JP2016017489A - Internal combustion engine control unit - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide an internal combustion engine control unit capable of achieving improvements in fuel economy, exhaust emission performance, and the like by making both a mechanical compression ratio and a mechanical expansion ratio different depending on an engine operation state.SOLUTION: An internal combustion engine control unit comprises: an upper link 7 coupled to a piston 2 via a piston pin 3; a lower link 10 rockably coupled to the upper link 7 via a first coupling pin 8, and rotatably coupled to a crankpin 9 of a crankshaft 4; and a control link 14 rockably coupled to the lower link 10 via a second coupling pin 11, and rotatably coupled to an eccentric cam portion 13 provided on a control shaft 12, the control shaft 12 being configured to rotate at an angular velocity that is half of an angular velocity of the crankshaft 4, and the internal combustion engine control unit comprising a phase change mechanism 6 of a vane type capable of changing relative rotation phases of the crankshaft 4 and the control shaft 12 relative to each other.SELECTED DRAWING: Figure 1

Description

  The present invention relates to a control device for an internal combustion engine capable of changing a mechanical compression ratio and a mechanical expansion ratio by changing relative stroke positions of a compression stroke and an expansion stroke of a piston, respectively.

  As a conventional control device for an internal combustion engine, a control device described in Patent Document 1 including a variable compression ratio mechanism is known.

  That is, the variable compression ratio mechanism uses a multi-link type piston-crank mechanism. As shown in FIG. 8 of Patent Document 1, the left-side diagram shows the piston compression top dead in the high mechanical compression ratio control. At the point position (piston position slightly high), the figure on the right is the piston compression top dead center position (piston position slightly low) in the low mechanical compression ratio control.

  Here, regarding the correlation between the mechanical compression ratio C and the mechanical expansion ratio E, first, for example, the case of the high mechanical compression ratio control shown in the left diagram will be considered.

  Assuming that the cylinder internal volume at the piston compression top dead center position is VO, the piston intake bottom dead center is reached in the vicinity of the crankshaft rotating counterclockwise by about 180 ° from the piston compression top dead center position. If the internal volume is VC, VC ÷ VO is the mechanical compression ratio C.

  On the other hand, when the crankshaft is rotated approximately 180 ° clockwise from the piston compression top dead center position, the piston expansion bottom dead center is reached. When the cylinder volume at this time is VE, VE ÷ VO becomes the mechanical expansion ratio E. .

JP 2002-276446 A

  However, since the conventional variable compression ratio mechanism is a mechanism having one cycle (one cycle) at a crank angle of 360 °, the piston intake bottom dead center and the piston expansion bottom dead center are at the same piston position. Therefore, VC = VE, that is, mechanical compression ratio C = mechanical expansion ratio E.

  Similarly, in the case of the low mechanical compression ratio control, the mechanical compression ratio C = the mechanical expansion ratio E. That is, in this conventional variable compression ratio mechanism, although the mechanical compression ratio C can be changed, the mechanical expansion ratio E also changes to the same value as the mechanical compression ratio C.

  For this reason, various inconveniences in engine performance occur, for example, when it is desired to increase the expansion work by sufficiently increasing the mechanical expansion ratio E in the low load range (after engine warm-up) that is a normal operation and to improve the fuel efficiency. As a result, the mechanical compression ratio C becomes excessively high, so that the in-cylinder gas temperature at the compression top dead center is excessively increased, thereby increasing the cooling loss, and the fuel consumption cannot be sufficiently improved. There was a problem.

  Or, for example, if the mechanical compression ratio C is lowered in order to improve the knock resistance at the time of high engine load after completion of warming up, the mechanical expansion ratio E also decreases accordingly. As a result, the fuel consumption deteriorates and the torque cannot be sufficiently increased. In this case, the exhaust work at the time of high engine load increases due to the reduction of the expansion work due to the decrease in the mechanical expansion ratio E, and the thermal deterioration of the catalyst is promoted, so that the exhaust emission increases with time. The problem of end up occurring.

  Also, for example, at the time of cold start, when the mechanical expansion ratio E is lowered to reduce the expansion work of the engine, and the exhaust gas temperature is raised correspondingly to improve the exhaust emission conversion (purification) efficiency in the catalyst, Following this, the mechanical compression ratio C decreases.

  That is, due to the decrease in the mechanical compression ratio C, the in-cylinder gas temperature at the compression top dead center cannot be sufficiently increased during cold operation, the combustion deteriorates, and exhaust emissions (catalysts) exhausted from the engine itself are reduced. As a result, there has been a problem that exhaust emission discharged from the tail pipe (downstream of the catalyst) to the atmosphere after passing through the catalyst cannot be sufficiently reduced.

