Variable compression ratio system for internal combustion engine and method for controlling the system

A variable compression ratio system for an internal combustion engine, including a variable compression ratio mechanism for continuously varying a compression ratio of the engine, the variable compression ratio mechanism including a control shaft rotatably moveable to a rotational position corresponding to the compression ratio, a hydraulic actuator driving the control shaft to the rotational position depending on operating conditions of the engine, a hydraulic pressure source mechanically driven by the engine to produce a hydraulic pressure supplied to the hydraulic actuator, and a hydraulic control for variably controlling the hydraulic pressure supplied to the hydraulic actuator on the basis of the engine operating conditions.

BACKGROUND OF THE INVENTION

The present invention relates to a variable compression ratio system for an internal combustion engine which is capable of continuously and variably controlling a compression ratio of the engine depending on engine operating conditions, and a method for controlling the system.

U.S. Pat. No. 6,491,003 (corresponding to Japanese Patent Application First Publication No. 2002-115571) discloses a variable compression ratio system for a reciprocating internal combustion engine. The variable compression ratio system uses a multiple-link type piston-crank mechanism for varying a position of a piston bottom dead center (BDC). The multiple-link type piston-crank mechanism includes upper and lower links linking a piston pin of a piston to a crankpin, and a control link linking the lower link to an eccentric cam of a control shaft. An actuator drives the control shaft to vary the rotational position depending on the engine operating conditions, whereby the compression ratio is variably controlled. The actuator may be an electric actuator, namely, an electric motor, or a hydraulic actuator.

SUMMARY OF THE INVENTION

In such a variable compression ratio system as the above-described related art, a load applied to the control link during the engine operation is transmitted to the eccentric cam of the control shaft to cause a rotation moment acting on the control shaft. The actuator, therefore, is required to drive the control shaft in the rotation direction against the rotation moment during the compression ratio varying operation and during the compression ratio holding operation. This causes increase in energy consumed for driving the actuator. Especially, in a case where the electric motor is used, the energy consumption will be more increased due to a low efficiency in converting the power output of the engine to that of the electric motor.

Further, a force applied to the control shaft is largely influenced by a combustion pressure produced when combustion takes place in the engine cylinder, and is varied depending on engine load. When the engine load is large even though the engine speed is low, a large rotation moment is applied to the control shaft. Therefore, in a case where the hydraulic actuator is used, the hydraulic actuator must be designed to produce a large output using a high hydraulic pressure so as to operate the control shaft against the large rotation moment. However, if such a high hydraulic pressure is used, a leakage from the hydraulic actuator and other parts, for instance, a selector valve, will be increased. This causes undesired increase in energy loss.

Further, torque required for rotating the control shaft upon controlling the compression ratio varies depending on engine speed and engine load. For instance, the required torque is small in a low-speed and low-load range of the engine. In such a case, the leakage from the hydraulic actuator, the selector valve and the like can be suppressed by reducing the hydraulic pressure supplied from the oil pump to the hydraulic actuator to a necessary and sufficient extent. This decreases the energy loss caused due to the leakage. Meanwhile, an amount of hydraulic fluid leaking from clearances varies in proportion to a square of a hydraulic pressure thereof. Further, if a hydraulic pressure is reduced upon supplying an amount of hydraulic fluid to the hydraulic actuator, energy consumption in driving the hydraulic actuator becomes smaller than that in a case where the hydraulic pressure is not reduced.

It is an object of the present invention to provide a variable compression ratio system for an internal combustion engine, which includes a variable compression ratio mechanism for continuously varying a compression ratio of the engine and a hydraulic actuator for driving the variable compression ratio mechanism depending on operating conditions of the engine, which is capable of reducing energy consumption required for driving the hydraulic actuator.

In one aspect of the present invention, there is provided a variable compression ratio system for an internal combustion engine, comprising:a variable compression ratio mechanism for continuously varying a compression ratio of the internal combustion engine, the variable compression ratio mechanism including a control shaft rotatably moveable to a rotational position corresponding to the compression ratio;a hydraulic actuator driving the control shaft to the rotational position depending on operating conditions of the internal combustion engine;a hydraulic pressure source mechanically driven by the internal combustion engine to produce a hydraulic pressure supplied to the hydraulic actuator; andhydraulic control means for variably controlling the hydraulic pressure supplied to the hydraulic actuator on the basis of the operating conditions of the internal combustion engine.

