the rear. . A through hole (not shown) having a circular cross section is formed at a position slightly to the right of the center of the housing recess 50c, and the input shaft 5a inserted therethrough is rotatably supported.

  The input shaft 5a may be provided integrally with the front end portion of the crankshaft of the engine 1 or may be driven by a chain or the like as a separate body from the crankshaft. The input shaft 5a passes through the central portion of the drive rotor 51 and is fitted by, for example, a spline. The drive rotor 51 has a plurality of outer teeth 51a (11 in the illustrated example) having a trochoid curve or a curve approximated to a trochoid curve (for example, involute, cycloid, etc.) on the outer periphery.

  On the other hand, the driven rotor 52 is formed in an annular shape, and has inner teeth 52a that are one tooth larger than the inner teeth 52a (12 in the illustrated example) so as to mesh with the outer teeth 51a of the drive rotor 51. Is formed. The center of the driven rotor 52 is eccentric by a predetermined amount with respect to the center of the drive rotor 51, and the outer teeth 51 a of the drive rotor 51 and the inner teeth 52 a of the driven rotor 52 on the eccentric side (the upper left side in FIG. 4). Are engaged.

  The driven rotor 52 is slidably fitted and supported by an annular main body 53 a of the adjustment ring 53. In this example, the adjustment ring 53 has two projecting portions 53b and 53c that project radially outward from the outer periphery of the main body 53a in the circumferential direction over a predetermined angular range (about 50 ° in the illustrated example). In addition, an arm portion 53d extending greatly outward in the radial direction and a small protruding portion 53e are integrally formed.

  In this embodiment, a trochoid pump having 11 leaves and 12 nodes is configured by the drive rotor 51 and the driven rotor 52 held in the adjustment ring 53 in this manner, and the annular space between the two rotors 51 and 52 is formed in the annular space. Are formed with a plurality of working chambers R arranged in the circumferential direction between teeth engaged with each other. The volume of each working chamber R increases or decreases while moving along the outer periphery of the drive rotor 51 as the two rotors 51 and 52 rotate.

  That is, in a range extending from the position at which the teeth of the two rotors 51 and 52 mesh with each other to about 180 degrees in the rotor rotation direction indicated by the arrow in the drawing (the lower left side range in FIG. 4), the two rotors 51 and 52 rotate. Accordingly, the volume of the working chamber R gradually increases, and an intake range for sucking oil is obtained. On the other hand, in the remaining range of about 180 degrees (the upper right side range in FIG. 4), the volume of the working chamber R gradually decreases as the rotors 51 and 52 rotate, and oil is discharged while being pressurized. The discharge range.

  A suction port and a discharge port are formed in the housing 50 and the cover so as to correspond to the suction range and the discharge range, respectively. FIG. 4 shows only the suction port 50d and the discharge port 50e of the housing 50. The suction port 50d opens at the bottom surface of the housing recess 50c of the housing 50 so as to correspond to the suction region, and is also in the discharge region. The discharge port 50e is opened so as to correspond.

  The suction port 50d is located on the lower left side of the housing 50 in the drawing, and communicates with a suction port of a cover (not shown), and communicates with the suction line of the oil strainer via this. On the other hand, the discharge port 50e is located on the upper right side of the housing 50, communicates with a discharge port of a cover (not shown), and extends toward the right side of the drawing so as to correspond to the protruding portion 50a of the housing 50. It reaches the communication path 6a toward 45.

  With this configuration, the oil pump 5 rotates while the drive shaft 51 and the driven rotor 52 are engaged with each other due to the rotation of the input shaft 5a, and oil is sucked from the suction port 50d into the working chamber R formed between them. Compressed and discharged from the discharge port 50e. The flow rate of the oil basically increases as the rotational speed of the oil pump 5 (the rotational speed of the input shaft 5a), that is, the engine rotational speed Ne increases.

-Capacity variable mechanism-
The oil pump 5 of the present embodiment includes a variable capacity mechanism capable of changing the amount of oil discharged per rotation of the drive rotor 51, that is, the pump capacity. In the present embodiment, the adjustment ring 53 is displaced mainly by the hydraulic pressure (discharge pressure P) introduced mainly from the discharge port 50e, and the relative positions of the drive rotor 51 and the driven rotor 52 with respect to the suction port 50d and the discharge port 50e. By changing the flow rate of the oil sucked and discharged per one rotation.