  The present invention has been devised in view of the above-described conventional technical problems, and the mechanical compression ratio and the mechanical expansion ratio can be made different according to the engine operating state. An object of the present invention is to provide a control device for an internal combustion engine that can improve various performances of the engine such as improvement.

  The present invention has been devised in view of the conventional technical problem, and the piston position variable mechanism includes a first link connected to the piston via a piston pin, and a first connection to the first link. A second link that is swingably connected via a pin and is rotatably connected to a crankpin of the crankshaft; and a second link that is swingably connected to the second link via a second connection pin and is controlled A third link rotatably connected to an eccentric cam provided on the control shaft, and configured to rotate the control shaft at an angular speed half that of the crankshaft; and the crankshaft and the control -It has a phase change mechanism that makes it possible to change the relative rotational phase with the shaft.

  The piston position variable mechanism is characterized in that the mechanical compression ratio and the mechanical expansion ratio can be changed, and both can be changed differently.

  According to the present invention, since the mechanical compression ratio and the mechanical expansion ratio can be made different according to the engine operating state, various engine performances such as improvement in fuel consumption and exhaust emission performance can be improved.

1 is an overall schematic diagram (rear view) showing a first embodiment in which a control device according to the present invention is applied to a single cylinder internal combustion engine. It is a principal part side view which cuts and shows a part of control apparatus of this embodiment. It is the front view (front view) which removed the front cover of the phase change mechanism with which this embodiment is provided, A shows the most retarded angle control state, B shows the most advanced angle state. The operation of the control shaft phase conversion by the phase changing mechanism provided in this embodiment is shown (rear view), and at the crankshaft rotation angle (X = 360 °) with the crankpin near the compression top dead center facing directly upward, A is the state in which the eccentric rotation phase of the control shaft is controlled to α1 (eg 137 °), B is α2 (eg 180 °), C is α3 (eg 222 °), and D is controlled to α4 (eg 240 °). Show. It is a characteristic view which shows the height position change of a piston in relation to the rotation angle of the crankshaft in this embodiment. FIG. 4 is an operation explanatory diagram (rear view) of the piston position variable mechanism in the present embodiment, where A to D indicate the piston position when the vane rotor is in the most retarded state (control phase α4), and A is the intake top dead center. Position, B is an intake bottom dead center position, C compression top dead center position, D is an expansion bottom dead center position, and E to H are piston positions when the vane rotor is in the most advanced state (control phase α3). , E indicates an intake top dead center position, F indicates an intake bottom dead center position, G indicates a compression top dead center position, and H indicates an expansion bottom dead center position. It is a characteristic view which shows the height position change of a piston in relation to the rotation angle of the crankshaft in 2nd Embodiment. FIG. 4 is an operation explanatory diagram (rear view) of the piston position variable mechanism in the present embodiment, where A to D indicate the piston position when the vane rotor is in the most retarded state (control phase α3), and A is the intake top dead center. Position, B is an intake bottom dead center position, C compression top dead center position, D is an expansion bottom dead center position, and E to H are piston positions when the vane rotor is in the most advanced state (control phase α2). , E indicates an intake top dead center position, F indicates an intake bottom dead center position, G indicates a compression top dead center position, and H indicates an expansion bottom dead center position. It is a characteristic view which shows the height position change of a piston in relation to the rotation angle of the crankshaft in 3rd Embodiment. FIG. 5 is an operation explanatory diagram (rear view) of the piston position variable mechanism in the present embodiment, where A to D indicate the piston position when the vane rotor is in the most advanced state (control phase α1), and A is the intake top dead center. Position, B is an intake bottom dead center position, C compression top dead center position, and D is an expansion bottom dead center position. It is a whole schematic diagram which shows the link mechanism of the control apparatus of 4th Embodiment.

  Embodiments of an internal combustion engine control apparatus according to the present invention will be described below with reference to the drawings.

In this embodiment, the present invention is applied to a single-cylinder internal combustion engine 01 of a 4-stroke combustion cycle gasoline specification. The feature is that the mechanism 1 is provided.
[First Embodiment]
FIG. 1 (rear view) and FIG. 2 show a first embodiment of the present invention. An internal combustion engine 01 includes a piston 2 that reciprocates in a vertical direction along a cylinder bore 03 formed in a cylinder block 02, and the piston 2. , And a crankshaft 4 that is rotationally driven via a link mechanism 5 described later of the piston position variable mechanism 1. The space defined between the combustion chamber boundary line indicated by the alternate long and short dash line on the crown surface 2a of the piston 2 in FIG.