In a further aspect of the invention, there is provided a method for controlling a variable compression ratio system for an internal combustion engine, the variable compression ratio system including a variable compression ratio mechanism for continuously varying a compression ratio of the internal combustion engine, a hydraulic actuator driving the variable compression ratio mechanism, and a hydraulic pressure source mechanically driven by the internal combustion engine to produce a hydraulic pressure, the hydraulic actuator being supplied with the hydraulic pressure from the hydraulic pressure source via a hydraulic passage extending therebetween, the method comprising:detecting operating conditions of the internal combustion engine;determining a predetermined hydraulic pressure to be supplied to the hydraulic actuator on the basis of the detected operating conditions of the internal combustion engine;detecting a hydraulic pressure within the hydraulic passage; andcontrolling the hydraulic pressure supplied to the hydraulic actuator to the predetermined hydraulic pressure on the basis of the detected hydraulic pressure within the hydraulic passage.

DETAILED DESCRIPTION OF THE INVENTION

Referring toFIG. 1, there is shown a multiple-link type variable compression ratio mechanism10linked with a reciprocating internal combustion engine. Variable compression ratio mechanism10is operated by a hydraulic actuator explained later, so as to continuously vary a compression ratio of the engine. Here, the compression ratio is defined as the ratio of the volume in engine cylinder6above piston1when piston1is at bottom-dead-center (BDC) to the volume in engine cylinder6above piston1when piston1is at top-dead-center (TDC). Cylinder block5includes engine cylinders6one of which is illustrated inFIG. 1. Piston1is slidably disposed within engine cylinder6. Piston1defines a combustion chamber within engine cylinder6to thereby undergo a combustion pressure that is produced when combustion takes place in the combustion chamber. Crankshaft3is rotatably supported on cylinder block5via crankshaft bearing bracket7. Supercharger9may be used in the engine. Upper link11has one end pivotally coupled to piston1via piston pin2and an opposite end rotatably coupled to one end of lower link13via connecting pin12. Lower link13has a central portion pivotally supported on crankpin4of engine crankshaft3.

Lower link13has the other end to which one end of control link15is rotatably coupled to via connecting pin14. Control link15has an opposite end pivotally supported on a portion of the engine body integrally formed with cylinder block5. In order to vary the compression ratio of the engine, a pivot of the pivotal movement of the opposite end of control link15is arranged to be displaceable relative to the engine body. Specifically, control shaft18extending parallel to crankshaft3is provided with a generally cylindrical-shaped eccentric cam19whose center axis16is eccentric to a center axis of control shaft18. The opposite end of control link15is rotatably fitted to an outer circumferential surface of eccentric cam19. Control shaft18is rotatably supported between crankshaft bearing bracket7and control shaft bearing bracket8.

When control shaft18is rotated in order to vary the compression ratio, center axis16of eccentric cam19serving as the pivot of control link15is displaced relative to the engine body. Owing to the displacement of the pivot of control link15, the movement of each of lower link13and upper link11are varied. This causes change in stroke of piston1to thereby vary the compression ratio of the engine.

Referring now toFIG. 2, a relationship between a direction of movement of control shaft18and the compression ratio is explained. Reference characters Pc and Pe denote the center axis of control shaft18and the center axis of eccentric cam19, respectively. As control shaft18is rotated, center axis Pe of eccentric cam19is displaced around center axis Pc of control shaft18. In an initial position shown inFIG. 2, center axis Pe of eccentric cam19is positioned on the left side of center axis Pc of control shaft18. When control shaft18is rotated in direction A, namely, a clockwise direction, center axis Pe of eccentric cam19upwardly moves and control link15is also moved upwardly as indicated by arrow B. The movement of control link15causes lower link13to pivotally move in direction C, namely, a counterclockwise direction. The pivotal movement of lower link13causes upper link11to move downwardly as indicated by arrow D. As a result, piston1is moved downwardly as indicated by arrow E, so that the compression ratio is reduced. Namely, when control shaft18is rotated in the clockwise direction to move from the initial position shown inFIG. 2, the compression ratio is reduced. On the other hand, when control shaft18is rotated in the counterclockwise direction to move from the initial position shown inFIG. 2, the compression ratio is increased.

