Patent Application
20120037105
(1) Field of the Invention
The present invention relates to a variably operated valve apparatus for an internal combustion engine, a start system therefor, and a start control apparatus therefor which are capable of controlling at least a relationship between a closure timing of an intake valve and a stop position of a piston to enable an improvement in an engine startability.
(2) Description of Related Art
A previously proposed variably operated valve apparatus is exemplified by a Japanese Patent Application First Publication No. 2006-125276 published on May 18, 2006 (which corresponds to a Japanese Patent No. 4419800 issued on Dec. 11, 2009).
In the previously proposed variably operated valve apparatus, after an engine start condition is established, a retardation angle control for a closure timing of an intake valve(s) is performed through a valve timing control apparatus in order for a compression in a top dead center of one of engine cylinders which is in a piston compression stroke to be reduced (a decompression). Thus, a rise in rotation of cranking is made faster and a favorable self-start can be achieved.
However, in the previously proposed variably operated valve apparatus described hereinabove, a stop position of a piston (or crankshaft) of one of a plurality of cylinders of the engine which is in the compression stroke at the time of the stop of the engine is frequently placed at a substantially intermediate position between the top dead center and a bottom dead center due to a push down of the piston according to the compression within the corresponding cylinder. During a time duration to the subsequent engine restart, an atmospheric pressure is invaded into the inside of the cylinder through a gap between the piston and a cylinder bore. Therefore, if the cranking is started at the time of the restart, the piston in the compression stroke is raised from the crank position at the time of the stop of the engine so that the compression becomes large.
The crank position (piston position) at the time of the stop of the engine becomes easy to be dispersed. If the crank position is earlier than the closure timing of the intake valve, the compression is increased along with the thereafter rise in the piston with the closure timing of the intake valve as the atmospheric pressure. If the crank position is later than the closure timing of the intake valve, the compression is increased along with the thereafter rise in the piston with the closure timing of the intake valve as the atmospheric pressure.
In other words, a stop position of the crankshaft is dispersed and there are cases in which the compression is determined according to the closure timing of the intake valve and is determined according to the stop position of the crankshaft. Hence, both cases occur. Especially, in the latter case, there is a high possibility that a start cranking characteristic is dispersed due to the dispersion of the crank(shaft) stop position so that a stable engine startability cannot be obtained
It is, therefore, an object of the present invention to provide a variably operated valve apparatus for an internal combustion engine, a start system therefor, and a start control apparatus therefor which can achieve a stabilization of a top dead center compression by means of a piston, thus the start cranking characteristic being stabilized and an engine startability being improved.
According to one aspect of the present invention, there is provided a variably operated valve apparatus for use in an internal combustion engine, the internal combustion engine being configured for a stop position of a piston to be controlled at a position passed through a bottom dead center at a time of a stop of the engine, the variably operated valve apparatus comprising: a variable valve actuator configured to variably adjust at least a closure timing of an intake valve; and a mechanically stabilizing section configured to mechanically stabilize the closure timing of the intake valve at a position more retardation angle side than a most retardation angle position which is a limit of a control range of the stop position of the piston at the time of the stop of the engine.
According to another aspect of the present invention, there is provided a start system for an internal combustion engine, comprising: a variably valve operated apparatus configured to variably control at least a closure timing of an intake valve and, at a time of a stop of the engine, to stop the closure timing of the intake valve at a constant position more retardation angle side than a bottom dead center; and crank position control means for controlling a stop position of a piston at a position more advance angle side than the closure timing of the intake valve, at a time of the stop of the engine.
According to a still another aspect of the present invention, there is provided a start control apparatus for an internal combustion engine, comprising: a variably operated valve apparatus configured to variably control at least a closure timing of an intake valve; and crank position control means for controlling stop positions of a crankshaft and a piston at a time of a stop of the engine, wherein, at the time of the stop of the engine, the variably operated valve apparatus controls the closure timing of the intake valve at a constant position more retardation angle side than a bottom dead center and the crank position control means controls the stop position of the piston at a position more advance angle side than the closure timing of the intake valve.
Reference will hereinafter be made to the drawings in order to facilitate a better understanding of the present invention. That is to say, variably operated valve apparatus in preferred embodiments according to the present invention will, hereinafter, be described in details with reference to the accompanied drawings. In a first preferred embodiment of the variably operated valve apparatus according to the present invention, the present invention is applicable to a gasoline specification four-cycle, four-cylinder internal combustion engine.
First, a structure of the whole internal combustion engine to which the variably operated valve apparatus according to the present invention is applicable will be described with reference to
Piston 01 is linked with a crankshaft 02 via a connecting rod 03 and a combustion chamber 04 is formed between a space of a crown surface of piston 01 and a lower surface of cylinder head SH.
A butterfly type throttle valve SV is installed within an inside of intake manifold Ia of an intake air tube I located at upstream side of intake manifold Ia connected to intake port IP for controlling an intake air quantity of the engine and a fuel injection valve (not shown) is installed within the inside of the intake manifold Ia located at the downstream side thereof through which fuel is injected toward intake port IP. In addition, an ignition plug 05 is installed at an approximately center of cylinder head SH.
Crankshaft 02 is rotationally driven by means of an electrically driven drive motor 07 via a pinion gear mechanism 06 during the engine start (at a time of a start of the engine). This drive motor 07 constitutes crank position control means for controlling a crank angle (a crank position) at a time of engine stop and a slide position of piston 01 as well as for performing the cranking (the engine start).
It should be noted that a valve lift, a working angle, and a lift phase (an open-and-closure timing) of intake valves 4, 4 are variably controlled by means of the variably operated valve apparatus.