  The piston position variable mechanism 1 includes the link mechanism 5 composed of a plurality of links, a phase change mechanism 6 that changes the posture of the link mechanism 5, and the like.

  The link mechanism 5 is connected to the piston 1 via a piston pin 3 and is connected to the upper link 7 via a first connection pin 8. The upper link 7 is a first link connected to the piston 1 via a piston pin 3. A lower link 10 which is a second link rotatably connected to the crankpin 9 of the crankshaft 4, and is swingably connected to the lower link 10 via a second connecting pin 11 and is connected to the control shaft 12. A control link 14 is a third link that is rotatably connected to the eccentric cam portion 13.

  Further, as shown in FIGS. 1 and 2, a small-diameter first gear gear 15 that is a drive rotating body is fixed to the front end portion of the crankshaft 4, while the front end portion side of the control shaft 12 is driven. A large-diameter second gear gear 16 that is a rotating body is provided, and the first gear gear 15 and the second gear gear 16 mesh with each other, and the rotational force of the crankshaft 4 is applied to the control shaft 12 via the phase change mechanism 6. It is to be transmitted.

  The outer diameter of the first gear gear 15 is about half of the outer diameter of the second gear gear 16, and therefore the rotational speed of the crankshaft 4 is the same as that of the first gear gear 15 and the second gear gear 15. The gear 16 is transmitted to the control shaft 12 after being decelerated to a half angular velocity due to the difference in outer diameter of the gear 16.

  The phase of the control shaft 12 with respect to the second gear gear 16 is changed by the phase changing mechanism 6, that is, the relative rotational phase with respect to the crankshaft 4 is changed.

  The crankshaft 4 and the control shaft 12 are rotatably supported by two common front and rear bearings 17 and 18 provided in the cylinder block. Further, the eccentric cam portion 13 is rotatably connected to a large diameter portion formed at the lower end portion of the control link 14 via a needle bearing 19.

  The phase change mechanism 6 will be briefly described because it has the same structure as the hydraulic (vane type) phase change mechanism described in Japanese Patent Application Laid-Open No. 2012-225287 filed by the present applicant.

  That is, as shown in FIGS. 2 and 3A, B (front view), the phase changing mechanism 6 is housed in a housing 20 to which the second gear gear 16 is fixed, and accommodated in the housing 20 so as to be relatively rotatable. And a vane rotor 21 fixed to one end of the control shaft 12 and a hydraulic circuit 22 for rotating the vane rotor 21 forward and backward by hydraulic pressure.

  In the housing 20, a front end opening of a cylindrical housing body 20 a is closed by a disk-shaped front cover 23, and a rear end opening is closed by a disk-shaped rear cover 24. Further, at a position of about 90 ° in the circumferential direction of the inner peripheral surface of the housing main body 20a, shoes 20b that are four partition walls project from the inner side.

  The rear cover 24 is integrally provided at the center position of the second gear gear 16, and the outer peripheral portion thereof is fastened and fixed to the housing body 20 a and the front cover 23 by four bolts 25. In addition, a large-diameter bearing hole 24 a that is supported on the outer periphery of the cylindrical portion of the vane rotor 21 is formed in the center of the rear cover 24 in the axial direction.

  The vane rotor 21 includes a cylindrical rotor 26 having a bolt insertion hole in the center, and four vanes 27 that are integrally provided at approximately 90 ° in the circumferential direction of the outer peripheral surface of the rotor 26.

  In the rotor 26, a small-diameter cylindrical portion 26a on the front end side is rotatably supported by the central support hole of the front cover 23, while a small-diameter cylindrical portion 26b on the rear end side rotates in the bearing hole 24a of the rear cover 24. It is supported freely.

  The vane rotor 21 is fixed to the front end portion of the control shaft 12 from the axial direction by a fixing bolt 28 inserted through the bolt insertion hole of the rotor 26 from the axial direction.

  Each vane 27 is disposed between the shoes 20b, and a seal member that slides in contact with the inner peripheral surface of the housing body 20a in an elongated holding groove formed in the axial direction of each outer surface. Each of the leaf springs that press in the direction of the inner peripheral surface of the housing body is fitted and held.

  Further, four advance chambers 40 and retard chambers 41 are respectively formed between both sides of each vane 27 and both sides of each shoe 20b.

  As shown in FIG. 2, the hydraulic circuit 22 includes a first hydraulic passage 28 that supplies and discharges hydraulic oil pressure to and from each advance chamber 40, and hydraulic oil pressure to each retard chamber 41. There are two systems of hydraulic passages, a second hydraulic passage 29 for supplying and discharging gas, and a supply passage 30 and a drain passage 31 are respectively connected to both of the hydraulic passages 28 and 29 via an electromagnetic switching valve 32 for passage switching. Connected. The supply passage 30 is provided with a one-way oil pump 34 that pumps the oil in the oil pan 33, while the downstream end of the drain passage 31 communicates with the oil pan 33.