Referring toFIG. 3, there is shown a hydraulic circuit for operating hydraulic actuator31which drives control shaft18in a rotation direction. In this embodiment, hydraulic actuator31is in the form of a double acting piston-cylinder mechanism including rod51which is linearly moveable in an axial direction thereof. A pair of levers50are fixedly arranged on control shaft18with a predetermined space therebetween in an axial direction of control shaft18. Each of levers50has slit50aextending in a radial direction of control shaft18. Lever50and rod51are coupled to each other via generally cylindrical pin52which is moveably received in slit50a. Specifically, pin52has two parallel surfaces52ain a diametrically opposed relation to each other. Parallel surfaces52aare formed on a circumferential surface of each of the opposite end portions of pin52so as to be slidably engaged in slit50aof lever50. Pin52has a cylindrical middle portion rotatably supported in pin hole51bwhich is formed on one axial end portion51aof rod51. Rod51has large-diameter portion51cslidably fitted to sleeve54aextending outwardly from actuator housing54. Rod51has disk-shaped piston53at an end of large-diameter portion51cwhich is axially opposed to one axial end portion51awith pin hole51b. Actuator housing54is divided by piston53into first oil chamber55positioned on the side of control shaft18and second oil chamber56positioned on the side opposite to control shaft18. Rod51extends through first oil chamber55and sleeve54atoward control shaft18.

Hydraulic actuator31is operated by hydraulic pressure discharged from oil pump60acting as a hydraulic pressure source. Oil pump60has hydraulic fluid and is mechanically coupled to and driven by crank pulley63of the engine via belt64to produce the hydraulic pressure supplied to hydraulic actuator31. First and second oil chambers55and56of hydraulic actuator31are fluidly communicated with oil pump60and oil pan68via hydraulic path therebetween. Directional control valve59is disposed within the hydraulic path and electronically connected to engine control unit (ECU)40, hereinafter referred to as a controller. Directional control valve59is operative to switch supply of the hydraulic pressure discharged from oil pump60to hydraulic actuator31. In this embodiment, directional control valve59is in the form of a four-port three-position solenoid-operated valve. Directional control valve59selectively allows the fluid communication between each of first and second oil chambers55and56and oil pump60and the fluid communication between each of first and second oil chambers55and56and oil pan68.

Specifically, directional control valve59is connected with first oil chamber55via hydraulic passage57and with second oil chamber56via hydraulic passage58. Directional control valve59is also connected with a discharge port of oil pump60via supply passage61and with oil pan68via drain passage62. Directional control valve59has a first open position where the fluid communication between first oil chamber55and oil pump60and the fluid communication between second oil chamber56and oil pan68are established. Directional control valve59has a second open position where the fluid communication between first oil chamber55and oil pan68and the fluid communication between second oil chamber56and oil pump60are established. Directional control valve59has a closed position where the fluid communication between each of first and second oil chambers55and56and each of oil pump60and oil pan68are blocked. Directional control valve59is controlled by controller40to shift between the first and second open positions and the closed position.

Variable relief valve66is disposed within relief passage65branched from supply passage61. Variable relief valve66is electronically connected to controller40and operated to release an amount of the hydraulic fluid discharged from oil pump60. Pressure sensor67is arranged to detect the hydraulic pressure in the hydraulic path upstream of selector valve59, namely, in supply passage61. Pressure sensor67is electronically connected to controller40and operated to transmit signal Ps indicative of the detected hydraulic pressure in supply passage61.

In addition to pressure sensor67, a plurality of sensors are electronically connected to controller40. The sensors includes engine speed sensor42, intake air flow sensor44, and control shaft angle sensor46. Engine speed sensor42detects engine speed, i.e., the number of engine revolution, and generates signal Ne indicative of the detected engine speed. Engine speed sensor42may be a crank angle sensor. Intake air flow sensor44detects an amount of intake air flowing into the combustion chamber of the engine and generates signal Qa indicative of the detected intake air amount. Intake air flow sensor44may be an intake airflow meter. Control shaft angle sensor46detects a rotational angle of control shaft18and generates signal εr indicative of the detected rotational angle. Controller40receives signals Ne, Qa and εr generated from sensors42,44and46and processes signals Ne, Qa and εr to obtain engine operating conditions. Depending on the engine operating conditions, controller40executes various controls including control of selector valve59. Controller40may be a microcomputer including a central processing unit (CPU), input and output ports (I/O), a read-only memory (ROM) as an electronic storage medium for executable programs and calibration values, a random access memory (RAM), a keep alive memory (KAM), and a common data bus.