  Specifically, as shown in FIG. 4, a pressing force from the compression coil spring 54 acts on the arm portion 53 d that extends radially outward from the main body portion 53 a of the adjustment ring 53, and thereby the adjustment ring 53. Is biased so as to be displaced slightly upward while rotating clockwise in the figure. The locus of the adjustment ring 53 at the time of such displacement is defined by the projecting portions 53b and 53c and the guide pins 55 and 56 engaged therewith.

  The adjusting ring 53 thus displaced divides the housing recess 50c into a high-pressure space TH on the upper right side in the drawing and a low-pressure space TL from the left side to the lower side, and is operated by receiving the hydraulic pressure in the high-pressure space TH. The That is, the high-pressure space TH is surrounded by the outer periphery of the protruding portion 53 c of the adjustment ring 53 and the wall portion of the housing 50 in the housing recess 50 c of the housing 50, and the first and second sealing members 57 and 58. Is formed in a region where the flow of oil is restricted.

  A part of the opening of the discharge port 50 e faces the high pressure space TH, and the discharge pressure P of the oil pump 5 is guided to the high pressure space TH and acts on the outer peripheral surface of the adjustment ring 53. On the other hand, since the atmospheric pressure is generally applied to the low pressure space TL that communicates with the suction port 50d, the adjustment ring 53 is attached so as to rotate counterclockwise in the figure by the hydraulic pressure from the high pressure space TH. Be forced.

  On the other hand, the adjustment ring 53 is urged clockwise by receiving the resilient force of the coil spring 54 acting on the arm portion 53d as described above. Therefore, for example, when the engine speed Ne is low as in idling, the adjustment ring 53 is urged to the maximum capacity position in FIG. 4 by the elastic force of the coil spring 54. At this time, per one rotation of the drive rotor 51 and the driven rotor 52, the amount of oil sucked from the suction port 50d and discharged from the discharge port 50e, that is, the pump capacity is maximized.

  When the engine speed Ne increases from this state, the discharge pressure P also tends to increase due to the increase in the oil discharge amount. Therefore, the adjustment ring 53 resists the resilience of the coil spring 54 in response to the oil pressure in the high-pressure space TH. As a result, it is displaced counterclockwise. As a result, the pump capacity is reduced, so that the increase in the discharge amount and the discharge pressure P is suppressed even if the rotational speed increases. As shown in FIG. 5, when the adjustment ring 53 is positioned at the minimum capacity position, the discharge amount per one rotation is minimized.

  Further, in the present embodiment, as shown in FIGS. 4 and 5, a control space TC is provided in the housing 50 so as to be adjacent to the high-pressure space TH, and an electronically controlled control valve 60 (Oil Control Vale: Hereinafter, the control hydraulic pressure from the OCV) is supplied. The hydraulic pressure in the control space TC generates a force that displaces the adjustment ring 53 in the direction in which the capacity of the oil pump 5 decreases as described above.

  Specifically, in the middle of the two overhanging portions 53b and 53c of the adjustment ring 53, a second sealing material 58 is disposed on the outer periphery thereof, and the inner surface of the wall portion of the housing 50 surrounding the housing recess 50c. It comes to be in sliding contact with. The second seal material 58 is a seal portion between the high-pressure space TH and the control space TC, and moves along the inner surface of the wall portion of the housing 50 in accordance with the displacement of the adjustment ring 53 as described above. become.

  Similarly, a third seal material 59 is disposed at the tip of the arm portion 53d of the adjustment ring 53 so as to be in sliding contact with the inner surface of the wall portion of the housing 50 facing the adjustment ring 53. The second and third sealing materials 58 and 59 and the first sealing material 57 described above all have the same thickness as the adjustment ring 53 (dimension in the direction perpendicular to the paper surface of FIGS. 4 and 5). It is made of a metal material or resin material having a size of about and excellent in wear resistance.