  The first and second hydraulic passages 28 and 29 are formed inside a passage constituting portion provided on the front cover 23 side, and each one end portion is a small diameter cylindrical portion 26a of the rotor 26 of the passage constituting portion. The other end portion is connected to the electromagnetic switching valve 32 while communicating with the rotor 26 via a cylindrical portion 35 inserted and disposed in an internal support hole.

  The first hydraulic passage 28 includes four branch passages (not shown) communicating with the advance chambers 40, while the second hydraulic passage 29 is a second oil passage communicating with the retard chambers 41. And.

  The electromagnetic switching valve 32 is a four-port, three-position type, and an internal valve body is configured to relatively switch and control each of the hydraulic passages 28, 29, the supply passage 30, and the drain passage 31, Switching operation is performed by a control signal from the control unit 36.

  The vane rotor 21 (control shaft 12) is supplied to the crankshaft 4 by selectively supplying hydraulic oil to each advance chamber 40 and each retard chamber 41 by the switching operation of the electromagnetic switching valve 32. The relative rotational phase is changed.

  In each retarding angle chamber 41, four coil springs 42 that constantly bias the vane rotor 21 in the retarding direction are mounted.

  4A to 4D (rear views) show a case where the relative rotational phase between the second gear gear 16 and the control shaft 12 is changed. In this figure, the first and second gear gears 15 and 16 are omitted.

  In this embodiment, the relative rotational phase can be changed by the relative rotational phase conversion control by the phase changing mechanism 6 described above, but the second gear gear 16 and the control shaft 12 (eccentric cam portion 13). It can also be performed by relatively changing the mounting relationship.

  4, the crankshaft 4 is rotated clockwise without changing the relative phase between the second gear gear 16 and the control shaft 12 shown in FIG. (Crank angle X = 0 °; near intake (exhaust) top dead center), and one more turn to show the posture at the position where the crankpin 9 faces again upward (X = 360 °; near compression top dead center) . At this time, since the position (height) of the piston 2 is near the compression top dead center, it is at a high position. At this time, for example, as shown in FIG. 3A, the eccentric direction of the eccentric cam portion 13 is a position that is retarded in the clockwise direction by α1 (for example, 137 °) from the directly above direction.

That is, the rotation direction of the eccentric cam portion 13 in FIG. 4 (rear view) is the counterclockwise direction opposite to the crankshaft, so in the case shown in FIG. 4A, it is retarded by α1 from the direct upward direction. In such a case, it will be referred to as a control phase α1.
FIG. 4B shows a position where the phase of the control shaft 12 (eccentric cam portion 13) is further retarded to α2 (for example, 180 °) on the retard side, that is, the eccentric direction of the eccentric cam portion 13 with respect to FIG. 4A. Near the bottom. (Control phase α2)
4C, the control phase α3 (eg, 222 °), the control phase α4 (eg, 240 °) in FIG. 4D, and the phase of the control shaft 12 (eccentric cam portion 13) are retarded clockwise. It is in the position where it went.

  Here, for example, the operation of the phase changing mechanism 6 capable of converting between the control phase α3 shown in FIG. 4C and the control phase α4 shown in FIG. 4D will be considered based on FIGS. 3A and 3B.

  Since FIG. 3 is a front view, the rotation direction of the second gear gear 16 is clockwise in FIG. 3A shows the most retarded angle position (corresponding to the control phase α4) of the vane rotor 21 of the phase change mechanism 6, and FIG. 3B shows the most advanced angle position (corresponding to the control phase α3). Both sides of the widest vane 27 (27a) at both angular positions come into contact with one side surface and the other side surface of each adjacent shoe 20b and are regulated by stoppers (retard side stopper, advance side stopper). ing.

Here, the vane rotor 21 is mechanically stabilized in the vicinity of the most retarded position as shown in FIG. 3A by the spring force of each coil spring 42. (In other words, the default position is the most retarded position)
If the phase conversion angle αT of the phase change mechanism 6 is αT = α4-α3, for example, 18 ° (= 240 ° -222 °), a desired conversion angle αT (18 °) in the conversion of α3⇔α4 is realized. it can.

  Further, when the attachment phase between the vane rotor 21 and the control shaft 12 is set so that the aforementioned most retarded angle position (default position) coincides with α4 shown in FIG. 4D, the desired phase conversion of FIGS. (Α3⇔α4) can be realized.