Controller40executes feedback control based on signal εr generated by control shaft angle sensor46and transmits the control signal to selector valve59. In response to the control signal, selector valve59shifts between the open positions so that the pressurized hydraulic fluid produced by oil pump60is introduced into one of first and second oil chambers55and56, and at the same time, the hydraulic fluid within the other of first and second oil chambers55and56is drained. This causes pressure difference between first and second oil chambers55and56to thereby move piston53and rod51of hydraulic actuator31closer to control shaft18and away therefrom. As a result, control shaft18is driven to a desired rotational position corresponding to a target compression ratio.

Controller40is programmed to determine a desired opening degree of variable relief valve66based on signal Ps generated by pressure sensor67. Namely, controller40is programmed to determine the amount of hydraulic fluid which is released through variable relief valve66when detected hydraulic pressure Ps within supply passage61is more than target hydraulic pressure Pt. Controller40transmits a control signal to variable relief valve66. In response to the control signal, variable relief valve66is operated to the desired opening degree to release the amount of hydraulic fluid into oil pan68. The hydraulic pressure within supply passage61is thus adjusted at target hydraulic pressure Pt.

Controller40is programmed to determine target hydraulic pressure Pt by selecting a larger one of a first hydraulic pressure required for satisfying responsivity of control shaft18upon varying the compression ratio of the engine and a second hydraulic pressure required for holding control shaft18at the rotational position to maintain the compression ratio of the internal combustion engine. The first hydraulic pressure is determined by calculating an amount of hydraulic fluid to be supplied to hydraulic actuator31during a target response period in which control shaft18must be operated from a certain stationary position to a rotational position. The responsivity of control shaft18is required for the main purpose of preventing occurrence of knocking when the engine load is increased. In order to prevent the occurrence of knocking, the compression ratio must be varied from a larger side to a smaller side. Upon the variation of the compression ratio, control shaft18is rotated in the same direction as the rotation moment applied thereto due to the combustion pressure generated in the combustion chamber of the engine. Therefore, the responsivity of control shaft18is more influenced by the hydraulic quantity supplied to hydraulic actuator31than by the hydraulic pressure supplied thereto. That is, the hydraulic quantity required for operating hydraulic actuator31is determined in relation to the responsivity of control shaft18. As a result, by determining the hydraulic quantity required for operating hydraulic actuator31in transition of the compression ratio, the hydraulic pressure required for operating hydraulic actuator31can be determined based on characteristics of the hydraulic system including hydraulic actuator31. On the other hand, the second hydraulic pressure means a hydraulic pressure required for holding control shaft18against the rotation force applied thereto in the same direction as the rotation moment applied thereto due to the combustion pressure. In other words, the second hydraulic pressure means the hydraulic pressure required for holding control shaft18against the rotation force applied thereto upon varying the compression ratio from the larger side to the smaller side. Control shaft18undergoes the rotation moment or load caused by the combustion pressure in many operating ranges of the engine.

Owing to the determination of target hydraulic pressure Pt by selecting the larger one of the first and second hydraulic pressures, the hydraulic pressure immediately upstream of directional control valve59can be reduced to a lower limit without adversely affecting the responsivity of control shaft18upon transition of the compression ratio. This serves for reducing energy consumption. Especially, an energy required for driving oil pump60can be decreased by reducing the hydraulic pressure immediately upstream of directional control valve59. Further, an amount of the hydraulic fluid leaking from directional control valve59and hydraulic actuator31can be reduced, so that energy consumption required for replenishing the leakage amount of the hydraulic fluid can be suppressed.