  Thus, the control space TC is surrounded by the outer periphery of the adjustment ring 53 (specifically, the outer periphery of the projecting portion 53b), the arm portion 53d, and the wall portion of the housing 50 facing them in the housing recess 50c of the housing 50, In addition, the second and third seal members 58 and 59 are formed in a region where the oil flow is restricted. Then, the control oil pressure is supplied from the OCV 60 through the control oil passage 61 that opens to the bottom surface of the housing recess 50c in the control space TC.

  One end of the control oil passage 61 opens as the round hole 61a facing the control space TC as described above, and the other end communicates with the control port 60a of the OCV 60. The OCV 60 receives a signal from the ECU 100, which will be described later, and the position of the spool is changed so that the oil from the supply port 60b is sent from the control port 60a to the control oil passage 61 and the oil discharged from the control oil passage 61 is discharged. The control port 60a is switched to the state of receiving and discharging from the drain port 60c.

  Further, as an example, in the OCV 60 that is a linear solenoid valve, the position of the spool continuously changes in response to a signal from the ECU 100, and the pressure of the oil sent from the control port 60a to the control oil passage 61 as described above (control oil pressure). Can be increased or decreased linearly. By adjusting the control oil pressure, the adjustment ring 53 can be displaced to adjust the capacity of the oil pump 5 as described below, and the discharge pressure P and thus the oil pressure of the main gallery 47 can be controlled.

  For example, when the control hydraulic pressure from the OCV 60 is increased, when the hydraulic pressure in the control space TC increases, the adjustment ring 53 is displaced counterclockwise in FIG. 4 and the pump capacity is reduced, so that the discharge amount of the oil pump 5 is reduced. The discharge pressure P decreases. On the other hand, when the control hydraulic pressure is lowered, the adjustment ring 53 is displaced clockwise in FIG. 4, the pump capacity is increased, and the discharge pressure P is also increased.

-Hydraulic control-
Next, the operation of the hydraulic control apparatus that suitably maintains the hydraulic pressure of the main gallery 47 by changing the capacity of the oil pump 5 and adjusting the discharge pressure P as described above will be described in more detail. FIG. 6 shows a relationship between a command signal from the ECU 100 to the OCV 60, that is, an OCV current command value (control duty as an example) and the discharge pressure P of the oil pump 5. From this figure, if the OCV current command value is increased, the discharge pressure P can be kept low even if the pump rotational speed is increased. On the other hand, if the OCV current command value is decreased, the discharge pressure P is increased. It can be seen that the discharge pressure P of 5 can be arbitrarily controlled.

  Here, in order to reduce the fuel consumption of the engine 1, it is preferable to reduce the driving load of the oil pump 5 as much as possible. On the other hand, in order to properly supply oil to the lubricated parts such as the piston 13 and the crank journal. Must maintain the hydraulic pressure of the main gallery 47 to some extent. That is, in order to supply a sufficient amount of oil and hydraulic pressure to the lubricated portion at a high rotation or high load of the engine 1, the main gallery 47 needs to be maintained at a high hydraulic pressure, while at a low rotation or low load. I want to reduce the drive load of the oil pump 5 by keeping the oil pressure low.

  Therefore, conventionally, the current command value to the OCV 60 is changed according to the operating state of the engine 1 (for example, the engine speed Ne, the intake air amount Ga, the load factor KL, etc.), and the capacity of the oil pump 5 is adjusted as described above. Therefore, it has been proposed to control the discharge pressure P as low as possible. Further, as shown in FIG. 7, between the operating state of the engine 1 and the temperature state of the piston 13 (piston temperature Tp), the higher the engine speed Ne and the higher the load factor KL, There is a relationship that the piston temperature Tp also increases.

  From this, if the discharge pressure P of the oil pump 5 is adjusted according to the operating state of the engine 1 as described above, the higher the piston temperature Tp is, the higher the hydraulic pressure of the main gallery 47 is maintained. In general, a necessary hydraulic pressure is supplied to the oil jet 30. Therefore, the amount of oil injected from the oil jet 30 increases at high rotations or loads with a large amount of heat received by the piston 13.

  On the other hand, if the discharge pressure P of the oil pump 5 is adjusted according to the operating state of the engine 1 as described above, the oil pressure of the main gallery 47 is naturally lowered as the piston temperature Tp is lower, and the oil pump is driven. The load can be reduced.