  FIG. 5 shows the piston position change characteristic. Here, when the crank angle X is 0 °, the crank pin 9 is located directly above, and in this vicinity, the intake (or exhaust) top dead center of the piston 2 is reached.

  When the crank angle X starts to rotate clockwise from 0 °, the exhaust valve (not shown) is completely closed, the intake valve (not shown) starts to open, and fresh air (mixture) is started from the intake port (not shown). ). Next, the piston intake bottom dead center is reached in the vicinity of the crank angle X of 180 °, and the intake valve is closed in this vicinity. Here, the intake stroke is from the intake top dead center to the intake bottom dead center.

  Further, when the crankshaft 4 rotates, the intake valve is completely closed and the in-cylinder air-fuel mixture is compressed, near the position where the crank angle X becomes 360 ° (the crankpin 9 is again directly above). , Piston compression top dead center. Here, the range from the intake bottom dead center to the compression top dead center is referred to as a compression stroke.

  Thereafter, spark ignition (or compression ignition) is performed and combustion is started. The combustion pressure pushes down the piston 2, and an expansion bottom dead center is reached when the crank angle X is around 540 °. Here, the process from the compression top dead center to the expansion bottom dead center is referred to as an expansion stroke.

Near this expansion bottom dead center, the exhaust valve starts to open, and when the piston 2 rises again, the combustion gas (exhaust gas) is discharged from the exhaust port (not shown), and again near the top dead center of the exhaust (intake) That is, the crank angle X returns to a position where the crank angle X is 720 ° (0 °) (the crank pin 9 is located directly above). (Here, the range from the expansion bottom dead center to the exhaust (intake) top dead center is called the exhaust stroke.)
As described above, the operation as a four-cycle engine is performed, and the operation is periodic with a crank angle (X) of 720 ° as one cycle.

  In FIG. 5, the solid line indicates the piston position characteristic (α3 characteristic) at the control phase α3 in FIG. 4C, and the broken line indicates the piston position characteristic (α4 characteristic) at the control phase α4 in FIG. 4D. In both characteristics, the piston position at the compression top dead center is substantially the same (Y0), and the intake bottom dead center position is different in both characteristics. That is, the cylinder internal volume V at the compression top dead center is determined by the Y0 for both characteristics and is substantially the same (V0).

  This V0 is the volume surrounded by the shape of the combustion chamber on the cylinder head side, the shape of the crown surface 2a of the piston 2, the inner diameter of the cylinder block 02, the inner diameter of the head gasket, not shown, and the like at the compression top dead center. That is, it becomes the volume of the gas (air mixture) at the compression top dead center.

  In the α3 characteristic shown in FIG. 5, the piston position at the intake bottom dead center is YC3, the length from that to the compression top dead center (compression stroke) is LC3, and the piston position at the expansion bottom dead center is YE3. The length (expansion stroke) from the compression top dead center to this is LE3.

  Here, C3 which is the mechanical compression ratio C in the α3 characteristic and E3 which is the mechanical expansion ratio E will be considered.

  When the area of the bore (cylinder inner diameter) is S, the cylinder internal volume VC3 at the intake bottom dead center is VC3 = V0 + S × LC3.

  Therefore, the mechanical compression ratio C3 = VC3 ÷ V0 = (V0 + S × LC3) ÷ V0.

  On the other hand, mechanical expansion ratio E3 = VE3 ÷ V0 = (V0 + S × LE3) ÷ V0.

Here, VE3 is the cylinder volume at the expansion bottom dead center.
In the case of the α3 characteristic, as shown in FIG. 5, since LC3≈LE3, mechanical compression ratio C3≈mechanical expansion ratio E3. Here, the relative ratio D = mechanical expansion ratio E ÷ mechanical compression ratio C is defined.

  The relative ratio D3 in the α3 characteristic is E3 ÷ C3≈1, and the mechanical expansion ratio E and the mechanical compression ratio C are substantially the same and almost standard characteristics.

  That is, the α3 characteristic is close to a normal piston position change characteristic (E = C, D = 1) of a general engine.

  Next, C4 which is the mechanical compression ratio C in the α4 characteristic and E4 which is the mechanical expansion ratio E will be considered.

  Similar to the α3 characteristic, mechanical compression ratio C4 = VC4 ÷ V0 = (V0 + S × LC4) ÷ V0, and mechanical expansion ratio E4 = VE4 ÷ V0 = (V0 + S × LE4) ÷ V0.

