JP4135394B2 - Control device for internal combustion engine - Google Patents

Description

[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a control device for an internal combustion engine that includes a variable compression ratio mechanism that variably controls an engine compression ratio, and more particularly to a fuel efficiency improvement technique at the time of partial load in a gasoline engine that includes a variable valve mechanism on the intake valve side.
[0002]
[Prior art]
The present applicant has previously proposed a variable valve mechanism capable of continuously expanding and reducing the lift and operating angle of the intake valve, and further combining it with a mechanism for delaying the phase of the center angle of the lift. Thus, there have been proposed variable valve mechanisms that can obtain a large degree of freedom in lift characteristics (see, for example, Japanese Patent Application Laid-Open Nos. 2002-89303 and 2002-89341).
[0003]
Further, the present applicant uses a multi-link type piston-crank mechanism as a variable compression ratio mechanism of a reciprocating internal combustion engine, and changes a piston top dead center position by moving a part of the link configuration. Various proposals have been made (for example, JP-A-2002-215952). This type of variable compression ratio mechanism changes the mechanical compression ratio of the internal combustion engine, that is, the nominal compression ratio, and is generally controlled at a high compression ratio to improve thermal efficiency during partial load, and at high load, The compression ratio is controlled to be low in order to avoid knocking.
[0004]
[Problems to be solved by the invention]
When the multi-link type piston-crank mechanism as described above is used, in order to reduce the friction loss and wear due to the piston thrust load, it is connected to the piston as described in JP-A-2002-215952. It is desirable to set the link configuration so that the angle (first link tilt angle) θ formed by the first link and the cylinder axis is smaller in the first half of the expansion stroke than in the second half of the compression stroke. That is, the combustion pressure acting on the piston becomes maximum in the first half of the expansion stroke, and this is larger than the reaction force acting on the piston during the compression stroke, so that the tilt angle θ of the first link is the compression top dead center. The piston thrust load caused by the combustion pressure is reduced by making it the smallest in the later position, that is, the first half of the expansion stroke in which the combustion pressure acts. In other words, it is desirable that the first link be in the state closest to the vertical posture in the vicinity where the combustion pressure becomes maximum after the piston top dead center.
[0005]
However, when the link configuration is set such that the tilt angle θ of the first link is minimized at a position delayed from the top dead center, the tilt angle θ of the first link in the latter half of the compression stroke is reversed. It grows that much. Therefore, the piston thrust load generated in the latter half of the compression stroke is rather increased as compared with the link configuration in which the tilt angle θ is minimized at the piston top dead center position. Of course, in terms of friction loss throughout the cycle, it is advantageous that the tilt angle θ is minimized in the first half of the expansion stroke, but particularly in a state where the variable compression ratio mechanism has a high compression ratio, it is accompanied by compression. Since the reaction force to the piston becomes large, the piston thrust load in the latter half of the compression stroke becomes relatively large, and there is still room for improvement in terms of friction loss and wear.
[0006]
Therefore, the present invention uses a variable valve mechanism provided for reducing pump loss and the like to suppress the piston thrust load in the latter half of the compression stroke and further improve the friction loss and wear due to the piston thrust load. With the goal.
[0007]
The present invention provides a variable compression ratio mechanism in which a piston top dead center is changed by moving a part of the link structure using a multi-link type piston-crank mechanism as in claim 1, and at least a closing timing of the intake valve An internal combustion engine control device that controls the compression ratio and the intake valve closing timing in accordance with the operating conditions.
The variable compression ratio mechanism has a link configuration so that the angle θ formed between the first link connected to the piston and the cylinder axis is smaller in the first half of the expansion stroke than in the second half of the compression stroke and is minimized in the first half of the expansion stroke. Is set,
The intake valve so that the piston thrust load in the latter half of the compression stroke when the angle θ is relatively large at the high compression ratio is equal to or smaller than the piston thrust load in the first half of the expansion stroke where the angle θ is minimized. by controlling the closing timing, the actual compression ratio determined by the intake valve closing timing, is characterized by smaller than the actual expansion ratio determined by the opening time exhaust valve.