FIG. 4illustrates characteristic of compression ratio to be controlled relative to engine operating conditions, namely, engine speed and engine torque (load). In a range of low engine torque, the compression ratio is controlled to higher in order to enhance thermal efficiency. In contrast, in a range of high engine torque, the compression ratio is controlled to lower in order to prevent occurrence of knocking. Basically, as the engine torque becomes lower, the compression ratio is controlled to higher.

FIG. 5illustrates characteristic of a maximum torque required for driving control shaft18, relative to engine speed and engine torque (load). As shown inFIG. 5, as the engine torque becomes lower, the required torque of control shaft18becomes larger. Meanwhile, since oil pump60is rotated synchronously with crankshaft3of the engine, the hydraulic pressure produced increases as the engine speed becomes higher.

Referring toFIG. 6, there is shown a second embodiment of the variable compression ratio system which differs in the hydraulic control from the first embodiment. Like reference numerals denote like parts, and therefore, detailed explanations therefor are omitted. Check valve71is disposed within supply passage61between oil pump60and directional control valve59. Hydraulic accumulator72is disposed between check valve71and directional control valve59and stores the hydraulic pressure discharged from oil pump60through check valve61. Pressure sensor67detects the hydraulic pressure between check valve71and directional control valve59, namely, the hydraulic pressure within hydraulic accumulator72. Relief passage65is branched from an upstream portion of supply passage61which is located between check valve71and oil pump60. Unloading valve73is disposed within relief passage65. Unloading valve73is electronically connected to controller40and operated to release the hydraulic pressure discharged from oil pump60when the hydraulic pressure within hydraulic accumulator72is not less than a predetermined hydraulic pressure. The hydraulic pressure released from unloading valve73is fed to oil pan68. With this arrangement, difference between the hydraulic pressure on the upstream side of oil pump60and the hydraulic pressure on the downstream side of oil pump60can be reduced so that energy consumption in driving oil pump60can be lowered.

Referring toFIG. 7, there is shown a flow of the hydraulic control operation implemented by controller40in the second embodiment ofFIG. 6. Logic flow starts and goes to block S1where actual operating conditions of the engine are read. In this embodiment, the operating conditions are engine speed Ne, intake air amount Qa and compression ratio ea determined based on the detected rotational angle of control shaft18. The logic flow goes to block S2where upper limit pressure P1and lower limit pressure P2of hydraulic accumulator72are determined based on the operating conditions read at block S1. Here, assuming that target hydraulic pressure Pt is indicated at P0, the relationship between target hydraulic pressure P0and upper and lower limit pressures P1and P2is expressed as follows: P0<P2<P1. The logic flow goes to block S3where hydraulic pressure Pn within hydraulic accumulator72which is detected by pressure sensor67is read, and then goes to block S4. At block S4, an interrogation is made whether or not unloading valve73is open to allow release of the hydraulic pressure discharged from oil pump60. If, at block S4, the interrogation is in negative, indicating that unloading valve73is closed to prevent release of the hydraulic pressure discharged from oil pump60, the logic flow goes to block S5. At block S5, an interrogation is made whether or not detected hydraulic pressure Pn within hydraulic accumulator72is more than upper limit pressure P1. If, at block S5, the interrogation is in affirmative, the logic flow goes to block S6where unloading valve73is opened. If, at block S5, the interrogation is in negative, the logic flow goes to end.

On the other hand, if, at block S4, the interrogation is in affirmative, indicating that unloading valve73is open, the logic flow goes to block S7. At block S7, an interrogation is made whether or not detected hydraulic pressure Pn within hydraulic accumulator72is less than lower limit pressure P2. If, at block S7, the interrogation is in affirmative, the logic flow goes to block S8where unloading valve73is closed. If, at block S7, the interrogation is in negative, the logic flow jumps to end. Thus, hydraulic pressure Pn within hydraulic accumulator72can be always maintained between upper limit pressure P1and lower limit pressure P2.

Next, referring toFIG. 8, there is shown a third embodiment of the variable compression ratio system which differs in that, instead of unloading valve73of the second embodiment, clutch mechanism81is provided for coupling oil pump60to the engine, from the second embodiment. Oil pump60is driven by engine crank pulley63through clutch mechanism81. Clutch mechanism81may be formed by an electromagnetic clutch assembly. Clutch mechanism81is electronically connected to controller40and operated to allow the coupling between oil pump60and the engine to thereby drive oil pump60and prevent the coupling therebetween to thereby stop oil pump60. With this arrangement, energy consumption in driving oil pump60can be reduced.