That is to say, as appreciated from
Intake VEL 1 has a structure in the same way as described in a Japanese Patent Application First Publication No. 2003-172112 filed by the same Applicant (assignee) in Japan on Dec. 4, 2001 and published on Jun. 20, 2003. A brief explanation thereof will be made. That is to say, intake VEL 1 includes: a hollow drive axle 6 rotatably supported on a bearing located at an upper part of cylinder head SH; a drive cam 7 which is an eccentric rotation cam which is fixedly installed on drive axle 6 by means of a press fit thereof into drive axle 6; two swing cams 9, 9 causing open operations of respective intake valves 4, 4 by slidably contacting on upper surfaces of valve lifters 8, 8 disposed on upper end portions of respective intake valves 4, 4; and a transmission mechanism linked between drive cam 7 and swing cams 9, 9 for transmitting a rotational force of drive cam 7 as a swing force of swings cams 9, 9.
The rotating force is transmitted from crankshaft 02 to drive axle 6 via a timing sprocket 30 and a timing chain installed on one end of drive axle 6. A rotational direction of drive axle 6 is set in a clockwise direction as shown in
Drive cam 7 is substantially of a ring shape, is fixed on drive axle 6 via a drive axle inserting hole formed along an internal axis direction and an axial core of a cam main body is offset by a predetermined quantity in a radial direction from the axial core of drive axle 6.
Each of both of swing cams 9, 9, as shown in
The transmission mechanism includes: a rocker arm 11 disposed on an upper side of drive axle 6; a link arm 12 linking one end portion 11a of rocker arm 11 and drive cam 7; and a link rod 13 linking the other end portion 11b of rocker arm 11 and each of swing cams 9, 9.
A center positioned cylindrical base portion of rocker arm 11 is rotatably supported on a control cam which will be described later via a supporting hole and one end 11a of rocker arm 11 is rotatably linked with link arm 12 by means of a pin 14. On the other hand, the other end portion 11b is rotatably linked with one end portion 13a of link rod 13 via a pin 15.
Link arm 12 is formed with a fitting hole into which a cam main body of drive cam 7 is rotatably fitted at a center position of a circular ring-shaped base portion 12a having a relatively large diameter. On the other hand, a projection end 12b is linked with one end portion 11a of rocker arm 11 by means of pin 14.
Link rod 13 has the other end portion 13b linked rotatably with the cam nose portion of each of swing cams 9, 9 via a pin 16 (as shown in
In addition, a control axle 17 is rotatably supported by means of the same bearing on an upper position of drive axle 6 and a control cam 18 is fixed on an outer periphery of control axle 17, with control cam 18 rotatably fitted into a supporting hole of rocker arm 11. Control cam 18 serves as a swing fulcrum of rocker arm 11.
Control axle 17 is disposed in a forward-and-rearward direction of the engine in parallel to drive axle 6 and is rotatably controlled by means of a drive mechanism 19. On the other hand, control cam 18 is of a cylindrical shape. The axial center position thereof is deflected by a predetermined distance from the axial center of control axle 17.
Drive mechanism 19 includes, as shown in
Electrically driven motor 20 is constituted by a proportional type DC motor and is driven by means of a control signal from an electronic controller 22 (ECU) for detecting an engine driving condition. It should be noted that an intake drive axle rotational angle signal from an intake drive axle rotation sensor 28 and a control axle rotational angle signal from a control axle rotational angle sensor 29 are inputted to electronic controller 22.
Electronic controller 22 detects the present engine driving condition according to detection signals from various kinds of sensors such as a crank angle sensor detecting an engine revolution number, an airflow meter detecting an intake air quantity, an engine coolant temperature sensor, an intake air temperature sensor, a knocking sensor, a vehicle speed sensor, and an accelerator opening angle sensor and outputs control signals to throttle valve SV, fuel injection valve, electrically driven motor 20, and so forth.
Ball screw transmission means 21 mainly includes: a ball-screw axle 23 disposed substantially coaxially with the drive shaft of electrically driven motor 20; a ball nut 24 which is a movable member spirally meshed with an outer periphery of ball screw axle 23; an linkage arm 25 linked on one end portion of control axle 17 along a diameter direction thereof; and a link member 26 for linking both linkage arm 25 and ball nut 24.
Ball screw axle 23 has a ball circulating groove 23a of a predetermined width formed spirally continuously over a whole outer peripheral surface thereof except both end portions. One end portion of ball screw axle 23 is coupled to drive shaft 20a of electrically driven motor 20. This coupling transmits the rotational drive force of electrically driven motor 20 to ball screw axle 23 and allows a slight movement of ball screw axle 23 in the axial direction.
Ball nut 24 is formed substantially cylindrically. A guide groove 24a for holding a plurality of balls 27 to enable a rolling motion thereof in conjunction with ball circulating groove 23a is spirally continuously formed on an inner peripheral surface of ball nut 24. A rotational motion of ball screw axle 23 is converted into a rectilinear motion of ball nut 24 via each of balls 27 so as to provide an axial movement force.
In addition, this ball nut 24 is biased toward an opposite side of electrically driven motor 20 by means of a spring force of a coil spring 31 so as to eliminate a backlash between ball nut 24 and ball screw axle 23 and this spring force serves to always (at all times) bias intake valves 4, 4 toward directions of a maximum lift and a maximum working angle via control axle 17.
Hence, after the engine has stopped, the spring force of coil spring 31 (positively) biases intake valves 4, 4 toward the maximum working angle side (most retardation angle side) without failure so as to maintain the engine in a stable state.
Next, an operation of intake VEL 1 will briefly be explained.