  Here, in the case of the α4 characteristic, as shown in FIG. 5, since LC4> LE4, the mechanical compression ratio C4> the mechanical expansion ratio E4. That is, the relative ratio D4 = LE4 ÷ LC4 <1, which means that the mechanical compression ratio is relatively larger than the mechanical expansion ratio. Further, in comparison with the α3 characteristic, C4> C3 and the mechanical compression ratio are large, and E4 <E3 and the mechanical expansion ratio are small.

6A to 6D, when the crank angle at the control phase α4 (the most retarded angle default position of the vane rotor 21, for example, 240 °) is changed, the control phase α3 (the most advanced angle position of the vane rotor 21 such as 222 for example) is changed to FIGS. FIG. 9 shows respective mechanism posture change diagrams when the crank angle at °) is changed. (All rear view)
Here, A and E in FIG. 6 are postures at the top dead center of intake (exhaust), B and F are postures at the bottom dead center of intake, C and G are postures at the top dead center of compression, and D and H are expansions. Each posture is shown at the bottom dead center.

  In the case of the α3 characteristic shown in FIGS. 6E to H, since LC3≈LE3 as described above, the mechanical compression ratio C3≈mechanical expansion ratio E3 (relative ratio D3≈1).

  In the case of the α4 characteristic shown in FIGS. 6A to 6D, since LC4> LE4 as described above, mechanical compression ratio C4> mechanical expansion ratio E4 (relative ratio D4 <1). 6A to 6H, LC4> LC3 and LE4 <LE3 as described above.

  The reason why such a piston position change characteristic is considered will be considered. Looking at the eccentric rotation direction αC of the eccentric cam portion 13 at the intake bottom dead center, αC4 in the α4 characteristic shown in FIG. 6B is clockwise (retarded direction) with respect to αC3 in the α3 characteristic shown in FIG. 6F. In other words, the eccentric circle center of the eccentric cam portion 13 has moved to the upper right relative to the α3 characteristic, so that the control link 14 moves the second connecting pin 11 to the upper right. The lower link 10 is pushed up and rotated clockwise with the crank pin 9 as a fulcrum, whereby the position of the first connecting pin 8 is lowered, and the piston 2 is pulled down by the upper link 7. Thereby, LC4> LC3.

  On the other hand, when looking at the eccentric rotation direction αE of the eccentric cam portion 13 at the expansion bottom dead center, αE4 in the α4 characteristic shown in FIG. 6D is similarly clockwise (αE3 in the α3 characteristic shown in FIG. 6H). In other words, the center of the eccentric circle has moved relatively downward, so that the control link 14 pulls the second connecting pin 11 downward to the left and lower link 10 is connected to the crank pin 9. And the position of the first connecting pin 8 is raised, and the piston 2 is pushed upward by the upper link 7. Thus, LE4 <LE3.

  That is, the difference in the piston position change characteristic between the control phase α3 and the control phase α4 shown in FIG. 5 is generated by the difference in the link posture due to the difference in the eccentric phase of the eccentric cam portion 13 shown in FIG.

  On the other hand, when looking at the compression top dead center position, the position of the piston 2 in the control phase α3 and the control phase α4 is substantially the same as described above. This is due to the following reason. That is, as shown in the compression top dead center posture shown in FIGS. 6C and 6G, the crank pin 9, the first connecting pin 8, and the piston pin 3 are arranged substantially in a straight line in both the α3 characteristic and the α4 characteristic. This is because even if the first connecting pin 8 is rotated by the rotation of the lower link 10, the position change of the piston pin 2 is slightly suppressed.

  For this reason, the compression top dead center piston position of α3 characteristic (Y03 in FIG. 5) and the compression top dead center piston position of α4 characteristic (Y04 in FIG. 5) are substantially the same position, and this is Y0 described above. is there.

However, if there is a significant difference between Y03 and Y04, the in-cylinder volumes are used as V03 and V04, respectively, instead of the above-mentioned V0, and the respective mechanical compression ratios C3 and C4, the respective mechanical expansion ratios E3 and E4, What is necessary is just to obtain | require each total contrast D3 and D4.
[Performance effect of this embodiment]
When the engine is stopped, the vane rotor 21 of the phase change mechanism 6 is pressed and stabilized by the spring force of each coil spring 42 at the most retarded position (counterclockwise direction) shown in FIG. 3A (default position). Is the aforementioned α4.

