JP5236786B2 - Variable valve system and variable valve apparatus for internal combustion engine - Google Patents

Description

  The present invention relates to a variable valve system and a variable valve apparatus for an internal combustion engine that can improve exhaust emission performance when the engine is started.

  As a conventional variable valve operating apparatus for an internal combustion engine, for example, one described in Patent Document 1 below is known.

  Briefly, this variable valve operating apparatus selectively supplies and discharges hydraulic pressure to and from an advance angle side oil chamber and a retard angle side oil chamber that are separated in a housing, and a vane connected to a camshaft. By rotating forward and backward, the opening / closing timing (valve timing) of the intake valve is changed according to the engine operating state.

Before the engine stops, the vane is fixed by a lock pin while being controlled to an intermediate position advanced to some extent to restrict free relative rotation between the housing and the vane. Create a reasonable valve overlap for the valve. As a result, the starting performance such as the exhaust emission performance at the time of cold start is improved.
JP 2005-233049 A

  However, in the conventional variable valve operating system for an internal combustion engine, in various sudden engine stop modes such as sudden braking of the vehicle and engine stop, the intermediate position is not a mechanically stable position. It becomes difficult to always lock the lock pin into the lock hole. Therefore, it cannot be held at an appropriate advance angle position, and there is a possibility that the exhaust emission reduction effect at the time of cold start cannot be sufficiently obtained.

  Further, as another measure, it is conceivable that the hydraulic pressure is supplied to the advance side oil chamber at the start of the cooler and the vane is appropriately rotated to the advance side to slightly increase the valve overlap. At the time of start-up, since the temperature of the hydraulic oil is low and the viscosity of the oil is high, the hydraulic oil cannot be quickly supplied to the advance side oil chamber due to the viscous resistance. As a result, the relative rotation between the vane and the housing cannot be performed smoothly, and exhaust emission may increase.

  The present invention has been devised in view of the technical problem of the conventional variable valve operating device, and when the engine is stopped, an appropriate valve overlap is preliminarily achieved by operating both the exhaust side phase change mechanism and the intake side phase change mechanism. It is an object of the present invention to provide a variable valve system and a variable valve apparatus for an internal combustion engine that can be stably created mechanically and reliably reduce exhaust emissions at the time of cold start.

The present invention is a variable valve operating device applied to a direct injection internal combustion engine that directly injects fuel into a combustion chamber,
An intake side phase conversion mechanism that makes the opening and closing timing of the intake valve variable,
An exhaust side phase conversion mechanism that makes the opening / closing timing of the exhaust valve variable,
When the engine is stopped, the intake-side phase conversion mechanism causes the intake valve opening / closing timing to be at a first position where the valve overlap between the intake valve and the exhaust valve is increased to a maximum and mechanically held. And the exhaust-side phase conversion mechanism changes the opening / closing timing of the exhaust valve to a second position where the valve overlap is mechanically held with the valve overlap reduced to a minimum,
Closing timing of the intake valve is held at the first position, and the valve overlap when the opening and closing timing of the exhaust valve is then held in the second position, the opening and closing timing of the intake valve is most The valve when the opening / closing timing of the exhaust valve is smaller than the valve overlap when the opening / closing timing of the exhaust valve is the most retarded angle, the opening / closing timing of the intake valve is the most retarded, and the opening / closing timing of the exhaust valve is the most advanced angle Set to be larger than the overlap,
When starting the engine, the intake-side phase conversion mechanism and the exhaust-side phase conversion mechanism hold the opening / closing timing of the intake valve at the first position, and the opening / closing timing of the exhaust valve at the second position. It is characterized by holding .

  According to the present invention, as in the prior art, it is not fixed to an intermediate position that is not mechanically stable when the engine is stopped, or is not forced to rotate relative to the operating hydraulic pressure when the engine is started. When the engine is stopped, an appropriate valve overlap between the exhaust valve and the intake valve can be created mechanically and stably by using both phase conversion mechanisms in advance, so that an optimal valve overlap can be ensured from the beginning of the engine start. As a result, exhaust emission can be reliably and effectively reduced.

  Embodiments of a variable valve operating apparatus for an internal combustion engine according to the present invention will be described below in detail with reference to the drawings. This embodiment shows what is applied to what is called a 4-cycle gasoline engine.