[0008]
Here, the actual compression ratio is a compression ratio when compression is performed up to the piston top dead center on the assumption that compression starts from the intake valve closing timing, and the actual expansion ratio expands from the piston top dead center and exhausts. These expansion ratios assume that the expansion ends at the valve opening timing, and the specific values of both are dependent on the mechanical compression ratio. The ratio has no effect. In other words, the relationship that “the actual compression ratio is smaller than the actual expansion ratio” is “(actual compression ratio / actual expansion ratio) <1”, and the mechanical compression ratios that are the basis of each are the same. Therefore, there is no need to consider, and the above magnitude relationship is determined by the intake valve closing timing. The relationship “(actual compression ratio / actual expansion ratio) <1” means that the actual compression ratio is positively suppressed.
[0009]
In order to lower the actual compression ratio in this way, either the so-called early closing method of closing the intake valve earlier than the intake bottom dead center or the so-called delayed closing method of closing later than the intake bottom dead center may be used. .
[0010]
For example, in the invention of claim 4, in order to reduce the actual compression ratio, the intake valve operating angle is set to 180 ° or less in crank angle, and the intake valve closing timing is advanced from the intake bottom dead center. .
[0011]
Alternatively, in the invention of claim 5, in order to reduce the actual compression ratio, the intake valve operating angle is set to 180 ° or more in crank angle, and the intake valve closing timing is retarded from the intake bottom dead center. .
[0012]
That is, the nominal compression ratio (mechanical compression ratio) is changed by the variable compression ratio mechanism, but even if the same nominal compression ratio, for example, if the intake valve is closed at a position advanced from the intake bottom dead center, the actual compression ratio will be The compression ratio decreases. The same applies to the case of late closing that closes after the intake bottom dead center. By reducing the actual compression ratio in this way, the reaction force acting on the piston in the latter half of the compression stroke is reduced, and the piston thrust load is suppressed.
[0013]
The compression ratio control by the variable compression ratio mechanism is generally controlled to a high compression ratio on the low load side where knocking does not become a problem and to a relatively low compression ratio on the high load side. A decrease in the compression ratio is not a problem, and it is advantageous in terms of pump loss by suppressing the actual compression ratio by the intake valve closing timing while increasing the nominal compression ratio.
[0014]
The invention of claim 2 further reduced from the invention of claim 1 comprises a supercharger,
In the supercharging region where the cylinder pressure during the intake stroke is higher than atmospheric pressure, the compression ratio is controlled to be low,
In the non-supercharging region where the cylinder pressure in the intake stroke is lower than atmospheric pressure, the compression ratio is controlled to be high, and the intake valve closing timing is controlled so that the actual compression ratio is smaller than the actual expansion ratio. It is characterized by that.
[0015]
The supercharger may be a turbocharger or a mechanical supercharger.
[0016]
Since knocking becomes a problem in the supercharging region, the compression ratio is controlled to be low. However, in the non-supercharging region, it is desirable to increase the compression ratio in terms of thermal efficiency, combustion stability, and the like. In the present invention, the intake valve closing timing is controlled so that the actual compression ratio becomes smaller than the actual expansion ratio in this non-supercharging region, and the piston thrust load is suppressed.
[0017]
Desirably, the compression ratio is controlled to be lower as the supercharging pressure becomes higher.
[0018]
In addition, the variable valve mechanism, as described in claim 6, delays the phase of the lift / working angle variable mechanism capable of continuously expanding / reducing the lift / working angle of the intake valve and the center angle of the lift of the intake valve. The phase variable mechanism to be advanced can be configured by either one or a combination of both.
[0019]
As the lift / operating angle variable mechanism, as in claim 7, an eccentric cam that is rotationally driven by a drive shaft, a link arm that is fitted to the outer periphery of the eccentric cam so as to be relatively rotatable, and a parallel to the drive shaft. A pivotable control shaft provided with an eccentric cam portion, a rocker arm rotatably mounted on the eccentric cam portion of the control shaft and swung by the link arm, and rotated on the drive shaft And an oscillating cam that presses the tappet of the intake valve by oscillating along with the rocker arm and oscillating along with the rocker arm. It is possible to use a configuration in which the lift / operating angle of the intake valve is increased or decreased simultaneously by changing the rotation position of the part.
[0020]
As the variable compression ratio mechanism, as in claim 8, a first link connected to the piston via a piston pin, a swing pin connected to the first link and a crank pin of the crankshaft. A second link rotatably connected to the part, and a third link swingably connected to the second link and supported swingably on the engine body. It is possible to use one configured to perform variable control of the compression ratio by changing the fulcrum position of the link relative to the engine body.