Therefore, at the time of cold start, the characteristic of α4 which is the most retarded angle position of the vane rotor 21 (broken line in FIG. 5) is set in advance. can get. (Also, this position can be maintained even when the electrical system of the electromagnetic switching valve 32 of the phase change mechanism 6 has a failure such as disconnection, so that the exhaust emission reduction effect described above can be obtained even in that case, so-called mechanical fail-safe effect. Also have.)
That is, regarding the exhaust emission reduction effect due to this characteristic, first, since the mechanical expansion ratio E4 is small, the temperature of exhaust gas discharged from the internal combustion engine increases as the expansion work is reduced. The catalyst warm-up is promoted to improve the emission conversion rate. (A)
On the other hand, second, since the mechanical compression ratio C4 is increased, the in-cylinder gas temperature at the compression top dead center is increased, which improves the combustion failure that becomes a problem during cold operation, and therefore from the internal combustion engine itself. Emissions can be reduced. (I)
By the above E4 small, C4 large, and D4 small (= E4 ÷ C4), the amount of exhaust emission released to the atmosphere from the tail pipe downstream of the catalyst is reduced as a synergistic effect of the effects of (a) and (b) it can.

  Here, the relative ratio D4 is a small value less than 1, which means that the smaller this is, the smaller the mechanical expansion ratio and the larger the mechanical compression ratio. It can be regarded as an index indicating good performance.

  By the way, when the warm-up of the engine is completed, the fuel consumption deteriorates in the state where C4, E4, and D4 remain unchanged. Because the mechanical expansion ratio E4 is low, the expansion work by the piston 2 is reduced (c), and the mechanical compression ratio C4 is high, so that the compression top dead center temperature becomes excessively high after warm-up, so-called cooling loss occurs. The fuel consumption increases due to the loss due to the above (c) and (d).

  Further, if the engine operating state is a high load, abnormal combustion such as knocking and pre-ignition is further induced, and the fuel consumption is further deteriorated and the torque is also reduced.

  Therefore, after the warm-up, the vane rotor 21 is converted to the most advanced position by the control hydraulic pressure from the electromagnetic switching valve 32 of the phase change mechanism 6 and switched to the characteristic of α3 (solid line in FIG. 5).

  As a result, the standard mechanical expansion ratio E3 and the standard mechanical compression ratio C3 are restored, and the relative ratio D3 is almost 1, which is equivalent to the normal piston position change characteristic. As a result, it is possible to suppress the deterioration of fuel consumption and the induction of abnormal combustion.

  If the engine temperature is intermediate between the cold engine and the warm-up completion, the vane rotor 21 is changed to the retard side as the temperature becomes lower (closer to α4), and the phase changes as the temperature becomes higher. By changing to the advance side by the mechanism 6 (closer to α3), the exhaust emission performance and the fuel consumption performance can be optimally balanced for each change in temperature. For example, fuel consumption deterioration can be suppressed as much as possible while suppressing the emission to a sufficiently low predetermined value.

  As described above, the position change characteristic of the piston 2 is cyclically operated with a crank angle of 720 ° as a cycle, and the top dead center is 2 degrees where the crank angle is around 0 ° and around 360 °. Appears.

The top dead center near the crank angle of 360 ° (Y0 mentioned above) is the compression top dead center where the intake valve and the exhaust valve are completely closed, and the top dead center near the crank angle 0 ° (Y '03, Y'04) is an intake (exhaust) top dead center which is another top dead center where the exhaust valve is closed and the intake valve operation is started.
The intake (exhaust) top dead center position (Y′03, Y′04) is lower than the compression top dead center (Y0). As shown in the intake top dead center posture in FIGS. 6A and 6E, the crank pin 9, the first connecting pin 8, and the piston pin 3 are not arranged in a straight line but are bent in a reverse letter shape for both α3 and α4. Due to this arrangement, the position of the piston 2 is lower than the above-mentioned Y0, and the top of the piston 2 is caused by the phase difference of the control shaft 12 between α3 and α4, that is, by the difference in the angle of the inverted letter. There is a significant difference between Y'03 and Y'04 at the dead center position. (Fig. 5)
Here, assuming that the operation timing of the intake / exhaust valve is shifted by 360 ° in the crank angle, the compression top dead center and the intake (exhaust) top dead center are interchanged. The position change characteristic of the piston 2 at the top dead center and the position change characteristic of the piston 2 at the intake (exhaust) top dead center are interchanged. Even in such a case, the present invention as described above is used. The basic effect can be obtained.

  On the other hand, when this setting is not shifted, that is, in the first embodiment (FIGS. 1 to 6) itself described above, the following special effects can be obtained.

  That is, since the piston top dead center position of Y0 is high at the compression top dead center, the mechanical compression ratio C and the mechanical expansion ratio E can be set large, so that the effect of the present invention can be sufficiently enhanced.

  Moreover, even if such a high piston position is set, the intake / exhaust valve does not operate at the compression top dead center and the closed state continues, so that the problem of interference between the piston 2 and the intake / exhaust valve does not occur.