  First, as shown in FIGS. 1 to 5, the variable valve operating apparatus for an internal combustion engine according to the present invention includes an intake side timing pulley 04 to which rotational force is transmitted from a drive pulley 02 of a crankshaft 01 via a timing chain 03 and An exhaust side timing pulley 05, and an intake side camshaft 06 and an exhaust side camshaft 07 to which driving rotational force is transmitted from the respective timing pulleys 04, 05 are provided. There are provided intake-side cams 08 and 08 and exhaust-side cams 09 and 09 that open two intake valves and two exhaust valves, not shown, provided per cylinder, against the spring force of the valve spring.

  Further, between the timing pulleys 04 and 05 and the intake and exhaust camshafts 06 and 07, there is one phase conversion mechanism for controlling the opening and closing timings of the intake and exhaust valves according to the engine operating state. An exhaust side phase change mechanism (exhaust VTC) 1 and an intake side phase conversion mechanism (intake VTC) 2 which is the other phase conversion mechanism are provided.

  The exhaust VTC1 and the intake VTC2 are each of the vane type and have the same basic structure. First, the structure of the exhaust VTC 1 will be described. As shown in FIGS. 2 and 3, the timing pulley 05 that transmits the rotational force to the exhaust camshaft 07 and the end of the exhaust camshaft 07 are fixed to the timing. A vane member 3 rotatably accommodated in a pulley 05 and a hydraulic circuit 4 that rotates the vane member 3 forward and backward by hydraulic pressure are provided.

  The timing pulley 05 includes a housing 5 that rotatably accommodates the vane member 3, a disc-shaped front cover 6 that closes the front end opening of the housing 5, and a substantially disc that closes the rear end opening of the housing 5. The housing 5 and the front cover 6 and the rear cover 7 are integrally fastened together by four small-diameter bolts 8 from the axial direction of the exhaust camshaft 07.

  The housing 5 has a cylindrical shape in which both front and rear ends are formed, and shoes 5a, which are four partition walls, project inwardly at a position of about 90 ° in the circumferential direction of the inner peripheral surface. Each shoe 5a has a substantially trapezoidal cross section, and four bolt insertion holes through which the shaft portions of the respective bolts 8 pass are formed at substantially the center position, and the high-order portions of the respective inner end surfaces. A U-shaped seal member 9 and a leaf spring (not shown) that presses the seal member 9 inwardly are fitted and held in a holding groove that is cut out at a position along the axial direction.

  The front cover 6 is formed in a disk plate shape, and a relatively large-diameter support hole 6a is formed in the center, and the outer peripheral portion is not shown at a position corresponding to each bolt insertion hole of the housing 5. These four bolt holes are drilled.

  The rear cover 7 is integrally provided with a gear portion 7a that meshes with the timing chain on the rear end side, and a large-diameter bearing hole 7b is formed in the axial direction so as to penetrate therethrough.

  The vane member 3 includes an annular vane rotor 3a having a bolt insertion hole in the center, and four vanes 3b provided integrally at a position approximately 90 ° in the circumferential direction of the outer peripheral surface of the vane rotor 3a.

  In the vane rotor 3a, a small-diameter cylindrical portion on the front end side is rotatably supported by the support hole 6a of the front cover 6, while a small-diameter cylindrical portion on the rear end side is freely rotatable in the bearing hole 7b of the rear cover 7. It is supported.

  The vane member 3 is fixed to the front end portion of the exhaust camshaft 07 from the axial direction by a fixing bolt 9 inserted from the axial direction into the bolt insertion hole of the vane rotor 3a.

  Each of the vanes 3b is formed in a relatively long and narrow rectangular shape, and the other one is formed in a relatively large trapezoidal shape. The three vanes 3b are substantially the same in width and length. In contrast, the width of one vane 3b is set to be larger than that of the three vanes 3 to balance the weight of the vane member 3 as a whole.

  Each vane 3b is disposed between the shoes 5a, and has a U-shaped seal member 10 slidably contacting the inner peripheral surface of the housing 5 in an elongated holding groove formed in the axial direction of each outer surface. Leaf springs that press the seal member 10 toward the inner peripheral surface of the housing 5 are fitted and held. Further, two substantially circular concave grooves 3c are formed on one side of each vane 3b on the same side as the rotation direction of the exhaust camshaft 07, respectively.

  Further, four advance-side oil chambers 11 and retard-side oil chambers 12 are respectively separated between both sides of each vane 3b and both sides of each shoe 5a.