FIG. 9illustrates a flow of the hydraulic control operation implemented by controller40in the third embodiment ofFIG. 8. The flow differs in blocks S104to S108from the flow of the second embodiment. Similar to the second embodiment, there is the relationship P0<P2<P1between target hydraulic pressure P0and upper and lower limit pressures P1and P2determined at block S2. Subsequent to block S3, logic flow goes to block S104where an interrogation is made whether or not clutch mechanism81is applied to allow the coupling between oil pump60and the engine. If, at block S104, the interrogation is in affirmative, the logic flow goes to block S105. At block S105, an interrogation is made whether or not detected hydraulic pressure Pn within hydraulic accumulator72is more than upper limit pressure P1. If, at block S105, the interrogation is in affirmative, the logic flow goes to block S106where clutch mechanism81is released to prevent the coupling between oil pump60and the engine and thereby stop oil pump60. If, at block S105, the interrogation is negative, the logic flow goes to end.

On the other hand, if, at block S104, the interrogation is in negative, indicating that clutch mechanism81is released, the logic flow goes to block S107. At block S107, an interrogation is made whether or not detected hydraulic pressure Pn within hydraulic accumulator72is less than lower limit pressure P2. If, at block S107, the interrogation is in affirmative, the logic flow goes to block S108where clutch mechanism81is applied to allow the coupling between oil pump60and the engine and thereby restart oil pump60. If, at block S107, the interrogation is in negative, the logic flow goes to end. Thus, hydraulic pressure Pn within hydraulic accumulator72can be always maintained between upper limit pressure P1and lower limit pressure P2.

Referring toFIGS. 10 and 11, a modification of the third embodiment of the variable compression ratio system is explained.FIG. 10illustrates characteristic of compression ratio to be controlled with respect to engine operating conditions, namely, engine speed and engine torque (load), which is used in the modification. In the modification, the compression ratio is controlled to a minimum at a predetermined high speed of the engine. The predetermined high speed may be 4000 rpm and be in a range from 3600 rpm to 4000 rpm. Variable compression ratio mechanism10may be provided with a stop which is arranged to stop control shaft18in a rotational position where the compression ratio is the minimum. In such a case, it will eliminate the hydraulic pressure which is required for holding control shaft18in the rotational position at the predetermined high speed of the engine. This is because the rotation moment applied to control shaft18due to the combustion pressure acts to rotate control shaft18in such a direction as to vary the compression ratio from the larger side to the smaller side, as explained above. Controller40is programmed to control the hydraulic pressure supplied to hydraulic actuator31so as to minimize the compression ratio and operate clutch mechanism81to prevent the coupling between oil pump60and the engine, when the engine is operated at the predetermined high speed.

FIG. 11illustrates a flow of the hydraulic control implemented by controller40in the modification of the third embodiment. The flow differs in blocks S201and S210from the flow of the third embodiment. Subsequent to block S1, logic flow goes to block S201where an interrogation is made whether or not detected engine speed Ne exceeds predetermined high speed N1. If, at block S201, the interrogation is in affirmative, the logic flow goes to block S210. At block S210, clutch mechanism81is released to prevent the coupling between oil pump60and the engine and stop oil pump60. The logic flow then goes to end. If, at block S201, the interrogation is in negative, the logic flow goes to block S2.

In the modification, a maximum speed of oil pump60can be set at a lower value. This serves for reducing the size and weight of oil pump60.

As explained in the embodiments and modification of the present invention, the hydraulic actuator is operated by the oil pump mechanically driven by the internal combustion engine. This can serve for increasing efficiency in using the engine output. Further, the hydraulic pressure supplied to the hydraulic actuator can be variably controlled to an adequate hydraulic pressure depending on the engine operating conditions. This can serve for suppressing energy consumption in driving the hydraulic actuator.

This application is based on a prior Japanese Patent Application No. 2002-320758 filed on Nov. 5, 2002. The entire contents of the Japanese Patent Application No. 2002-320758 is hereby incorporated by reference.