  However, the operating state of the engine 1 does not represent the piston temperature Tp as it is, and during a transition in which the operating state of the engine 1 changes suddenly more than a predetermined value, there is a time between the operating state and the piston temperature Tp. Therefore, if the oil pressure is controlled only in accordance with the operating state of the engine 1 as described above, the oil injection amount to the piston 13 by the oil jet 30 temporarily becomes excessive, or conversely insufficient. Sometimes.

  That is, for example, when the load factor KL of the engine 1 suddenly rises in response to depression of the accelerator pedal, the piston temperature Tp rises with a delay as shown by the solid line or broken line graph in FIG. In the example shown in the figure, the load factor KL rises in a step shape from time t1, and when the oil pressure is increased accordingly, the oil pump 5 is up to time t2 even though there is a response delay in increasing the piston temperature Tp. The driving load increases.

  Although illustration is omitted, for example, even if the accelerator pedal is released after operation at a high load and the load factor KL suddenly decreases, the piston temperature Tp does not decrease immediately, and the cooling of the piston is not performed for a while. Therefore, a certain amount of oil is required. At this time, if the hydraulic pressure is reduced in accordance with the reduction in the load factor KL, the amount of oil injected to the piston 13 by the oil jet 30 is insufficient, and the piston temperature Tp may temporarily rise excessively. Yes, there is a risk of damage.

  Considering this point, in the present embodiment, the ECU 100 (target hydraulic pressure setting unit) first sets the discharge pressure P of the oil pump 5 so as to correspond to the estimated piston temperature Tp based on the operating state of the engine 1. A target value Pt (target oil pressure) is calculated, and if the engine 1 is in a steady operation state, the discharge pressure P of the oil pump 5 is controlled based on the target oil pressure Pt.

  On the other hand, at the time of a transient in which the operating state of the engine 1 changes suddenly more than a predetermined value, the response time of the change of the piston temperature Tp with respect to the change of the operating state of the engine 1 is taken into account, and a piston time constant τ representing this response delay is used. The target hydraulic pressure Pt is corrected. Thereby, the suitable target oil pressure Pt corresponding to the actual piston temperature Tp at the time of transition can be set.

-Specific control action-
Hereinafter, the control (hydraulic control) of the oil pump 5 performed by the ECU 100 will be described in detail with reference to the flowchart of FIG. The illustrated control routine is executed at regular intervals (for example, about several milliseconds to several tens of milliseconds) during operation of the engine 1.

  FIG. 9A shows the overall control operation of the hydraulic pressure control. First, in step ST1 after the start, predetermined information regarding the operating state of the engine 1 is acquired. For example, the engine speed Ne is calculated from the signal from the crank angle sensor 101, the intake air amount Ga is calculated from the signal from the air flow meter 102, and the engine 1 is calculated from the engine speed Ne and the intake air amount Ga or the accelerator operation amount. The load factor KL is calculated.

  In the subsequent step ST2, based on the engine speed Ne, the intake air amount Ga, the load factor KL, etc., that is, based on the operating state of the engine 1, the target of the discharge pressure P of the oil pump 5 is described in detail later. A value (target oil pressure Pt) is set. In steps ST3 and ST4, the oil temperature and the oil pressure are calculated based on signals from the oil temperature sensor 104 and the oil pressure sensor 103, respectively (acquisition of oil temperature and oil pressure).

  Then, feedback control is performed so that the actual discharge pressure P of the oil pump 5 becomes the target hydraulic pressure Pt. That is, in step ST5, for example, a deviation between the actual pump discharge pressure P and the target oil pressure Pt is calculated from the information on the oil pressure, and the pump discharge pressure P is converged to the target oil pressure Pt according to the PID rule or the like according to this deviation. A target value of such pump capacity is calculated (feedback control calculation).

  In step ST6, the control hydraulic pressure supplied to the control space TC of the oil pump 5 is calculated so as to be the target value of the pump displacement, and the spool is operated so that the OCV 60 outputs this control hydraulic pressure. Command signal, that is, an OCV current command value (control duty), is output to the OCV 60 and the process returns. Thereby, the discharge pressure P of the oil pump 5 and the hydraulic pressure of the main gallery 47 are controlled.