  As shown in FIG. 2, the hydraulic circuit 33 is connected to the first hydraulic passages 13 for supplying and discharging the hydraulic oil pressure to and from the advance side oil chambers 11, and to the retard side oil chambers 12. There are two systems of hydraulic passages, a second hydraulic passage 14 for supplying and discharging the hydraulic oil pressure, and a supply passage 15 and a drain passage 16 are provided in both the hydraulic passages 13 and 14 for electromagnetic switching for passage switching. It is connected via a valve 17. The supply passage 15 is provided with a one-way oil pump 19 for pumping oil in the oil pan 18, while the downstream end of the drain passage 16 communicates with the oil pan 18.

  The first and second hydraulic passages 13 and 14 are formed inside a cylindrical passage constituting portion 20, and one end portion of the passage constituting portion 20 extends from the small diameter cylindrical portion of the vane rotor 3 a to the inside support hole 3 d. The other end is connected to the electromagnetic switching valve 17.

  Further, between the outer peripheral surface of the one end portion of the passage constituting portion 20 and the inner peripheral surface of the support hole 3d, three annular seal members 21 for separating and sealing between one end sides of the hydraulic passages 13 and 14 are fitted. It is fixed.

  The first hydraulic passage 13 is formed in an oil chamber 13a formed at the end of the support hole 3d on the drive shaft 6 side, and substantially radially inside the vane rotor 3a, so that the oil chamber 13a and each advance side oil chamber are formed. And four branch passages 13 b communicating with 11.

  On the other hand, the second hydraulic passage 14 is stopped within one end portion of the passage constituting portion 20, and is formed into an annular chamber 14a formed on the outer peripheral surface of the one end portion and a substantially L-shape inside the vane rotor 3a. The annular chamber 14a and a second oil passage 14b communicating with each retarded angle side oil chamber 12 are provided.

  The electromagnetic switching valve 17 is a four-port three-position type, and an internal valve body is configured to relatively switch and control each of the hydraulic passages 13 and 14, the supply passage 15 and the drain passage 16, and Switching operation is performed by a control signal from the controller 22.

  The controller 22 detects the engine operating state, and at the same time, receives signals from the crank angle sensor 23, the cam angle sensor 24 of the intake cam 08, and the cam angle sensor 25 of the exhaust cam 09, and the timing pulleys 04, 05 and the camshaft 06. , 07 and relative rotation position are detected.

  Then, by the switching operation of the electromagnetic switching valve 17, hydraulic oil is supplied to the retard side oil chamber 12 when the engine is started, and thereafter, hydraulic oil is supplied to the advance side oil chamber 11.

  Further, between the vane member 3 and the housing 5, there is provided a lock mechanism which is a fixing means for restraining the rotation of the vane member 3 with respect to the housing 5 and releasing the restraint.

  That is, as shown in FIG. 2, this locking mechanism is provided between the one vane 3b having a large width and the thick rear cover 7, and the axial direction of the exhaust camshaft 07 inside the vane 3b. A sliding hole 26 formed along the sliding hole 26, a covered cylindrical lock pin 27 slidably provided in the sliding hole 26, and a transverse section fixed in a fixing hole provided in the rear cover 7. An engagement hole 28 a provided in the cup-shaped engagement hole constituting portion 28, in which the tapered tip end portion 27 a of the lock pin 27 engages and disengages, and a spring retainer 29 fixed to the bottom surface side of the sliding hole 26. The coil spring-like spring member 30 is held and biases the lock pin 27 toward the engagement hole 28a. The engagement hole 28a is supplied with hydraulic pressure from the retarded angle side oil chamber 12 through an oil hole (not shown).

  The lock pin 27 is engaged with the engagement hole 28a by the spring force of the spring member 30 at the position (first position) where the vane member 3 is rotated to the most retarded angle side. The relative rotation between the timing pulley 05 and the exhaust camshaft 07 is locked. Further, the lock pin 27 is moved backward by the hydraulic pressure supplied from the retard side oil chamber 12 into the engagement hole 28a, and the engagement with the engagement hole 28a is released.

  Further, a pair of coil springs 31 which are urging means for urging and rotating the vane member 3 to the retard side are provided between one side surface of each vane 3b and the opposite surface of each shoe 5a facing the one side surface. Are arranged respectively.

  The two coil springs 31 are formed independently of each other and formed in parallel with each other. The axial lengths (coil lengths) of the two coil springs 31 are one side surface of the vane 3b and the opposite surface of the shoe 5a. Is set to be larger than the length between the two, and both are set to the same length.