[0021]
【The invention's effect】
According to the present invention, in the link configuration of the variable compression ratio mechanism using the multi-link type piston-crank mechanism, the angle θ formed by the first link connected to the piston and the cylinder axis is expanded more than the latter half of the compression stroke. By setting it to be small in the first half of the stroke, the piston thrust load due to combustion pressure can be reduced, and conversely, the piston thrust load in the latter half of the compression stroke that tends to increase is suppressed by controlling the intake valve closing timing can do. Therefore, the piston thrust load can be reduced as a whole, and friction loss and wear can be reduced.
[0022]
In particular, when combined with a supercharger, a significant reduction in fuel consumption can be achieved in a non-supercharged region.
[0023]
DETAILED DESCRIPTION OF THE INVENTION
Hereinafter, preferred embodiments of the present invention will be described in detail with reference to the drawings.
[0024]
FIG. 1 shows an embodiment of a control device for an internal combustion engine according to the present invention. The internal combustion engine includes a variable valve mechanism 101 for variably controlling the intake valve opening / closing timing, a compression ratio variable mechanism 102 for variably controlling the nominal compression ratio (mechanical compression ratio) ε of the internal combustion engine, and an ignition timing control. An ignition advance control device 103 that performs this operation and a turbocharger 104 that will be described later are provided.
[0025]
FIG. 2 is an explanatory diagram showing the configuration of the variable valve mechanism 101. The variable valve mechanism includes a lift / operating angle variable mechanism 1 that changes the lift / operating angle of the intake valve 12, and the lift. A phase variable mechanism 2 that advances or retards the phase of the central angle (phase relative to the crankshaft) is combined.
[0026]
First, the lift / operating angle variable mechanism 1 will be described with reference to the operation explanatory diagram of FIG. The lift / operating angle variable mechanism 1 was previously proposed by the applicant of the present invention, but is known together with the phase variable mechanism 2 from the above-mentioned JP-A-2002-89303 and JP-A-2002-89341. Therefore, only the outline will be explained.
[0027]
The lift / operating angle variable mechanism 1 includes a hollow drive shaft 13 rotatably supported by a cam bracket (not shown) above the cylinder head, an eccentric cam 15 fixed to the drive shaft 13 by press-fitting or the like, A control shaft 16 that is rotatably supported by the same cam bracket at a position above the drive shaft 13 and is arranged in parallel with the drive shaft 13, and is swingably supported by an eccentric cam portion 17 of the control shaft 16. A rocker arm 18 and a swing cam 20 that abuts against a tappet 19 disposed at the upper end of each intake valve 12 are provided. The eccentric cam 15 and the rocker arm 18 are linked by a link arm 25, and the rocker arm 18 and the swing cam 20 are linked by a link member 26.
[0028]
As will be described later, the drive shaft 13 is driven by a crankshaft of an engine via a timing chain or a timing belt.
[0029]
The eccentric cam 15 has a circular outer peripheral surface, the center of the outer peripheral surface is offset from the shaft center of the drive shaft 13 by a predetermined amount, and the annular portion 25a of the link arm 25 is rotatable on the outer peripheral surface. Is fitted.
[0030]
The rocker arm 18 has a substantially central portion supported by the eccentric cam portion 17, and an extension portion 25 b of the link arm 25 is linked to one end portion of the rocker arm 18, and the link member 26 is connected to the other end portion. The upper end is linked. The eccentric cam portion 17 is eccentric from the axis of the control shaft 16, and accordingly, the rocking center of the rocker arm 18 changes according to the angular position of the control shaft 16.
[0031]
The rocking cam 20 is fitted to the outer periphery of the drive shaft 13 and is rotatably supported. A lower end portion of the link member 26 is linked to an end portion 20a extending to the side. On the lower surface of the swing cam 20, a base circle surface 24a that forms a concentric arc with the drive shaft 13, a cam surface 24b extending in a predetermined curve from the base circle surface 24a to the end portion 20a, Are formed continuously, and the base circle surface 24 a and the cam surface 24 b come into contact with the upper surface of the tappet 19 according to the swing position of the swing cam 20.
[0032]
That is, the base circle surface 24a is a section where the lift amount becomes 0 as a base circle section. As shown in FIG. 3, when the swing cam 20 swings and the cam surface 24b comes into contact with the tappet 19, gradually. It will lift to. A slight ramp section is provided between the base circle section and the lift section.