At a time of the engine stop, a power supply from electronic controller 22 to electrically driven motor 20 is interrupted. At this time, ball nut 24 is linearly moved to a maximum rightward (a rightmost) direction (a direction in which ball nut 24 becomes spaced apart from electrically driven motor 20) according to the spring force of coil spring 31. Thus, control axle 17 is rotated in one direction via link member 39 and linkage arm 25. That is to say, in a case where a conversion power is not acted upon by means of electrically driven motor 20, the lift and working angle characteristics of intake valves 4, 4 are stably held mechanically to the maximum working angle. This maximum working angle provides a default position. It should be noted that left and right maximum rotation positions of control axle 17 are limited by means of rotation limiting stoppers ST. It should also be noted that one of two rotation limiting stoppers ST for limiting the rightward direction of control axle 17 is shown in
Hence, an axial center of control cam 18 is rotated about an axial center of control axle 17 by the same radius, as shown in
Thus, drive cam 7 is rotated to push up one end portion 11a of rocker arm 11 via link arm 12. Thus, the lift (quantity) thereof is transmitted to each of swing cams 9, 9 and each valve lifter 8 via link rod 13. Thus, intake valves 4, 4 have valve lifts of a large lift (L3), as shown in a characteristic graph of
Next, at a time of a start of engine, an ignition switch is turned to ON to rotationally activate drive motor 07 so that a cranking rotation of crankshaft 02 is started. However, at an initial stage of the cranking, the biasing force of coil spring 31 causes the valve lift to be maintained at maximum lift and working angle D3 to be maintained at the maximum working angle. In addition, closure timing (IC) of intake valves 4, 4 also indicates a more retardation angle side than a bottom dead center (BDC).
Then, the cranking rotation is raised to a predetermined rotation. At this time, the control signal from electronic controller 22 causes electrically driven motor 20 to be reversely rotated. This rotating torque is transmitted to ball screw axle 23 to be rotated. Along with this rotation, ball nut 24 is linearly moved from a position (not shown) to the direction approaching to electrically driven motor 20 against the spring force of coil spring 31 (near an intermediate position between the positions shown in
Thus, the axial center of control cam 18 is held at a rotational angle position near an intermediate position of
Then, when the engine driving condition is transferred to an ordinary driving after an end of the start, electronic controller 22 controls the valve lift, for example, from a small lift (L1) to a middle lift (L2) and controls the working angle from a small working angle (D1) to a middle working angle (D2). In addition, the lift phase is, for example, controlled to be oriented toward an advance angle side. Thus, the closure timing of intake valves 4, 4 is made earlier, a valve overlap with exhaust valves 5, 5 becomes large (wide), and a pumping loss accordingly becomes reduced. Consequently, a fuel economy is accordingly improved.
Then, in a case where the vehicle driver depresses an accelerator pedal (not shown) so that the engine is transferred from the ordinary driving to a high-load-and-high-rotation drive region, the control signal from electronic controller 22 rotates electrically driven motor 20 toward one direction and control axle 17 causes control cam 18 to be rotated in an anticlockwise direction and its axial center of control cam 18 is pivoted in the lower direction, as shown in
Thus, when drive cam 7 is rotated and one end portion 11a of rocker arm 11 is pushed upward via link arm 12, the valve lift is transmitted to each swing cam 9, 9 and each valve lifter 8, 8 via link rod 13. The valve lift becomes continuously large to L3 shown in
That is to say, the lift of intake valves 4, 4 is continuously varied from small lift of L1 to the maximum lift of L3 in accordance with the driving condition of the engine. Hence, the working angle of each intake valve 4, 4 is continuously varied from small lift D1 to maximum lift D3.
It should be noted that intake VTC 2 is of, so-called, a vane type and has substantially the same structure as disclosed in a Japanese Patent Application First Publication No. 2007-198367 published on Aug. 9, 2007 (which corresponds to a U.S. Pat. No. 7,703,424 issued on Apr. 27, 2010, the disclosure of which is herein incorporated by reference). Intake VTC 2 will, hereinafter, briefly be explained on a basis of
That is to say, intake VTC 2 includes: a timing sprocket 30 for transmitting a rotational force thereof to drive axle 6; a vane member 32 fixed on a terminal portion of drive axle 6 and rotatably housed within timing sprocket 30; and a hydraulic pressure circuit 33 for rotating vane member 32 in a normal direction or in a reverse direction according to the hydraulic pressure.
Timing sprocket 30 includes: a housing 34 in which vane member 32 is rotatably housed: a front cover 35 (refer to
Housing 34 is of a cylindrical shape and its front and rear ends thereof are opened. Four shoes 34a are projected toward an inner direction thereof which are partitioning walls thereof and are installed at about 90° positions of a peripheral direction of an inner peripheral surface thereof.
Each shoe 34a is of a substantially trapezoid shape over a laterally cross sectioned surface and, at its substantial center position of each shoe 34a, a bolt penetrating hole 34b through which an axle portion of each bolt 37 is penetrated is penetrated through each shoe 34a in an axial direction of housing 34 (totally four bolt penetrating holes). A letter C shaped seal member 38 and a plate spring (not shown) which presses seal member 38 toward the inner direction are fitted into and held by a holding groove cut out along an axial direction of each inner end surface of shoe 34a.
Front cover 35 is formed in a disc plate shape and a supporting hole 35a having a relatively large diameter is fitted into the center of the front cover, and four bolt holes (not shown) are fitted at positions of the outer peripheral portion of front cover 35 corresponding to respective bolt penetrating holes 34b of housing 34.
Rear cover 36 has its rear end side installed integrally with a gear portion 36a with which the timing chain is meshed and a bearing hole 36b having a large diameter is penetrated axially at a substantially center position of rear cover 36.
Vane member 32 includes a vane rotor 32a in a circular ring shape having a bolt penetrating hole at the center thereof and four vanes 32b integrally installed at substantially 90° positions in a peripheral direction of the outer peripheral surface of vane rotor 32a.
Vane rotor 32a has its small diameter cylindrical portion at the front end side thereof rotatably supported on supporting hole 35a of front cover 35 and has its small diameter cylindrical portion at the rear end side thereof rotatably supported on a bearing hole 36b of rear cover 36.
In addition, vane member 32 is fixed onto the front end portion of drive axle 6 through the axial direction thereof by means of a fixture bolt 39 penetrated through a bolt penetrating hole formed on vane rotor 32a.
Each of three of respective vanes 32b is formed in a relatively elongated rectangular shape and the remaining one of vanes 32b is formed in a relatively large trapezoid shape. Each width of three vanes 32b in the rectangular shape is set to be the mutually same and the remaining one in the trapezoid shape has its width set to be larger than the three of vanes 32b in the rectangular shape. A weight balance of whole vane member 32 is taken.