  The correspondence relationship of parameters such as the pump capacity, control hydraulic pressure, and OCV current command value is previously adapted by experiments and simulations and stored in the ROM of the ECU 100 as a map. An OCV current command value for realizing a target pump capacity is calculated with reference to a simple map. Also, parameter correspondences can be set as calculation formulas instead of maps.

-Target oil pressure setting-
Next, the setting of the target hydraulic pressure Pt in step ST2 will be described in detail. FIG. 9B shows an example of a routine for setting the target oil pressure Pt. By executing this routine, the ECU 100 realizes a function as a target oil pressure setting unit.

  In step ST21 after the start in the figure, information on the intake air amount Ga is acquired, and in the subsequent step ST22, the engine water temperature is calculated based on a signal from the water temperature sensor 105 (acquisition of water temperature). Note that the information on the intake air amount Ga is calculated based on the signal from the air flow meter 102 as described above, temporarily stored in the RAM or backup RAM of the ECU 100, and updated every predetermined cycle. Similarly, the engine water temperature may be read temporarily stored in the RAM or the backup RAM.

  In step ST23, a piston time constant τ for calculating the target oil pressure Pt at the time of transition in which the operating state of the engine 1 suddenly changes to a predetermined value or more is calculated. That is, as described above, since the piston temperature Tp changes with a delay with respect to the change in the operating state of the engine 1 during the transition, the time constant τ is used to compensate for this delay. Since the piston time constant τ mainly depends on the heat capacity of the piston 13, a value adapted in advance may be used. However, the amount of heat received and radiated by the piston 13 may vary depending on the operating state of the engine 1. It is preferable to reflect the value of the constant τ.

  As an example, the piston time constant τ may be calculated with reference to a map shown in FIG. The response delay of the change in the piston temperature Tp has a correlation with the intake air amount Ga of the engine 1, and when the intake air amount Ga is small, the heat generation amount in the combustion chamber 11 of the engine 1 also decreases, and the response delay of the temperature increase of the piston 13 occurs. Is relatively large. Further, when the intake air amount Ga is small, the cooling action of the piston 13 by the intake air at the time of fuel cut becomes weak, so that the response delay of the temperature decrease becomes relatively large.

  On the other hand, when the intake air amount Ga is large, the amount of heat generated in the combustion chamber 11 also increases. Therefore, the response delay of the temperature rise of the piston 13 becomes relatively small, and the cooling action of the piston 13 by intake air at the time of fuel cut becomes strong. Therefore, in the map shown in FIG. 10, the piston time constant τ is set to a larger value as the intake air amount Ga is smaller, while the piston time constant τ is set to a smaller value as the intake air amount Ga is larger. Such a relationship between the piston time constant τ and the intake air amount Ga is adapted in advance through experiments and simulations, and is stored in the ROM of the ECU 100 as the map.

  More specifically, in the map of FIG. 10, the piston time constant τ is set to a smaller value when the temperature of the intake air Ga increases and when the temperature decreases. This is due to the acceleration transient of the engine 1 in which the amount of heat received by the piston 13 increases as the intake air amount Ga increases, and the deceleration transient in which the amount of heat received by the piston 13 decreases as the intake air amount Ga decreases. This corresponds to the response delay of the change in the piston temperature Tp.

  That is, during the acceleration transition of the engine 1, the amount of heat generated in the combustion chamber 11 rapidly increases, and the piston temperature Tp also rises relatively quickly in response to this, so the response delay of the change becomes relatively small. On the other hand, although the amount of heat received by the piston 13 rapidly decreases during the deceleration transition, the increase in the amount of heat release is moderate compared to the increase in the amount of heat received during the acceleration transient, and the piston temperature Tp tends to decrease gradually. The response delay is relatively large.