[0033]
As shown in FIGS. 1 and 2, the control shaft 16 is configured to rotate within a predetermined angle range by a lift / operation angle control hydraulic actuator 31 provided at one end. The hydraulic pressure supply to the lift / operating angle control hydraulic actuator 31 is controlled by the first hydraulic control unit 32 based on a control signal from the engine control unit 33.
[0034]
The operation of the variable lift / operating angle mechanism 1 will be described. When the drive shaft 13 rotates, the link arm 25 moves up and down by the cam action of the eccentric cam 15, and the rocker arm 18 swings accordingly. The swing of the rocker arm 18 is transmitted to the swing cam 20 via the link member 26, and the swing cam 20 swings. The tappet 19 is pressed by the cam action of the swing cam 20, and the intake valve 12 is lifted.
[0035]
Here, when the angle of the control shaft 16 changes via the lift / operating angle control hydraulic actuator 31, the initial position of the rocker arm 18 changes, and consequently, the initial swing position of the swing cam 20 changes.
[0036]
For example, if the eccentric cam portion 17 is positioned upward as shown in FIG. 3A, the rocker arm 18 is positioned upward as a whole, and the end portion 20a of the swing cam 20 is relatively lifted upward. It becomes a state. That is, the initial position of the swing cam 20 is inclined in the direction in which the cam surface 24 b is separated from the tappet 19. Therefore, when the swing cam 20 swings with the rotation of the drive shaft 13, the base circle surface 24a continues to contact the tappet 19 for a long time, and the period during which the cam surface 24b contacts the tappet 19 is short. Therefore, the lift amount is reduced as a whole, and the angle range from the opening timing to the closing timing, that is, the operating angle is also reduced.
[0037]
Conversely, if the eccentric cam portion 17 is positioned downward as shown in FIG. 3B, the rocker arm 18 is positioned downward as a whole, and the end portion 20a of the swing cam 20 is pushed downward relatively. It will be in the state. That is, the initial position of the swing cam 20 is inclined in the direction in which the cam surface 24 b approaches the tappet 19. Accordingly, when the swing cam 20 swings with the rotation of the drive shaft 13, the portion that contacts the tappet 19 immediately shifts from the base circle surface 24a to the cam surface 24b. Therefore, the lift amount is increased as a whole, and the operating angle is increased.
[0038]
Since the initial position of the eccentric cam portion 17 can be continuously changed, the valve lift characteristics change continuously as shown in FIG. That is, the lift and the operating angle can be continuously expanded and contracted simultaneously. In this embodiment, the opening timing and closing timing of the intake valve 12 change substantially symmetrically as the lift and operating angle change.
[0039]
Next, as shown in FIG. 2, the phase variable mechanism 2 is configured so that the sprocket 35 provided at the front end of the drive shaft 13 and the sprocket 35 and the drive shaft 13 are relatively moved within a predetermined angle range. And a hydraulic actuator 36 for phase control that is rotated to the right. The sprocket 35 is linked to the crankshaft via a timing chain or timing belt (not shown). The hydraulic pressure supply to the phase control hydraulic actuator 36 is controlled by the second hydraulic pressure control unit 37 based on a control signal from the engine control unit 33. By the hydraulic pressure control to the phase control hydraulic actuator 36, the sprocket 35 and the drive shaft 13 are relatively rotated, and the lift center angle is retarded as shown in FIG. That is, the lift characteristic curve itself does not change, and the whole advances or retards. This change can also be obtained continuously. The phase variable mechanism 2 is not limited to a hydraulic one, and various configurations such as one using an electromagnetic actuator are possible.
[0040]
The lift / working angle variable mechanism 1 and the phase variable mechanism 2 may be controlled by providing a sensor for detecting an actual lift / working angle or phase and performing a closed loop control or depending on operating conditions. It is also possible to simply perform open loop control.
[0041]
Thus, according to the variable valve mechanism 101 combining the lift / operating angle variable mechanism 1 and the phase variable mechanism 2, both the intake valve opening timing and the intake valve closing timing can be arbitrarily controlled independently. Is possible.
[0042]
FIG. 6 is a diagram illustrating a configuration of the variable compression ratio mechanism 102.