In addition, each vane 32b is disposed between each shoe 34a, a letter C shaped seal member 40 is fitted into an elongated holding groove formed in an axial direction of each outer surface of vanes 32b and is slidably contacted on an inner peripheral surface of housing 34, and a plate spring which presses seal member 40 in the inner peripheral surface direction is fitted into the elongated holding groove. In addition, substantially circular-shaped two recessed grooves 32c are respectively formed on one side surface of each vane 32b in the rotation direction of timing sprocket 30.
A pair of coil springs 55, 56 are respectively disposed between recessed groove 32c of each vane 32b and an opposing surface 10b of each shoe 34a and are biasing means for rotationally biasing vane member 32 toward the retardation angle side. In other words, at the time of the stop of the engine, in a case where the conversion power is not acted upon by the hydraulic pressure with no supply of the hydraulic pressure from hydraulic pressure circuit 33, vane member 32 is stably and mechanically biased at the most retardation angle position.
These two coil springs 55, 56 are formed independently of each other and juxtaposed to each other. An axial length (coil length) of each of two coil springs 55, 56 is set to be longer than the length between one side surface of vane 32b and an opposing surface of shoe 34a. Then, both lengths of respective coil springs 55, 56 are set to be the same length.
Each coil spring 55, 56 is juxtaposed with an inter-axle distance at which each coil spring 55, 56 is mutually not contacted on each other during a maximum compression deformation and each end portion thereof is linked via a thin-plate like retainer (not shown) fitted into recessed groove 32c of each shoe 34a.
In addition, four advance angle chambers 41 and four retardation angle chambers 42 are partitioned respectively between both sides of respective vanes 32b and respective side surfaces of shoes 34a, respectively.
Hydraulic pressure circuits 33, as shown in
First and second hydraulic pressure passages 43, 44 are formed in an inside of a column shaped passage constituting section 39. One end portion of this passage constituting section 39 is inserted within a holding hole 32 from the small-diameter cylindrical portion of vane rotor 32a and the other end portion thereof is connected to electromagnetic switching valve 47.
In addition, three annular seal members 60 are fitted into a space between an outer peripheral surface of one end portion of passage constituting section 39 and an inner peripheral surface of holding hole 32d for partitioning and sealing between one end sides of respective first and second hydraulic pressure passages 43, 44.
First hydraulic pressure passage 43 includes an oil chamber 43a formed at a terminal portion of holding hole 32d at drive axle 6 side and four branch passages 43b formed substantially radially on an inside of vane rotor 32a for communicating oil chamber 43a and respective advance angle chambers 41.
On the other hand, second hydraulic pressure passage 44 includes an annular chamber 44a which is ended within an end portion of passage constituting section 39 and formed on the outer peripheral surface of the one end portion of passage constituting section 39 and a second oil passage 44b formed in a substantially letter L shape in the inside of vane rotor 32 for communicating annular chamber 44a and respective retardation angle chambers 42.
Electromagnetic switching valve 47 is of a four-port three-position type. An inner valve body is switching controlled to enable switching relatively between respective hydraulic pressure passages 43, 44 and supply passage 45 and drain passage 46 and switched in response to the control signal from electronic controller 22.
This electronic controller 22 is common to intake VEL 1, detects the engine driving condition and detects the relative rotational position between timing sprocket 30 and drive axle 6 according to the signals from crank angle sensor and intake drive axle angle sensor.
Then, the switching operation of electromagnetic switching valve 47 supplies the working oil to retardation angle chambers 42 at the time of the engine start and, thereafter, supplies the working oil to advance angle chambers 41.
It should be noted that a lock mechanism is interposed between vane member 32 and housing 34 for constraining a rotation of vane member 32 and releasing the constraint thereof.
That is to say, this lock mechanism, as typically shown in FIG. 9,includes: a sliding hole 50 installed between the remaining one of vanes 32b whose width is larger (wider) and rear cover 36 and formed along the axial direction of drive axle 6 within the inside of vane 32b; a lid provided cylindrical lock pin 51 slidably installed within the inside of sliding hole 50; an engagement hole 52a installed on an engagement hole constituting section 52 of a laterally cross sectioned cup shape and fixed within a fixture hole of sliding hole 50 and through which a tapered tip end 51a of lock pin 51 is disengageably engaged; and a spring member 54 retained on a spring retainer 53 fixed on a bottom surface side of sliding hole 50 for biasing lock pin 51 in a direction of engagement hole 52a.
In addition, the hydraulic pressure within retardation angle side chamber 42 or the hydraulic pressure of the oil pump are supplied to engagement hole 52a via an oil hole (not shown).
Lock pin 51 locks a relative rotation between timing sprocket 30 and drive axle 6 when tip end portion 51a of lock pin 51 is engaged with engagement hole 52a by means of the spring force of spring member 54 at a position at which vane member 32 is rotated at the most retardation angle side. In addition, lock pin 51 is retracted by means of the hydraulic pressure supplied from retardation angle chambers 42 within engagement hole 52a and the hydraulic pressure of the oil pump so that the engagement of lock pin 51 with engagement hole 52a is released.
It should be noted that
It should also be noted that an operation of intake VTC 2 is the same as described in a Japanese Patent Application First Publication No. 2007-198367 (which corresponds to U.S. Pat. No. 7,703,424 (, the disclosure of which is herein incorporated by reference)).
Exhaust VTC 3 has exactly the same structure as intake VTC 2 described above. In a case where the hydraulic pressure of hydraulic pressure circuit 33 is not acted upon at the time of the stop of engine, the spring forces of coil springs 55, 56 stably blase vane member 32 at the pivotal position of the most retardation angle side. This exhaust VTC 3 inputs an exhaust valve lift phase control signal which is open-and-closure timings of exhaust valves 5, 5 from electronic controller 22 on a basis of information signal from an exhaust drive axle angle sensor to control valve timing of exhaust valves 5, 5.