In step ST24, the target hydraulic pressure Pt is calculated. The method of calculating the target oil pressure Pt differs depending on whether the operating state of the engine 1 is steady or transient. The fact that the engine 1 is in a steady operating state can be determined from, for example, the amount of change per hour in the engine speed Ne and the load factor KL. First, if the engine 1 is in a steady operation state, the target oil pressure Pt is calculated in consideration of the engine water temperature based on the operation state, for example, the engine speed Ne and the load factor KL. This target oil pressure Pt is suitable for the piston temperature Tp in the steady operation state of the engine 1. Specifically, the target oil pressure Pt is calculated with reference to a map stored in the ROM of the ECU 100. Although not shown in the map, for example, the target hydraulic pressure Pt corresponding to the load factor KL of the engine 1 and the engine speed Ne is adapted in advance by experiments and simulations so as to reflect the relationship of FIG. It is set. In other words, a suitable target oil pressure Pt corresponding to the piston temperature Tp estimated based on the load factor KL of the engine 1 and the engine speed Ne is set.

  In the illustrated map, the target oil pressure Pt is set to a higher value as the load factor KL is higher and the engine speed Ne is higher. Therefore, the target oil pressure Pt is set higher as the estimated piston temperature Tp increases as the amount of heat generated in the combustion chamber 11 increases and the amount of heat received by the piston 13 increases. Further, since the heat radiation amount of the piston 13 varies depending on the engine water temperature, the target hydraulic pressure Pt is corrected based on the engine water temperature in the present embodiment.

  That is, although not shown in the drawings, the target hydraulic pressure Pt is corrected so as to become a higher value as the engine water temperature is higher, according to a map or calculation formula that has been adapted in advance through experiments and simulations. In other words, in the present embodiment, the target hydraulic pressure Pt is basically set based on the load factor KL of the engine 1 and the engine speed Ne, and the estimated piston temperature Tp increases as the engine water temperature increases. Thus, the value is corrected to the value on the high pressure side.

  By the way, instead of such a steady operation state, for example, at the time of a transient in which the operation state of the engine 1 changes greatly by a predetermined value or more during acceleration operation or deceleration operation, the piston temperature Tp is delayed with respect to the change of the operation state. Will change. Therefore, the target hydraulic pressure Pt calculated (and corrected) as described above is further corrected using the piston time constant τ so as to compensate for the delay.

  Specifically, for example, based on the operating state of the engine 1 when switching from the steady state to the transient state, the target oil pressure Pt calculated from the map of the target oil pressure Pt and corrected according to the engine water temperature is set as the basic value Ptb. To do. Then, by repeating the calculation of the following formula (1) using the basic value Ptb as an initial value of the target hydraulic pressure Pt, an appropriate target hydraulic pressure Pt can be obtained even during a transient when the operating state of the engine 1 changes more than a predetermined value. Can be calculated.

Pt (n) <-Pt (n-1) + (Ptb (n) -Pt (n-1)) / τ (1)
In the above equation (1), Pt (n) is the target oil pressure Pt in a certain control cycle (n), that is, the current value of the target oil pressure Pt to be calculated, and Pt (n-1) is the previous control cycle. This is the calculated target oil pressure (previous value). Ptb (n) is a basic value Ptb of the target hydraulic pressure Pt calculated from the target hydraulic pressure map and corrected by the engine water temperature in the current control cycle, and τ is a piston time constant. As described above, the initial value of the target hydraulic pressure Pt (n) is the initial value of the basic value Ptb (n).

  By sequentially calculating the target oil pressure Pt using the above equation (1), even when acceleration or deceleration transients in which the operating state changes greatly by more than a predetermined value, the amount of delay in changing the piston temperature Tp is compensated, It is possible to calculate an appropriate target hydraulic pressure Pt that matches the actual piston temperature Tp. If the piston time constant τ = 1, Pt (n) ← Ptb (n) in the above equation (1).

  Therefore, according to the hydraulic control device for the engine 1 according to the present embodiment, first, the target oil pressure Pt is set according to the operating state of the engine 1 (including the engine water temperature, etc.). Control can be performed so that the discharge pressure P of the pump 5 and thus the hydraulic pressure of the main gallery 47 is increased. As a result, a sufficient amount of oil and hydraulic pressure can be supplied to the oil jet gallery 31, and sufficient oil can be injected from the piston jet nozzle 32 toward the piston 13 for cooling.