[0043]
The crankshaft 51 includes a plurality of journal portions 52 and a crankpin portion 53, and the journal portion 52 is rotatably supported by the main bearing of the cylinder block 50. The crankpin portion 53 is eccentric from the journal portion 52 by a predetermined amount, and a lower link 54 serving as a second link is rotatably connected thereto.
[0044]
The lower link 54 is configured to be divided into a plurality of members, and the crank pin portion 53 is fitted into a substantially central connecting hole.
[0045]
The upper link 55 serving as the first link has a lower end side rotatably connected to one end of the lower link 54 by a connecting pin 56, and an upper end side rotatably connected to a piston 58 by a piston pin 57. The piston 58 receives the combustion pressure and reciprocates in the cylinder 59 of the cylinder block 50. Note that the intake valve 12 and an exhaust valve (not shown) are disposed above the cylinder 59.
[0046]
The control link 60 serving as the third link is pivotally connected at its upper end side to the other end of the lower link 54 by a connecting pin 61, and the lower end side of the lower part of the cylinder block 50 that forms part of the engine body via the control shaft 62. It is connected to the pivotable. Specifically, the control shaft 62 is rotatably supported by the engine body and has an eccentric cam portion 62a that is eccentric from the center of rotation, and the lower end portion of the control link 60 is rotatable on the eccentric cam portion 62a. Is fitted.
[0047]
The rotational position of the control shaft 62 is controlled by a compression ratio control actuator 63 using an electric motor based on a control signal from the engine control unit 33 (see FIG. 1).
[0048]
In the variable compression ratio mechanism 102 using the multi-link type piston-crank mechanism as described above, when the control shaft 62 is rotated by the compression ratio control actuator 63, the center position of the eccentric cam portion 62a, in particular, the engine. The relative position with respect to the main body changes. Thereby, the rocking | fluctuation support position of the lower end of the control link 60 changes. When the swing support position of the control link 60 changes, the stroke of the piston 58 changes, and the position of the piston 58 at the piston top dead center (TDC) becomes higher or lower as shown in FIG. This makes it possible to change the engine compression ratio. FIG. 7 representatively shows a high compression ratio state and a low compression ratio state, but the compression ratio can be continuously changed between them.
[0049]
As will be described later, this compression ratio is controlled based on a control map given in advance according to engine operating conditions. In this embodiment, as shown in FIG. 1, a knocking sensor for detecting the occurrence of knocking is used. 71 is provided in the cylinder block 50 or the like, and the compression ratio is also corrected by this knocking.
[0050]
Further, the variable compression ratio mechanism 102 using the above-described multi-link type piston-crank mechanism has a large degree of freedom in the link configuration. In particular, in the present invention, as described above, the piston thrust load in the first half of the expansion stroke is reduced. Therefore, the link configuration is such that the angle formed by the upper link 55 connected to the piston 58 and the cylinder axis m (the tilt angle of the upper link 55) θ is smaller in the first half of the expansion stroke than in the second half of the compression stroke. Is set. FIG. 8 defines the thrust load generated when a unit load is applied to the piston 58 as a piston thrust load factor, and shows the change with the crank angle. The piston thrust load factor is proportional to tan θ. As shown in the drawing, in this example, the upper link 55 is substantially vertical, that is, θ≈0 in the first half of the expansion stroke delayed from the piston top dead center position. This position substantially corresponds to the position where the combustion pressure is maximum. Before and after that, the tilt angle θ is a positive value, but its magnitude is larger in the second half of the compression stroke than in the first half of the expansion stroke with the piston top dead center as the center. That is, when comparing at two points that are the same angle away from the piston top dead center, the tilt angle θ is larger before the top dead center than after the top dead center, and the piston thrust load factor is also increased accordingly.
[0051]
FIG. 9 shows the configuration of the turbocharger 104. The turbocharger 104 has a configuration in which a turbine 82 located in the exhaust passage 81 of the internal combustion engine and a compressor 84 located in the intake passage 83 are coaxially connected, and controls the supercharging pressure according to operating conditions. Therefore, an exhaust bypass valve 85 that bypasses a part of the exhaust from the upstream side of the turbine 82 is provided. Further, a supercharging pressure sensor 86 for detecting a supercharging pressure is disposed on the downstream side of the compressor 84 in the intake passage 83, and the knocking sensor 71 is disposed on the cylinder block 50 as described above. Reference numeral 87 denotes an exhaust valve.