Hereinafter, an action of the first embodiment will be described on a basis of
The crank angle during the start of the engine will be considered as follows:
Suppose that #1 cylinder (first cylinder) is the cylinder in the compression stroke. At this time, there are many cases in which a rotational phase (crank angle) of a crankshaft 02 of #1 cylinder, namely, a position of piston 01 is such that piston 01 is stopped at an intermediate position between top dead center TDC and bottom dead center BDC. This is because, if piston 01 approaches to top dead center (TDC), a compressive pressure acted upon piston 01 causes piston 01 to be returned toward bottom dead center (BDC). At this time, crankshaft 02 is rotated in the counterclockwise direction as viewed from
Then, after the engine is stopped, the atmosphere is immediately invaded into combustion chamber 04 through a piston ring of piston 01 and an inner wall surface of the cylinder so that an inside of the cylinder indicates an atmospheric pressure.
Next, a case of the engine start will be considered as follows:
That is to say, when crankshaft 02 is rotated in the clockwise direction from the state of the atmosphere from the atmospheric pressure state of the inside of the cylinder to raise piston 01, the compression becomes peak at the top dead center (TDC). This top dead center compression becomes larger as an initial position of piston 01 becomes lower and becomes smaller as the initial position thereof becomes higher. In other words, this top dead center compression receives an influence of a stop crank angle (a stop piston position) of first cylinder (#1 cylinder).
It should be noted that a dispersion of the stop crank angle (the stop piston position) is essentially large since influences of a minute pressure balance and a friction balance at the time of engine stop are received. Even if the stop crank position is positively controlled using drive motor 07 which is the crank position control means, the dispersion is left to some degree and, thus, the dispersion of this top dead center compression is essentially left.
Therefore, the cranking becomes unstable due to the dispersion in the first time top dead center compression at the time of cranking of the engine or the dispersion becomes large and, consequently, brings out an unstablization of the engine startability, in the case of the comparative example.
On the contrary, in the first embodiment, the top dead center compression is determined according to the closure timing of intake valves 4, 4, namely, default IC without influence of the unstable stop crank position. Thus, the stabilization of engine startability can be improved.
That is to say, as shown in
It should be noted that, if a target (position) of stop crank angle Z is Z0 and the dispersion thereof is set as ±α, a position of Z0+α which is the most retardation angle becomes a position which is advanced than default IC (Z0+α<default IC).
Hence, if a control range is set as Z0−α˜Z0+α, the stop crank angle (position) can be set to always be open at the stop crank position even if the stop crank position becomes dispersed.
Then, since the cranking is started to raise piston 01 but intake valves 4, 4 are opened, an inner cylinder pressure is not immediately raised but the atmospheric pressure is maintained therein. The compression is generated from a time point at which intake valves 4, 4 are closed at default IC. It should herein be noted that default IC is the mechanically stable position and the dispersion is almost not present thereat. Thus, the top dead center compression becomes stable. Consequently, the cranking becomes stabile and a favorable startability can be achieved.
In addition, an advancing of stop crank angle Z means that the position of piston 01 at another cylinder in the suction stroke (third cylinder (#3 cylinder)) is raised as shown in
As a result of this, an atomization time of fuel can sufficiently be taken to increase a degree of homogeneity of air mixture fuel. Hence, a combustibility (a complete ignition performance) at the time of the engine start becomes favorable and the favorable startability can be obtained. On the other hand, the closure timing of intake valves 4, 4 is retardation angle controlled. Thus, the top dead center compression is sufficiently lowered, an excessive load torque increase can be suppressed at the time of the initial stage of the start, and the development of a start vibration can be suppressed.
It should be noted that
It should be noted that an ignition order is such that first cylinder (#1 cylinder), third cylinder (#3 cylinder), fourth cylinder (#4 cylinder), and second cylinder (#2 cylinder) and this order is repeated. If the compression stroke is carried out in first cylinder (#1 cylinder) (the first cylinder is in the compression stroke), stop crank angle Z is advanced with respect to retarded default IC (IC3) and the top dead center compression becomes stable in the operation mechanism as described above. On the other hand, #3 cylinder (third cylinder) in the suction stroke can assure a sufficient suction stroke of piston 01 from stop crank angle Z in the same way as described above since stop crank angle Z is advanced to a proximity to default IO (IO3) which is the open timing of intake valves 4, 4. Hence, since the atomization time of fuel can sufficiently be taken, the degree of homogeneity of air mixture fuel can be enhanced. Consequently, a start combustibility (a complete explosion performance) can become favorable and the startability can furthermore be improved.
On the other hand, since the closure timing (IC3) of intake valves 4, 4 is retarded maximally, the top dead center compression is sufficiently lowered and the increase in an excessive load torque at the time of initial stage of start of the engine can be suppressed and the start vibration can be suppressed.
In the next #4 cylinder (fourth cylinder), a sufficient suction stroke can be taken from open timing (IO3) of intake valves 4, 4. In addition, in the same way, closure timing (IC3) of intake valves 4, 4 is retarded maximally. Hence, the same vibration suppression effect as in the case of third cylinder (#3 cylinder) can be obtained.
Furthermore, in the next second cylinder (#2 cylinder), the working angle of intake valves 4, 4 is converted into intermediate working angle D2 by means of intake VEL 1 and the lift phase thereof is converted into the intermediate phase by means of intake VTC 2 until the second cylinder enters the suction stroke so that the closure timing of intake valves 4, 4 approaches to the bottom dead center (BDC) (IC2). Since #4 cylinder (fourth cylinder) is in the third time combustion and the rotation of engine becomes slightly high, the start vibration is not easy to occur. Thus, the closure timing (IC) of intake valves 4, 4 is approached to the bottom dead center to increase an intake charging efficiency and to increase a combustion torque. Thereby, a further rise in the rotation can be promoted.
Furthermore, in the next first cylinder (#1 cylinder), the closure timing (IC) of intake valves 4, 4 is further approached (IC1) to the bottom dead center by converting the working angle into intermediate working angle D2 by means of intake VEL 1 and the lift phase into the intermediate phase by means of intake VTC 2 until the suction stroke. Thus, the intake charging efficiency can furthermore be increased, the rise in the rotation can be accelerated, and a quick startability can be achieved.