  In addition, the coil springs 31 are arranged side by side with an inter-axis distance that does not come into contact with each other at the time of maximum compression deformation, and each end is fitted to a concave groove 3c of each shoe 5a. It is connected through.

  In this embodiment, the conversion angle of the vane member 3 on the exhaust side, that is, the difference θe between the most retarded angle (FIG. 3) and the most advanced angle (FIG. 4) is as shown in FIGS. For example, it is set to about 15 °.

  On the other hand, as shown in FIG. 5, the intake VTC2 is the same as the exhaust VTC1, and therefore, the same reference numerals are assigned to the common components and the detailed description thereof is omitted. The control position (second position) of the intake valve to the maximum retarded angle side by the intake VTC2, that is, the position where the vane member 3 is rotationally controlled to the maximum retarded angle side shown in FIG. The conversion angle θi at the opening time (IVO) at approximately 25 °.

  Hereinafter, the operation of the exhaust VTC 1 will be described. First, when the engine is stopped by turning off the ignition key, the output of the control current from the controller 22 to the electromagnetic switching valve 17 is stopped, and the valve body is pushed by the spring force of the spring to the position shown in FIG. Become. Therefore, the supply passage 15 and the retarded-side second hydraulic passage 14 are communicated. Therefore, the vane member 3 tries to rotate to the retard side by the supply hydraulic pressure, but when the engine speed approaches zero, the oil pressure of the oil pump 19 does not act, and the supply hydraulic pressure also becomes zero.

  Here, as shown in FIG. 3, the vane member 3 is moved to the exhaust camshaft 07 with respect to the timing pulley 05 by valve friction caused by an alternating torque acting on the exhaust camshaft 07 and the spring force of each coil spring 31. Relative rotation in the counterclockwise direction opposite to the rotation direction (arrow direction) is stabilized.

  Accordingly, the vane member 3 is held at a position (first position) where the vane 3b having the maximum width comes into contact with the side surface of the shoe 5a on the advance side oil chamber 11 side, so that the timing pulley 05 and the exhaust camshaft 07 are provided. Relative rotation phase to the maximum retard angle side.

  At the same time, at the first position, the distal end portion 27a of the lock pin 27 is engaged in the engagement hole 28a to restrict free relative rotation between the timing pulley 05 and the exhaust camshaft 07.

  Therefore, even when the rotational fluctuation is large as in cranking at the time of engine restart, the vane is mechanically stable at the maximum retard angle (first position) and is fixed by the lock pin 27. It is possible not only to stabilize the phase of the member 3, that is, the exhaust camshaft 07, but also to suppress its fluttering, and as a result, to suppress the instability of the valve timing control and to surely obtain good startability and reduction of cooler emissions. Can do.

  On the other hand, as shown in FIG. 5, the intake side VTC 2 is the same as the exhaust side VTC 1, and the vane 3 b having the maximum width of the vane member 3 is caused by the valve side friction and the spring force of each coil spring 31. The relative rotational phase between the timing pulley 05 and the intake camshaft 06 is changed and held at the maximum retarded angle side in contact with the side surface on the chamber 11 side, and at this time, the tip of the lock pin 27 is engaged as in the exhaust side. The vane member 3 engages with the hole 28 and is stably and reliably held at the most retarded rotation position (second position). Thus, the opening timing (IVO) of the intake valve in the intake stroke is held at the most retarded position near the top dead center.

  Therefore, as shown in FIG. 6, the exhaust valve closing timing (EVC) in the exhaust stroke of each cylinder is delayed by θe × 2 in terms of the crank angle, that is, after the top dead center, for example, about The position is on the retarded side of 30 °.

  Therefore, for example, the valve overlap with the intake valve becomes an appropriate size of about 30 °.

  Therefore, when the engine in the cold state is started in this state, the residual gas flows back to the intake system due to the appropriate valve overlap, and the unburned HC gas is reburned, or the residual gas temperature is The atomization of the warm fuel is promoted, and the generation of HC gas can be sufficiently suppressed.

  Further, at this time, if the valve overlap is excessively increased, the inert gas that is a residual gas in the combustion chamber is remarkably increased, so that a desired combustion torque cannot be generated. Although it is likely to cause instability, since it is the appropriate valve overlap of about 30 °, it is possible to improve the cold engine emission performance while avoiding instability of engine rotation.