  In particular, when the operating state of the engine 1 is in a transient state of acceleration or deceleration where the operating state suddenly changes, the piston time constant τ that compensates for the delay in the change in the piston temperature Tp with respect to the change in the operating state is used. By appropriately correcting the hydraulic pressure Pt, it is possible to realize an appropriate oil jet supply corresponding to the actual piston temperature Tp even during a transition. Therefore, not only in the steady operating state of the engine 1, but also during acceleration and deceleration transitions, the driving load of the oil pump 5 is reduced as much as possible while sufficiently ensuring the cooling performance of the piston 13 to improve fuel efficiency. Can be planned.

(Other embodiments)
As mentioned above, although embodiment demonstrated demonstrated the case where this invention was applied as a hydraulic control apparatus of the in-line 4-cylinder diesel engine 1 for motor vehicles, this invention is not limited to this, As a hydraulic control apparatus of engines other than a motor vehicle Is also applicable. Of course, the present invention is not limited to the number of cylinders or the type of engine (V type, horizontally opposed type, etc.), and is applicable to a gasoline engine.

In the above-described embodiment, the target oil pressure Pt is calculated based on the operating state of the engine 1 (as an example, the load factor KL and the rotational speed Ne) in consideration of the engine water temperature as shown in FIG. 9B. However, the target oil pressure Pt may be calculated by more accurately reflecting the amount of heat received and the amount of heat released by the piston 13.
For example, as shown in the flow of FIG. 11 (a), after acquiring the engine water temperature in step ST22, the oil temperature of the engine 1 is acquired in step ST221, and this is also taken into account to calculate the target oil pressure Pt. Also good. Also in this case, the target oil pressure Pt may be corrected so as to be higher as the oil temperature is higher, according to a set map or arithmetic expression that is adapted in advance through experiments and simulations.

  Further, as shown in the flow of FIG. 11 (b), information on the air-fuel ratio (excess air ratio) in the cylinder 12 is further acquired in step ST222, and this is also taken into account to calculate the target oil pressure Pt. Good. The air-fuel ratio information is calculated from the load factor KL and the fuel injection amount, temporarily stored in the RAM or backup RAM of the ECU 100, and updated every predetermined cycle.

  In the case of the diesel engine 1 as in the above embodiment, since the heat generation amount increases as the excess air ratio λ is closer to the stoichiometric value λ = 1, the smaller the excess air ratio λ is, the more the air excess ratio λ is estimated. The target oil pressure Pt may be set to a value on the high pressure side in response to the piston temperature Tp becoming higher. Similarly, in the case of a gasoline engine, the target oil pressure Pt may be set to a value on the high pressure side in response to the estimated piston temperature Tp becoming higher as the air fuel ratio is closer to the stoichiometric air fuel ratio. In other words, the target hydraulic pressure Pt is corrected according to the air-fuel ratio (excess air ratio) as described above according to a map or calculation formula that is adapted in advance through experiments and simulations.

  Further, when an in-cylinder pressure sensor for detecting the pressure in the combustion chamber 11 is disposed in the engine 1, as shown in the flow of FIG. 12 (a), after acquiring the air-fuel ratio information in step ST222, the step In ST223, the cylinder internal pressure may be acquired from the signal from the cylinder internal pressure sensor, and the target temperature Tp may be calculated in consideration of this. In this case, the target hydraulic pressure Pt may be set to a high-pressure side value in response to the estimated piston temperature Tp increasing as the cylinder internal pressure increases.

  Further, when an exhaust temperature sensor is disposed in the exhaust passage 3 of the engine 1, as shown in the flow of FIG. 12 (b), after acquiring air-fuel ratio information in step ST222, the exhaust temperature is detected in step ST224. The target temperature Tp may be calculated by acquiring the exhaust temperature from the sensor signal and taking this into consideration. In this case, the lower the exhaust gas temperature, the higher the estimated piston temperature Tp. Therefore, if the target oil pressure Pt is corrected to a value on the high pressure side according to a set map or an arithmetic expression, it is adapted in advance through experiments and simulations. Good.

  Further, when the engine 1 is a gasoline engine, as shown in the flow of FIG. 12 (c), after obtaining the air-fuel ratio information in step ST222, the ignition timing information is obtained in step ST225. Thus, the target temperature Tp may be calculated. In this case, the estimated piston temperature Tp becomes higher as the ignition timing is on the advance side. Therefore, the target oil pressure Pt is set to a value on the high-pressure side according to a set map or arithmetic expression, which is adapted in advance through experiments and simulations. It is sufficient to correct it.