At a first step S1 of
At the next step S3, electronic controller 22 detects actual positions of the closure timing and the open timing of intake valves 4, 4 set by means of intake VEL 1 and intake VTC 2.
At the next step S4, electronic controller 22 determines whether actual closure timing (IC) by means of intake VEL 1 and intake VTC 2 is sufficiently near to target closure timing of IC3. That is to say, electronic controller 22 determines whether an absolute value of actual closure timing difference of actual IC−IC3 is smaller (narrower) than a predetermined minute angle L IC. If the absolute value is smaller than ΔIC (Yes) at step S4, the routine goes to a step S5.
At step S5, electronic controller 22 performs a stop position control of crank angle Z through drive motor 07 (Z0±α) so as to control the closure timing toward more advance angle side of Z0 than closure timing IC3 even with the consideration of control dispersion±taken into consideration. In this state, the engine is stopped at a step S6.
Hence, as shown in
In addition, if electronic controller 22 determines that the absolute value of actual closure timing difference |actual IC−IC3| is not smaller than predetermined minute angle ΔIC at step S4 (No), the routine goes to a step S7. In details, electronic controller 22 recognizes that a, so-called, operation slow-down phenomenon is developed in intake VEL 1 and/or intake VTC 2 due to a failure therein, namely, determines that the decompression according to closure timing IC cannot be made and the routine goes to step S7. At step S7, electronic controller 22 modifies the crank angle stop position control with the crank stop position as target Z0 to the crank stop position control such that the crank stop position is set in such a way that the target crank stop position is Z1 which is more retardation angle side than Z0. Hence, in this case, the first time top dead center compression is reduced according to the stop crank angle not according to the closure timing IC and, thus, the startability which is a minimum requirement is secured. In addition, a release of the crank angle stop position control itself is made so that the crank stop position is shifted toward the retardation angle side although the dispersion of the stop crank angle becomes large. Thus, in the same way, the startability which is the to minimum requirement is secured with the first time top dead center compression reduced.
Next, the flowchart of
At step S13, electronic controller 22 outputs the conversion signal to convert the closure timing to closure timing IC3 even after the start of cranking in the same way as described at step S2. Thus, the closure timing can furthermore be coincident with closure timing IC3.
Then, the routine goes to a step S14. At step S14, electronic controller 22 determines whether the crank angle has reached to a predetermined crank angle (Za) in response to the cranking. If not reached to the predetermined crank angle (Za) (No), the routine returns to step S13. If the crank angle has reached to the predetermined crank angle (Za) (Yes), the routine goes to a step S15.
At step S15, electronic controller 22 outputs the signals to convert the working angle of intake valves 4, 4 to the intermediate working angle and to convert the lift phase thereof to the intermediate phase, respectively, through intake VEL 1 and intake VTC 2. Thus, in the second cylinder (#2 cylinder), which enters the suction stroke after the third cylinder (#3 cylinder) and the fourth cylinder (#4 cylinder) which have been in the suction stroke, the closure timing of intake valves 4, 4 is controlled to be the closure timing IC2 (intermediate phase and intermediate working angle). At this time point, since the second cylinder (#2 cylinder) is at the third time combustion (suction stroke), the influence of the vibration is reduced due to a slight increase in the cranking number of revolutions. Hence, the torque is increased so that the cranking rotation can furthermore be increased by improving the intake (air) charging efficiency with the closure timing (IC) approached to the bottom dead center.
Furthermore, at a step S16, electronic controller 22 determines whether the crank angle has reached to a predetermined crank angle (Zb). If not reached to the predetermined crank angle (No), the routine returns to step S15. If reached to the predetermined crank angle (Yes), the routine goes to a step S17. At step S17, electronic controller 22 outputs the control signals to convert the working angle of intake valves 4, 4 to the small working angle and the lift phase to the phase at the more advance angle side through intake VEL 1 and intake VTC 2. Thus, starting from the third cylinder (#3 cylinder) and the first cylinder (#1 cylinder) which becomes suction stroke after the fourth cylinder (#4 cylinder) and the second cylinder (#2 cylinder) which have been suction stroke, the crank angle is controlled to the closure timing IC1 (small working angle and the advance angle phase). At this time point, a fourth time combustion is entered so that the crankshaft rotation speed is furthermore increased and this makes the influence of the vibration be further reduced. Therefore, closure timing IC is furthermore approached to top dead center (IC1) and the intake charging efficiency is furthermore increased so that the torque can be increased and the cranking rotation can furthermore be increased. Thereby, favorable and quick engine startability can be achieved.
At the next step S18, electronic controller 22 determines whether an engine coolant temperature has reached to a predetermined coolant temperature. If not reached to the predetermined coolant temperature (No), the routine returns to step S17. If the engine temperature has reached to the predetermined coolant temperature (Yes), namely, if electronic controller 22 determines that an engine warm-up is ended, the routine goes to a step S19. At step S19, electronic controller 22 transfers to an ordinary control based on an engine-speed-and-load map.
It should be noted that, in a case where actual IC is determined to be largely separated from default IC3 (No at step S11), the operation slow-down phenomenon is determined to be developed in intake VEL 1 and intake VTC 2 due to the failure therein. Then, the routines goes to steps S20 through S22. At this time, electronic controller 22 continues to output the signal to be converted to the default IC3 and transfers a control mode into an engine fail-safe control mode. That is to say, at step S20, electronic controller 22 outputs the signal to start the cranking. At step S21, the conversion signal to the default position (IC3) is outputted to intake VEL 1 and intake VTC 2 and, at step S22, electronic controller 22 performs the fail-safe control such that a minimum drivability is secured on a basis of actual IC.
The crank position in the first cylinder (#1 cylinder) which is in the compression stroke shown in
It should herein be noted that the cylinder in the expansion stroke is #2 cylinder (second cylinder) and piston 01 in #1 cylinder (first cylinder) is advanced toward the bottom dead center side. Thus, piston 01 of #2 cylinder (second cylinder) is advanced toward the top dead center side. Hence, since an elongation of the piston stroke in the expansion stroke can be taken and a lot of expansion works according to the combustion can be carried out. Hence, an effective push-down of piston 01 permits the increase in the startability according to the unaided combustion.