  Further, when the engine is started, the hydraulic oil pumped from the oil pump 19 by the control current from the controller 22 to the electromagnetic switching valves 17 and 17 on both the exhaust side and the intake side is supplied to the retard side oil chambers 12 and 12, respectively. The vane members 3 and 3 are supplied and held together with the spring force of the coil springs 31 and 31 to the most retarded angle side, but as the hydraulic pressure in the retarded angle side oil chambers 12 and 12 increases, 27 comes out of the engagement hole 28a and allows the vane members 3 and 3 to freely rotate.

  Here, since the electromagnetic switching valve 17 on the intake side is continuously energized from the controller 22 and hydraulic oil is supplied to the retard side oil chamber 12, the intake side vane member 3 is a spring of the coil spring 31. The relative rotational position is held and controlled to the most retarded angle side by the force and the hydraulic pressure. Therefore, the opening / closing timing of the intake valve continues as it is, and as shown in FIG. 7, the opening timing (IVO) is at the top dead center position, and the closing timing (IVC) is sufficiently delayed from the bottom dead center. Controlled at different times.

  On the other hand, a switching control current is output from the controller 22 to the electromagnetic switching valve 17 on the exhaust side, and the hydraulic oil pressure-fed from the oil pump 19 is supplied to the retarding oil chamber 12 side. For this reason, the relative position of the vane member 3 is held and controlled to the most retarded angle side. Thus, as shown in FIG. 6, the exhaust valve is controlled so that the closing time (EVC) is approximately 30 ° after top dead center. Therefore, the above-mentioned exhaust emission reduction effect is sustained.

  Next, when the warm-up proceeds, the lock pin has already been removed, and control is performed as shown in FIG. 7 at low load. In particular, the exhaust side VTC is controlled to the advance side. Therefore, the valve overlap between the intake valve and the exhaust valve is almost zero. When this valve overlap is almost zero, the residual gas is small, a good combustion state is obtained, and the combustion is stabilized, so that the engine rotation can be stabilized and the exhaust emission is also improved.

  Next, for example, when the engine is in an intermediate load region or a low rotation / high load, a switching signal is output from the controller 22 to each of the electromagnetic switching valves 17 and 17 on the exhaust side and the intake side. Cuts off energization and connects the supply passage 15 and the second hydraulic passage 14 and also connects the first hydraulic passage 13 and the drain passage 16, while the supply passage 15 and the first hydraulic passage are connected to the intake side electromagnetic switching valve 17. 13 and the second hydraulic passage 14 and the drain passage 16 are communicated.

  Therefore, in the exhaust VTC1, the hydraulic pressure is supplied to each retarded angle oil chamber 12, the vane member 3 is relatively rotated to the most retarded angle side, and in the intake VTC2, the hydraulic pressure is applied to each advanced angle side oil chamber 11. Is supplied, and the vane member 3 rotates relative to the most advanced angle side. For this reason, the opening / closing timing of the exhaust valve has the characteristics shown in FIG. 8, and the closing timing (EVC) of the same level as that at the start becomes a conversion angle of about 30 ° after top dead center. The characteristics shown in FIG. 8 are obtained, and the opening time (IVO) is a conversion angle of about 50 ° before top dead center.

  Therefore, since the valve overlap is created as large as about 30 ° + about 50 ° = about 80 °, the pumping loss can be reduced and the fuel consumption can be improved. In other words, since the combustion torque is large in such a medium load region, there is little instability of combustion that is likely to occur in the light load to low load operation region, so the valve overlap can be increased. The fuel consumption improvement effect can be obtained. Here, the exhaust VTC1 does not necessarily have to be the most retarded angle, and the intake VTC2 does not necessarily have to be the most advanced angle.

  Hereinafter, specific control by the controller 22 at the time of engine start (at the time of cold start) will be described with reference to the control flowchart of FIG.

  First, in Step 1, it is determined whether or not the ignition switch is turned on, that is, whether or not the engine is started. If not, the process returns. Proceed to 2.

  In step 2, it is recognized that the engine is in the cranking state, but before this cranking, in principle, the exhaust valve and the intake valve are opened and closed by the exhaust VTC1 and the intake VTC2 in advance as shown in FIG. The vane member 3 is fixed and held by the lock pin 27.