[0052]
Next, control of each part of the internal combustion engine configured as described above will be described.
[0053]
The control characteristics of the mechanical compression ratio ε by the variable compression ratio mechanism 102 are shown in FIG. As shown in the figure, basically, the mechanical compression ratio ε is kept low in the high load region that is the supercharging region, and the mechanical compression ratio ε is controlled to be high in the low load region that is the non-supercharging region. The mechanical compression ratio ε decreases as the load increases. However, in the illustrated example, the region where the compression ratio is 10 or less is generally a supercharging region, and the region where the compression ratio is 12 or more is generally a non-supercharging region. This compression ratio is a geometric compression ratio ε determined only by the volume change of the combustion chamber due to the stroke of the piston 58. In the present invention in combination with the variable valve mechanism 101, the final actual compression ratio is This also depends on the valve lift characteristics of the intake valve 12.
[0054]
FIG. 11 shows control of the intake valve opening / closing timing by the variable valve mechanism 101 under typical operating conditions. Since the variable valve mechanism is not provided on the exhaust valve side, the exhaust valve opening timing (EVO) is always constant. As shown in the figure, in (1) idling and (2) partial load range (R / L range), the lift operating angle Φ is advanced while the small operating angle is set. Therefore, the intake valve closing timing (IVC) is considerably earlier than the bottom dead center. As a result, the pump loss can be greatly reduced. Here, if the nominal compression ratio ε is at a normal level, the actual compression ratio decreases and combustion worsens, but as shown in FIG. 10, the compression ratio ε increases in such a low load region. Combustion deterioration is avoided.
[0055]
(3) Since it is necessary to increase the intake charging efficiency in the acceleration region, the variable valve mechanism 101 is controlled so that the intake valve closing timing approaches the bottom dead center. Therefore, the compression ratio ε gradually decreases so as to avoid the occurrence of knocking. Note that point (3) in the figure is still in the non-supercharging region, and therefore the compression ratio ε is somewhat high, and the intake valve closing timing is considerably earlier than the bottom dead center.
[0056]
In the fully open output, which is the supercharging region of (4) and (5), in order to maximize the charging efficiency, the operating angle is sufficiently expanded, the intake valve opening timing is set near the top dead center, and the intake valve The closing time is near the bottom dead center. Accordingly, since the actual compression ratio tends to increase, the compression ratio ε by the variable compression ratio mechanism 102 is further controlled to avoid knocking.
[0057]
Here, in the above example, when the actual compression ratio is reduced in the partial load range (2) or (3) or the slow acceleration range, the operating angle is set to 180 ° or less, and the intake valve closing timing is set to the intake bottom dead. Although an example in which the angle is advanced from the point (so-called early closing) has been described, the operating angle may be 180 ° or more, and the intake valve closing timing may be delayed from the bottom dead center (so-called late closing). In the valve timing charts (2) and (3) in FIG. 11, the former example and the latter example are shown together.
[0058]
As described above, the actual compression ratio becomes smaller than the actual expansion ratio by reducing the actual compression ratio in the partial load region (2) or (3) or in the slow acceleration region. That is, the relationship is “(actual compression ratio / actual expansion ratio) <1”. Here, the actual compression ratio is a compression ratio when compression is performed up to the piston top dead center on the assumption that the compression starts from the intake valve closing timing (IVC), and the actual expansion ratio is expanded from the piston top dead center. Thus, the expansion ratio assumes that expansion ends at the exhaust valve opening timing (EVO).
[0059]
FIG. 12 shows the control characteristic of the compression ratio ε by the variable compression ratio mechanism 102 with respect to the load, the change in supercharging pressure, the control characteristic of the intake valve closing timing (IVC), and the above (actual compression ratio / actual The change in the ratio of the expansion ratio is shown in comparison. As the control characteristics of the intake valve closing timing (IVC), the characteristics of both the early closing and the late closing described above are shown. Thus, in the non-supercharging region where the supercharging pressure is equal to or lower than the atmospheric pressure, the compression ratio ε is given high, and at the same time, the actual compression ratio becomes low, and “(actual compression ratio / actual expansion ratio) <1” Become. Thus, by reducing the actual compression ratio in the non-supercharging region where the mechanical compression ratio ε is high, the piston thrust load in the latter half of the compression stroke is reduced as described above.