In addition, a default timing EO1 of the open timing (EO) of exhaust valves 5, 5 provides the most retardation angle and is placed at a position near to the bottom dead center (BDC). Consequently, a looseness in combustion pressure due to the open of exhaust valves 5, 5 during the combustion can be suppressed, the expansion work can furthermore be increased, and it becomes possible to effectively push down piston 01.
On the other hand, the position of piston 01 of first cylinder (#1 cylinder) is the same as a case of
Hence, a torque generated by the unaided combustion easily achieves a pass over the compression top dead center of first cylinder (#1 cylinder). As viewed from this point, an unaided engine startability can be increased.
Furthermore, since the compression of the first cylinder (#1 cylinder) can stably be reduced, this drive torque can easily get over the compression top dead center (the maximum compression) of first cylinder (#1 cylinder). That is to say, a positive rotation according to the unaided combustion can assuredly be obtained without the return of the piston toward a position before the top dead center.
Furthermore, the first cylinder (#1 cylinder) which is to be combusted subsequently to the second cylinder (#2 cylinder) will be considered as follows: At the initial stage of the compression stroke, stop crank position Z is advanced and the closure timing (IC3) of intake valves 4, 4 is retarded. Air within the cylinder is partially exhausted during the crank angle interval between stop crank angle (position) Z and closure timing IC3 to arrive at closure timing IC3. Thereafter, piston 01 is raised toward the top dead center and, in the midway through the top dead center, fuel is injected and ignited. Since a certain degree of compression is generated, the combustion torque equal to or larger than the unaided combustion of #2 cylinder (second cylinder) is obtained so that the rise in the rotation is promoted.
Next, the third cylinder (#3 cylinder) which is to be combusted subsequent to the first cylinder (#1 cylinder) will be considered as follows:
Since crank angle Z is placed at the advanced position at the initial stage of suction stroke, a relatively long suction stroke of piston 01 can be taken. Thus, the atomization of in-cylinder direct injection fuel is promoted so that the combustion becomes stabilized. In addition, closure timing IC3 of intake valves 4, 4 is retarded maximally. Thus, an abrupt increase in intake (air) charging efficiency in the intake air is suppressed and such a vibration which provides a major problem in an extremely low engine rotation region can be suppressed.
In the combustion of the next #4 cylinder (fourth cylinder), the injected fuel can be sucked into the cylinder from the open timing IO3 of intake valves 4, 4. Thus, the suction stroke can furthermore be elongated so that the atomization of fuel can be remarkably be improved and the vibration reduction effect by means of the retardation angle control of intake valve closure timing IC3 can be maintained.
In the combustion of the next #2 cylinder (second cylinder) combustion (the second time combustion), the working angle reduction control by means of intake VEL 1 and the advance angle control by means of intake VTC 2 are performed so that closure timing of intake valves 4, 4 is varied toward the bottom dead center side. Thus, the intake air charging efficiency (ηV) is improved and the torque is increased. Thus, the rise in the rotation of the engine is improved and the torque is accordingly increased. Thus, a quick startability can be obtained. In this state, the rotation is already to some degree raised. Thus, the start vibration falls in the region which does not provide the problem.
Subsequently, in the next #1 cylinder (first cylinder) combustion (the second time combustion), the working angle decrease control by means of intake valve VEL 1 and the advance angle control by means of intake VTC 2 are performed until the suction stroke so that the closure timing of intake valves 4, 4 is advanced in the proximity to the bottom dead center. Thus, the intake (air) charging efficiency (77V) is improved to increase the torque and the favorable startability due to a quick rotation can be promoted.
Subsequently, in the next third cylinder (#3 cylinder) combustion (the second time combustion), the closure timing of intake valves 4, 4 is, on the contrary, again retarded (IC3). This is because the torque is reduced to suppress an overshoot of rotation. Alternatively, it is possible to execute the decompression again by carrying out the same retardation angle control of closure timing IC at the rotation speed which is coincident with a resonance point of a vehicle drive system so as to make a countermeasure against the vibration.
In addition, using exhaust VTC 3, the lift phase of exhaust valves 5, 5 is advance angle controlled to the intermediate phase so that open timing EO of exhaust valves 5, 5 is advanced (EO1→EO2). This causes the combustion gas to be exhausted toward exhaust port side EP before the reduction of the catalytic temperature in the expansion stroke and the exhaust emission can be reduced. As the next step at which the quick start is achieved, the exhaust emission reduction is promoted. As described before, since the start time becomes short, the reduction effect of exhaust emission becomes large.
Then, after the engine coolant temperature is reached to the predetermined temperature, intake VEL 1, intake VTC 2, exhaust VTC 3 are controlled according to an engine-speed-and-engine-load map which accords with the requirement of driveability.
The control of electronic controller 22 in the case of second preferred embodiment will be described with reference to the control flowchart integrally shown in
At a first step S20, electronic controller 22 determines whether the engine start condition is satisfied. If not fall in the start condition (No), the process is ended. If the engine start condition is satisfied (Yes), the routine goes to a step S21.
At step S21, electronic controller 21 discriminates one of the cylinders which is presently in the expansion stroke from output signals of the crank angle sensor and the intake side drive axle angle sensor. Suppose that the expansion stroke cylinder is, for example, the second cylinder (#2 cylinder).
At a step S22, electronic controller 22 outputs the signal to convert the closure timing of intake valves 4, by means of intake VEL 1 and intake VTC 2 into closure timing IC3. Thus, the closure timing can become coincident with closure timing IC3.
At a step S23, the inner cylinder injection and the ignition are carried out in #2 (the second) cylinder so as to start the combustion therein. In details, the unaided combustion (the self or spontaneous combustion) is made.