  In step 3, the opening and closing timings of the intake valves and the exhaust valves have the characteristics shown in FIG. 6 depending on the spring force of the coil springs 31 and 31 by the exhaust VTC1 and the intake VTC2. Thus, a control signal is output to each of the electromagnetic switching valves 17 and 17 so that the opening / closing timings of the intake valve and the exhaust valve have the characteristics shown in FIG. That is, the hydraulic oil is forcibly supplied to each retarded angle side oil chamber 12, 12. In addition, since the hydraulic pressure is supplied to the engagement holes 28a and 28a as the hydraulic pressure is supplied to the retard angle side oil chambers 12 and 12, the engagement of the lock pins 27 and 27 is released, and the vane members 3 and 3 are released. As described above, the free rotation of 3 is allowed, but even in this state, the holding control is performed with the characteristics shown in FIG.

  In step 4, a control signal is output to the fuel injection valve and the spark plug to perform a complete explosion in the combustion chamber. Therefore, even during the complete explosion, the intake valve and the exhaust valve are maintained and controlled at the specific opening / closing timing shown in FIG. 6, so that the exhaust emission performance at the time of cold start described above can be improved.

  Next, in step 5, the current engine rotation state is detected from a crank angle sensor or the like. Subsequently, in step 6, it is determined whether or not the engine rotation is stabilized. If it is determined that the engine rotation is stable, the process proceeds to step 7, but if it is determined that it is unstable, Move on to step 8.

  In step 8, the hydraulic pressure is supplied to the advance side oil chamber 11 by the exhaust VTC1 to advance the exhaust valve closing timing (EVC), thereby controlling the valve overlap with the intake valve. Do. This stabilizes combustion and suppresses engine rotation fluctuations. This destabilization of the engine reduces the valve clearance with time, increases the actual valve overlap period, increases the exhaust system resistance, and increases the residual gas for the same valve overlap amount, etc. Caused by In this case, the above-described correction for reducing the valve overlap suppresses an unintended increase in residual gas, thereby ensuring the stability of the engine.

  In step 7, it is determined whether or not a predetermined time has elapsed since the start of cranking by a timer. If it is determined that the predetermined time has not elapsed, the process returns to step 5; It is determined that the cold machine control is finished, and the process proceeds to step 9. Here, the predetermined time can be changed according to the air temperature, the temperature or humidity of the engine before starting, or the like.

  In this step 9, the exhaust VTC1 and the intake VTC2 are subjected to normal control at the time of warming up and after warming up based on a control map prepared in advance according to engine operation, and then the process returns.

  The normal control in step 9 is, for example, a warm-up that tends to improve the fuel consumption by reducing the pumping loss as a valve overlap close to the large valve overlap shown in FIG. 8 at medium load, or to increase the residual gas. In idle mode, control is performed such that the small valve overlap is close to the valve overlap shown in FIG. 7 and the rotational stability of the engine is improved.

  Incidentally, the mechanically stable valve timing shown in FIG. 6 is that the conversion angle θe (about 15 °) of the exhaust valve to the maximum retarded side by the exhaust VTC1 is the maximum retarded side of the intake valve by the intake VTC2. Since it is smaller than the conversion angle θi (approximately 25 °), the valve overlap is relatively small. For this reason, not only can the valve overlap be advantageous for the exhaust emission performance suitable at the time of starting, but also the engine rotation can be stabilized to some extent even in a warm-up state at the time of fail-safe such as an abnormality in the electric system.

  Further, when the engine is started, in addition to being mechanically stable at the valve timing shown in FIG. 6, the free rotation of the vane member 3 is surely restricted by the lock mechanism. Even when there is a fluctuation in rotation, the actual valve timing (opening / closing timing) of the intake valve and the exhaust valve can be stabilized, and as a result, desired exhaust emission can be reliably reduced.

  In this embodiment, the vane members 3 and 3 of the exhaust VTC1 and the intake VTC2 are urged to rotate to the most retarded angle side by the spring force of the coil springs 31 and 31, respectively. And, the valve overlap can be surely created at the time of starting. Thereby, reduction of the exhaust emission at the time of cold machine starting can be obtained more reliably.

[Second Embodiment]
FIG. 10 and the subsequent figures show a second embodiment, in which one phase conversion mechanism is an intake VTC2 and the other phase conversion mechanism is an exhaust VTC1, and each exhaust valve and each intake valve are controlled by the exhaust VTC1 and the intake VTC2. The opening / closing timing of each is mechanically stabilized to the advance side when the engine is stopped.