[0060]
FIG. 13 shows the control characteristic of the compression ratio ε by the variable compression ratio mechanism 102 with respect to the load, the change in supercharging pressure, the change in the ratio of (actual compression ratio / actual expansion ratio), and the friction in the compression stroke and the expansion stroke. The loss is shown in comparison. Here, regarding friction loss, the present invention as described with reference to FIG. 8 is compared with the one having the characteristic that the tilt angle θ is substantially equal in the second half of the compression stroke and the first half of the expansion stroke with the top dead center as the center. It shows. The inclination angle in the latter half of the compression stroke is θc, the inclination angle in the first half of the expansion stroke is θe, (θc> θe) is a characteristic of the present invention, and (θc≈θe) is a general single link piston-crank mechanism. It is the characteristic of a prior art example. As shown in the figure, in the characteristic of (θc> θe) assumed by the present invention, as described above, the friction loss particularly in the expansion stroke is greatly reduced as compared with the case of (θc≈θe). On the other hand, in the compression stroke, since the tilt angle θ increases, the friction loss increases compared to the case of (θc≈θe). Here, if the actual compression ratio is not made small so that “(actual compression ratio / actual expansion ratio) <1” as in the present invention, as shown by a virtual line as a comparative example, Friction loss is even greater. In contrast, in the present invention, at the time of partial load, the compression ratio ε is set high and the expansion ratio becomes high, but the intake valve closing timing (IVC) is greatly advanced or retarded from the bottom dead center to “( By setting (actual compression ratio / actual expansion ratio) <1 ”, an increase in intake pressure and temperature due to compression during the compression stroke is mitigated, and friction loss in the compression stroke is suppressed as indicated by a solid line. In addition, since the decrease in the actual compression ratio is accompanied by a decrease in the charging efficiency, the throttle opening is widened accordingly, and the difference in the in-cylinder pressure itself is reduced by that amount, but the actual compression ratio is further decreased. The effect of pressure and temperature in the compression stroke due to is large.
[0061]
In the high load region, that is, the supercharging region, in order to increase the charging efficiency, the intake valve closing timing (IVC) is controlled so that the actual compression ratio becomes high. In this region, the mechanical compression ratio ε is greatly increased. As it decreases, the increase in friction loss is relatively small and excessive wear is avoided.
[0062]
FIG. 14 shows, for reference, friction loss and the like when the intake valve closing timing (IVC) is controlled so that the actual compression ratio is equal to the actual expansion ratio even in the non-supercharging region. As shown in the figure, in this case, the friction loss in the partial load region, particularly the friction loss in the compression stroke becomes large. This operating region is particularly frequently used as an internal combustion engine for automobiles, and thus has a great influence on practical fuel consumption.
[0063]
In the embodiment described above, the lift / operating angle variable mechanism 1 and the phase variable mechanism 2 are combined as the variable compression ratio mechanism. However, in the present invention, at least the intake valve closing timing (IVC) can be delayed. If it is sufficient, suppression of an actual compression ratio is realizable also in any one.
[0064]
FIG. 15 shows an example of variable control of the actual compression ratio when both the lift / operating angle variable mechanism 1 and the phase variable mechanism 2 are provided as in the above embodiment. FIG. 16 shows an example of variable control of the actual compression ratio when only the phase variable mechanism 2 is provided.
[Brief description of the drawings]
FIG. 1 is a configuration explanatory diagram showing an entire system of a control device according to the present invention.
FIG. 2 is a perspective view showing a variable valve mechanism in this embodiment.
FIG. 3 is an operation explanatory diagram of a lift / operating angle variable mechanism.
FIG. 4 is a characteristic diagram showing changes in lift / operating angle characteristics by a lift / operating angle variable mechanism;
FIG. 5 is a characteristic diagram showing a phase change of a valve lift characteristic by a phase variable mechanism.
FIG. 6 is a front view showing a variable compression ratio mechanism in this embodiment.
FIG. 7 is an operation explanatory diagram of a variable compression ratio mechanism.
FIG. 8 is a characteristic diagram showing characteristics of a piston thrust load factor in the variable compression ratio mechanism.
FIG. 9 is a diagram illustrating the configuration of a turbocharger.
FIG. 10 is a characteristic diagram showing compression ratio control characteristics.
FIG. 11 is a characteristic diagram showing valve lift characteristics under typical operating conditions.