This unaided combustion starts the rotation of the engine at a step S24 and, at a step S25, the fuel injection signal and the ignition signal are outputted in accordance with the ignition order.
At a step S26, electronic controller 22 determines whether crank angle Z has been rotated through a predetermined crank angle (Za). If not yet rotated (No), the routine returns to step S26 itself. If rotated (Yes), the routine goes to a step S27.
At step S27, electronic controller 22 outputs the signals to intake VEL 1 and intake VTC 2 respectively to provide the intermediate working angle for intake valves 4, 4 and to provide the intermediate phase for intake valves 4, 4 to convert the closure timing of intake valves 4, 4 into IC1 so as to be in time for the suction stroke of the next combustion (the second cycle) of the second cylinder (#2 cylinder).
At a step S28, electronic controller 22 determines whether crank angle Z has been rotated through predetermined crank angle (Zb) in the same way as step S26. If not yet rotated (No), the routine returns to step S27. If rotated (Yes), the routine transfers to a step S29.
At a step S29, electronic controller 22 outputs a small working angle control signal to intake VEL 1 and an advance phase control signal to intake VTC 2 respectively to convert the closure timing of intake valves 4, 4 to IC1 so as to be in time for the suction stroke of the next first cylinder (#1 cylinder) combustion (the second cycle).
At the next step S30, electronic controller 22 determines whether crank angle Z has been rotated through a predetermined angle (Zc), in the same way as step S28. If not yet rotated (No), the routine returns to step S29. If rotated (Yes), the routine goes to a step S31.
At step S31, electronic controller 22 outputs a maximum working angle control signal to intake VEL 1 and a most retardation angle phase control signal to intake VTC 2 respectively to convert the closure timing of intake valves to IC3 so as to be in time for the suction stroke of the next third cylinder (#3 cylinder) combustion (2 cycle).
At a step S32, electronic controller 22 determines whether the engine coolant temperature has been reached to the predetermined temperature. If not yet reached to the predetermined temperature (No), the routine returns to the same step S32. If reached to the predetermined temperature (Yes), electronic controller 22 determines that the engine warm-up is ended and the routine goes to a step S33.
At step S33, electronic controller 22 properly outputs the control signals to intake VEL 1, intake VTC 2, and exhaust VTC 3, respectively, in accordance with the engine-speed-and-load map to control the working angle and the lift phase of intake valves 4, 4 and lift phase of exhaust valves 5, 5, respectively, in order to exhibit an optimum engine performance in accordance with the engine driving condition.
These series of controls have the same advantages as described above.
The present invention is not limited to the structures of the first and second preferred embodiments. In addition, at intake valve sides 4, 4, the combination example of intake VEL 1 and intake VTC 2 has been indicated. However, either one may be accepted. In addition, the combination example of exhaust VTC 3 at the exhaust valves 5, 5 has been shown. However, this is not always necessary.
Closure timing IC and open timing IO of the intake valve(s) may correspond to lift start and lift end timings of the intake valves and the exhaust valves. These timings may be the lift start and the lift end timings in a state in which a, so-called, ramp region is eliminated which is the minute lift region having a smooth lift gradient. The latter case corresponds to the suction start and the suction end of the substantial combustion gas and the discharge start and discharge end of the combustion gas.
Technical ideas of the present invention grasped from the first and second embodiments except the independent claims 1 through 3 will be described hereinbelow. It should be noted that a variable valve actuator corresponds to either one or both of intake VEL 1 and intake VTC 2 and a mechanically stabilizing section corresponds to biasing means, namely, either or both of coil spring 31 and coil springs 55, 56.
According to the present invention, the two variable mechanisms are used to mechanically stabilize the closure timing of the intake valve toward the retardation angle side so that the closure timing of the intake valve can sufficiently be retarded. Consequently, the rise of the crank rotation can be made at a more earlier timing.
According to the present invention, the expansion work is sufficiently performed so as to enable an increase in a combustion torque. At a time of the start of the engine at which a friction of the engine is high, a sufficient rise in the rotation can be achieved.
According to the present invention, along with the control of the crank position of the cylinder which is in the compression stroke at the constant position at the retardation angle side by means of the crank stop position control means, the crank position of the cylinder which is in the compression stroke. Hence, the suction stroke of the piston can be elongated. Thus, the atomization time of fuel can sufficiently be taken and the degree of homogeneity of mixture fuel can be increased.
On the other hand, since the closure timing of the intake valve(s) is retarded, an excessive load torque increase due to the compression at the initial stage of start can be suppressed and the vibration of the engine at the time of engine start can be suppressed.
According to the present invention, in a case where the cylinder in the expansion stroke is combusted to make a self start (or an autonomous start), the top dead center compression of the cylinder in the expansion stroke is reduced and stable. Thus, the cylinder in the compression stroke can easily and stably climb over the top dead center even with a relatively small self (or spontaneous) combustion torque. Hence, a self (or spontaneous) startability (the startability according to the self (or spontaneous) combustion) is improved.
According to the present invention, the closure timing of the intake valve is stabilized at the time of the start of the engine and the top dead center compression is determined according to the closure timing of the intake valve. Hence, the top dead center compression becomes stabilized. Thus, the start cranking characteristic can be stabilized. Consequently, the stabilization of the start can be improved.
According to the present invention, the top dead center compression can, to some degree, be reduced even in a case where such an abnormality as the operation slow-down is developed. The startability which is a minimum requirement can be assured.
Although rightmost rotation limiting stopper ST is installed on ball screw axle 23 as shown in
This application is based on a prior Japanese Patent Application No. 2010-179156 filed in Japan on Aug. 10, 2010. The entire contents of this Japanese Patent Application No. 2010-179156 are hereby incorporated by reference. Although the invention has been described above by reference to certain embodiments of the invention, the invention is not limited to the embodiments described above. Modifications and variations of the embodiments described above will occur to those skilled in the art in light of the above teachings. The scope of the invention is defined with reference to the following claims.
| Number | Date | Country | Kind |
|---|---|---|---|
| 2010-179156 | Aug 2010 | JP | national |