  That is, FIGS. 10 to 12 show the structure of the exhaust VTC1, but the intake VTC2 shown in FIG. 13 is also formed in the same structure, and the basic structure is the same as that of the first embodiment. However, the electromagnetic switching valve is in the position shown in FIG. 10 by the spring, and the supply passage 15 of the hydraulic circuit 4 communicates with the first hydraulic passage 13 when the engine is stopped and started, and the second hydraulic passage 14 is connected to each of them. It communicates with the drain passage 16. Each advance side oil chamber 11 communicating with the first hydraulic passage 13 communicates with the engagement hole 28a of the lock mechanism.

  Further, the vane members 3 and 3 of the VTCs 1 and 2 are always urged to rotate toward the advance side by a pair of coil springs 31 disposed in each advance side oil chamber 11. Although the specific configuration of each coil spring 31 is the same as that of the first embodiment, each coil spring 31 urges the vane member 3 toward the advance side against the valve friction. Therefore, the spring load is set to be larger than that of the first embodiment.

  Further, as shown in FIG. 13, the conversion angle θi of the vane member 3 by the intake VTC2 is set to, for example, 25 °, and the conversion angle θe (for example, 15 °) of the vane member 3 by the exhaust VTC1 shown in FIG. Is set larger than. Therefore, the valve overlap amount at the default position (position at the time of engine stop and engine start) is, for example, 50 ° as shown in FIG. 14, and is larger than that of the first embodiment. Yes.

  Therefore, the residual gas in the combustion chamber increases during the operation of the engine, but this is required when applied to a direct injection engine that directly injects fuel into the combustion chamber as the internal combustion engine of the present embodiment. In a direct injection engine, a high compression ratio can be achieved in the combustion chamber due to a cooling effect. This high compression ratio can ensure the stability of combustion during cold operation, and can further increase the limit of the valve overlap amount. Therefore, exhaust emission at the time of cold start can be sufficiently reduced.

  This is because in the case of a direct injection engine, fuel can be supplied to the combustion chamber even when the intake valve is closed, and the degree of freedom of the injection pattern is large, so there is room for improving combustion.

  Further, during the idling operation after the warm-up is completed, the control of the intake VTC 2, that is, the controller 22 controls the switching of the electromagnetic switching valve 17 to connect the first hydraulic passage 13 and the drain passage 16, and the supply passage 15 and the second passage As shown in FIG. 12, the vane member 3 is communicated with the hydraulic passage 14 to supply hydraulic pressure into each retarded angle side oil chamber 12, and the timing pulley 05 is resisted against the spring force of each coil spring 31. Is rotated in the opposite direction to the rotation direction of the valve, and the opening / closing timing of the intake valve is controlled to the most retarded angle side. On the other hand, the exhaust VTC 1 continues the control as it was at the start, and controls the opening / closing timing of the exhaust valve to the most advanced angle side. Therefore, as shown in FIG. 15, the exhaust valve closing timing (EVC) is maintained at approximately the top dead center position, and the intake valve opening timing (IVO) is also controlled to approximately the top dead center position. As a result, the valve overlap becomes substantially zero, and the same effect as in the case shown in FIG. 7 can be obtained.

  Further, for example, when shifting to a medium load, low rotation, high load operation, the exhaust valve opening / closing timing is controlled to the most retarded angle side by the exhaust VTC1, and the intake valve VTC2 is used to control the intake valve opening timing. Is controlled to the most advanced angle side. As a result, the exhaust valve closing timing (EVC) is about 30 ° after the top dead center and the intake valve opening timing (IVO) is about 50 ° before the top dead center. Therefore, the valve overlap amount is about 80 °. Therefore, also in this case, the same effect as that shown in FIG. 8 can be obtained.

  Hereinafter, specific control of the controller 22 in the present embodiment will be described based on the flowchart shown in FIG. 17, but the basic flow is the same as that in FIG. 9 in the first embodiment. The difference is that when the engine is stopped or in step 13, both the exhaust valve and the intake valve are controlled to the most advanced angle side by the exhaust VTC1 and the intake VTC2, respectively, and an appropriate valve overlap is created. (FIG. 13) When the combustion is determined to be unstable in step 15, the intake valve closing timing (IVO) is corrected and controlled to the retard side in step 18 to reduce the valve overlap amount. Is a point.

  Since other processes and determinations are the same as those shown in FIG.

  Therefore, similarly to the first embodiment, the second embodiment can surely reduce the exhaust emission by creating an appropriate valve overlap at the time of cold start.