  Furthermore, in the above-described embodiment, the structure in which the drive rotor 51 and the driven rotor 52 are displaced by the adjustment ring 53 and the discharge amount per rotation of the input shaft 5a is changed as the variable capacity mechanism of the oil pump 5 has been described. It is not limited to such a structure. The oil pump is not limited to a gear pump regardless of whether it is an internal type or an external type, and may be a variable displacement type.

  Further, the invention is not limited to a structure equipped with a variable displacement type oil pump 5 for controlling the hydraulic pressure of the engine 1. For example, a linear solenoid valve or the like adjusts the opening steplessly to linearly adjust the hydraulic pressure. An oil amount and oil pressure of oil discharged from an oil pump having a fixed capacity may be adjusted using a valve that can be used.

  In the above-described embodiment, the target oil pressure Pt may be further corrected so as to increase as the oil temperature of the engine 1 increases. This is because the cooling performance of the piston 13 by the oil jet 30 decreases as the oil temperature of the engine 1 increases.

  According to the hydraulic control of the present invention, the driving load of the oil pump can be reduced as much as possible and the fuel consumption can be improved while sufficiently ensuring the cooling performance of the piston of the engine. is there.

1 Engine 5 Variable displacement oil pump 13 Piston 30 Oil jet 31 Oil jet gallery (oil jet)
32 Piston jet nozzle (oil jet)
100 ECU (target oil pressure setting unit)
Ga Intake amount Ne Engine speed Tp Piston temperature (piston temperature state)
P Oil pump discharge pressure Pt Target oil pressure (Hydraulic control target value)
τ Piston time constant

Claims (9)

An engine hydraulic control device having an oil jet that injects oil directly into a piston,
A target oil pressure setting unit that sets a control target value of oil pressure so that the higher the temperature state of the piston, the higher the value on the high pressure side, based on the operating state of the engine
The target hydraulic pressure setting unit calculates a piston time constant representing a response delay of a temperature change of the piston with respect to a change in the operating state of the engine, and corrects the control target value of the hydraulic pressure using the piston time constant. The engine hydraulic control device. The engine hydraulic control device according to claim 1,
The target hydraulic pressure setting unit calculates the piston time constant based on the intake air amount of the engine so that the piston time constant becomes larger as the intake air amount becomes smaller. The engine hydraulic control device according to claim 2,
The target hydraulic pressure setting unit calculates the piston time constant so that the piston time constant is smaller when the intake air amount is increasing than when it is decreasing. The engine hydraulic control apparatus according to any one of claims 1 to 3,
The target hydraulic pressure setting unit sets a control target value of hydraulic pressure so that the higher the estimated temperature state of the piston, the higher the pressure side value, based on at least the load factor and the rotational speed as the engine operating state. Hydraulic control device for the engine. The engine hydraulic control device according to claim 4,
The target hydraulic pressure setting unit sets the control target value of the hydraulic pressure so that the higher the estimated piston temperature state, the higher the pressure side value, taking into account at least one of the engine water temperature and the oil temperature. Hydraulic control device for the engine. The engine hydraulic control apparatus according to any one of claims 4 and 5,
The target hydraulic pressure setting unit sets an oil pressure control target value so that the higher the estimated piston temperature state, the higher the pressure side value, taking into account the air-fuel ratio of the engine combustion chamber. Hydraulic control device. The engine hydraulic control device according to any one of claims 4 to 6,
The target hydraulic pressure setting unit controls the hydraulic pressure so that the higher the estimated piston temperature state, the higher the value, taking into account at least one of engine ignition timing, cylinder internal pressure, and exhaust temperature. Hydraulic control device for the engine that sets the target value. The engine hydraulic control apparatus according to any one of claims 1 to 7,
The target hydraulic pressure setting unit sets an oil pressure control target value such that the higher the engine oil temperature, the higher the hydraulic pressure value. The engine hydraulic control apparatus according to any one of claims 1 to 8,
The engine is equipped with a variable capacity oil pump,
An engine hydraulic control device configured to change a capacity of the oil pump based on a hydraulic control target value set by the target hydraulic pressure setting unit and to control a discharge pressure thereof.

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