FIG. 12 is a characteristic diagram showing, for example, control characteristics of intake valve closing timing with respect to a load.
FIG. 13 is a characteristic diagram showing a comparison of characteristics such as friction loss with respect to a load.
14 is a characteristic diagram similar to FIG. 13 in the case of the reference example in which the actual compression ratio = the actual expansion ratio.
FIG. 15 is a characteristic diagram showing an example of variable control of the actual compression ratio when both a lift / operating angle variable mechanism and a phase variable mechanism are provided.
FIG. 16 is a characteristic diagram showing an example of variable control of an actual compression ratio when only a phase variable mechanism is provided.
[Explanation of symbols]
DESCRIPTION OF SYMBOLS 101 ... Variable valve mechanism 102 ... Variable compression ratio mechanism 104 ... Turbocharger 1 ... Lift / operating angle variable mechanism 2 ... Phase variable mechanism

Claims (8)

A variable compression ratio mechanism in which a piston top dead center is changed by moving a part of the link structure using a multi-link type piston-crank mechanism; a variable valve mechanism capable of variably controlling at least the closing timing of the intake valve; An internal combustion engine control device for controlling the compression ratio and the intake valve closing timing according to operating conditions,
The variable compression ratio mechanism has a link configuration so that the angle θ formed between the first link connected to the piston and the cylinder axis is smaller in the first half of the expansion stroke than in the second half of the compression stroke and is minimized in the first half of the expansion stroke. Is set,
The intake valve so that the piston thrust load in the latter half of the compression stroke when the angle θ is relatively large at the high compression ratio is equal to or smaller than the piston thrust load in the first half of the expansion stroke where the angle θ is minimized. by controlling the closing timing control apparatus for an internal combustion engine, characterized in that the actual compression ratio determined by the intake valve closing timing is smaller than the actual expansion ratio determined by the open exhaust valve timing. Equipped with a turbocharger,
In the supercharging region where the cylinder pressure during the intake stroke is higher than atmospheric pressure, the compression ratio is controlled to be low,
In the non-supercharging region where the cylinder pressure in the intake stroke is lower than atmospheric pressure, the compression ratio is controlled to be high, and the intake valve closing timing is controlled so that the actual compression ratio is smaller than the actual expansion ratio. The control apparatus for an internal combustion engine according to claim 1. The control apparatus for an internal combustion engine according to claim 2, wherein the compression ratio is controlled to be lower as the supercharging pressure becomes higher. The intake valve operating angle is set to 180 ° or less in crank angle in order to reduce the actual compression ratio, and the intake valve closing timing is advanced from the intake bottom dead center. A control device for an internal combustion engine according to claim 1. The intake valve operating angle is set to 180 ° or more in crank angle in order to reduce the actual compression ratio, and the intake valve closing timing is retarded from the intake bottom dead center. A control device for an internal combustion engine according to claim 1. The variable valve mechanism includes a lift / operating angle variable mechanism that can continuously increase / decrease the lift / operating angle of the intake valve, and a phase variable mechanism that delays the phase of the central angle of the lift of the intake valve. The control apparatus for an internal combustion engine according to any one of claims 1 to 5, wherein the control apparatus is configured by any one or a combination of both. The variable lift / operating angle mechanism includes an eccentric cam that is rotationally driven by a drive shaft, a link arm that is fitted to the outer periphery of the eccentric cam so as to be relatively rotatable, and an eccentric cam portion that is provided in parallel with the drive shaft. A pivotable control shaft comprising: a rocker arm that is rotatably mounted on an eccentric cam portion of the control shaft and is pivoted by the link arm; and is rotatably supported by the drive shaft, and A rocking cam connected to the rocker arm via a link, and rocking with the rocker arm to press the tappet of the intake valve, thereby changing the rotational position of the eccentric cam portion of the control shaft. 7. The control apparatus for an internal combustion engine according to claim 6, wherein the lift / operating angle of the intake valve is increased or decreased at the same time. The variable compression ratio mechanism includes a first link coupled to a piston via a piston pin, and is pivotably coupled to the first link and is rotatably coupled to a crankpin portion of a crankshaft. A fulcrum position of the third link with respect to the engine body, comprising: a second link; and a third link pivotably connected to the second link and supported by the engine body. The control device for an internal combustion engine according to claim 1, wherein the compression ratio is variably controlled by changing.

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