  The present invention is not limited to the configuration of each of the above embodiments. For example, the coil springs 31 and 31 in the first embodiment are not necessarily required, and even if they do not exist, the valve When the engine is stopped due to friction or the like, the vane members 3 and 3 are automatically rotated to the most retarded angle side. However, in the second embodiment, since the vane members 3 and 3 must be rotated to the most advanced angle side against the valve friction, the coil springs 31 and 31 are indispensable.

  Further, the first embodiment is not applied to a so-called direct injection internal combustion engine, but can be applied to this as in the second embodiment.

  In the second embodiment, a direct injection internal combustion engine is applied. However, it is also possible to use a port injection internal combustion engine in which a lift difference is provided in the intake two valves to improve combustion stability.

  Furthermore, in each of the above-described embodiments, the ignition method using a spark plug is used. However, the present invention can also be applied to, for example, an engine using an ignition method using compression. In this ignition method, ignition control is not included in the above-described complete explosion control, but the exhaust emission reduction effect at the time of cold start can be sufficiently obtained as in the spark plug method.

It is a perspective view which shows the main structural members of the internal combustion engine provided for embodiment of the variable valve apparatus which concerns on this invention. It is a longitudinal section showing exhaust VTC provided for a 1st embodiment. It is the sectional view on the AA line of FIG. 2 which shows the most retarded angle control state in this Embodiment. It is the sectional view on the AA line of FIG. 2 which shows the most advanced angle control state in this Embodiment. It is sectional drawing which shows the intake VTC in this embodiment. It is a characteristic figure which shows the open period of the intake / exhaust valve at the time of an engine stop and a start. It is a characteristic view which shows the open period of the intake / exhaust valve at the time of idling operation after completion of warm-up. It is a characteristic view which shows the open period of the intake / exhaust valve at the time of driving | running | working, such as medium load. It is a flowchart figure which shows the control by the controller of this embodiment. It is a longitudinal section showing exhaust VTC provided for a 2nd embodiment. It is the BB sectional view taken on the line of FIG. 10 which shows the most advanced angle control state in this Embodiment. It is the BB sectional view taken on the line of FIG. 10 which shows the most retarded angle control state in this Embodiment. It is sectional drawing which shows the intake VTC in this embodiment. It is a characteristic figure which shows the open period of the intake / exhaust valve at the time of an engine stop and a start. It is a characteristic view which shows the open period of the intake / exhaust valve at the time of idling operation after completion of warm-up. It is a characteristic view which shows the open period of the intake / exhaust valve at the time of driving | running | working, such as medium load. It is a flowchart figure which shows the control by the controller of this embodiment.

04 ... Intake side timing pulley 05 ... Exhaust side timing pulley 06 ... Intake camshaft 07 ... Exhaust camshaft 1 ... Exhaust VTC (Exhaust side phase conversion mechanism)
2 ... Intake VTC (Intake side phase change mechanism)
DESCRIPTION OF SYMBOLS 3 ... Vane member 4 ... Hydraulic circuit 5 ... Housing 22 ... Controller 27 ... Lock pin 28a ... Engagement hole 31 ... Coil spring

Claims (1)

A variable valve operating device applied to a direct injection internal combustion engine that directly injects fuel into a combustion chamber,
An intake side phase conversion mechanism that makes the opening and closing timing of the intake valve variable,
An exhaust side phase conversion mechanism that makes the opening / closing timing of the exhaust valve variable,
When the engine is stopped, the intake-side phase conversion mechanism causes the intake valve opening / closing timing to be at a first position where the valve overlap between the intake valve and the exhaust valve is increased to a maximum and mechanically held. And the exhaust-side phase conversion mechanism changes the opening / closing timing of the exhaust valve to a second position where the valve overlap is mechanically held with the valve overlap reduced to a minimum,
Closing timing of the intake valve is held at the first position, and the valve overlap when the opening and closing timing of the exhaust valve is then held in the second position, the opening and closing timing of the intake valve is most The valve when the opening / closing timing of the exhaust valve is smaller than the valve overlap when the opening / closing timing of the exhaust valve is the most retarded angle, the opening / closing timing of the intake valve is the most retarded, and the opening / closing timing of the exhaust valve is the most advanced angle Set to be larger than the overlap,
When starting the engine, the intake-side phase conversion mechanism and the exhaust-side phase conversion mechanism hold the opening / closing timing of the intake valve at the first position, and the opening / closing timing of the exhaust valve at the second position. A variable valve operating system for an internal combustion engine, characterized by being held .

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