  Also, at the intake (exhaust) top dead center, the exhaust valve closing operation and the intake valve opening operation are performed in this vicinity. Therefore, if the piston position is as high as Y0, interference between the intake / exhaust valve and the piston 2 will occur. As described above, since the intake (exhaust) top dead center position (Y′03, Y′04) of the piston 2 is lower than the compression top dead center position Y0 as described above, the interference can be avoided. It is.

Further, in the present embodiment, as shown in FIG. 2, the phase changing mechanism 6 is installed in the second gear gear 16 having a large diameter on the control shaft 12 side to be decelerated. When it is assumed that the phase change mechanism 6 is installed in the one gear gear 15, the outer diameter of the vane rotor 21 can be set appropriately large, and the conversion power of the vane rotor 21 by the phase change mechanism 6 can be increased. It is possible to improve the conversion responsiveness and increase the load resistance.
[Second Embodiment]
7 to 8 show a second embodiment of the present invention. Compared to the first embodiment, the conversion angle of the vane rotor 21 and the relative rotational phase of the vane rotor 21 and the control shaft 12 are changed, and FIGS. The control phase α2 and the control phase α3 shown in FIG.

  The conversion angle αT in FIG. 3 is α3−α2 (for example, 222 ° −180 ° = 42 °), and similarly, a coil spring 42 for urging the vane rotor 21 toward the retarded angle side is provided.

  By the way, in the second embodiment, the conversion angle of the vane rotor 21 is larger than that of the first embodiment, but by removing the vicinity of the stopper portion which is the protrusion of the housing body 20a and the side surface portion of the widened vane 27a. It only has to respond. Alternatively, even if the number of vanes 27 is reduced from 4 to 3, the conversion angle can be increased.

  Then, the mounting phase of the vane rotor 21 and the control shaft 12 is set so that the most retarded angle position (default position) where the retarded angle side stopper and the widening vane 27a contact each other coincides with the position α3 shown in FIG. 4C. It ’s fine.

  However, if the number of vanes 27 is reduced as the conversion angle of the vane rotor 21 is increased in this way, the conversion power (driving force) due to the hydraulic pressure of the phase change mechanism 6 is reduced, and there is a concern that the conversion response may deteriorate.

  However, as described above, since the phase changing mechanism 6 is provided toward the large-diameter second gear gear 16 on the speed reducing side, the outer diameter of the vane 27 and the like can be set appropriately large. Thus, the conversion power of the vane rotor 21 by the phase change mechanism 6 can be ensured, and the deterioration of the conversion response and the phase holding ability of the vane rotor 21 can be suppressed.

  FIG. 7 shows the piston position change characteristic, and the solid line shows the same characteristic (α3 characteristic) as the control phase α3 of FIG. 4C of the first embodiment, but in this embodiment, the vane rotor 21 is at the most retarded angle (default) position. It becomes a characteristic. The one-dot chain line in FIG. 7 is the characteristic (α2 characteristic) of the control phase α2 shown in FIG. 4B, which is the characteristic at the most advanced position of the vane rotor 21 of the present embodiment.

  4B, the position of the piston 2 at the compression top dead center is substantially the same (Y0 described above), but the intake bottom dead center position and the expansion bottom dead center position are different from the α3 characteristic. .

  That is, as shown in FIG. 7, since LC2 <LE2, the mechanical compression ratio C2 <mechanical expansion ratio E2, and the relative ratio D2 = LE2 ÷ LC2> 1, which is a mechanical expansion ratio of mechanical It means that it is relatively larger than the compression ratio.

  In comparison with α3, the mechanical compression ratio is small as C2 <C3, and the mechanical expansion ratio is large as LE2> LE3.

  8A to 8D show mechanism posture change diagrams at the control phase α2. As described above, since LC2 <LE2, mechanical compression ratio C2 <mechanical expansion ratio E2, that is, relative ratio D2> 1.

  Compared with the control phase α3 described in FIGS. 8E to 8H, as described above, LC2 <LC3, LE2> LE3.

  The reason why such characteristics are obtained will be discussed below. When comparing the eccentric rotation direction αC of the eccentric cam portion 13 in the intake bottom dead center posture shown in FIGS. 8B and 8F, αC2 in the α2 characteristic shown in FIG. 8B is equal to αC3 in the α3 characteristic shown in FIG. 8F. On the other hand, the phase changes in the counterclockwise direction (advance direction), that is, the center of the eccentric circle moves to the lower left, so that the control link 14 relatively lowers the second connecting pin 11 to the lower left. The lower link 10 is rotated counterclockwise with the crank pin 9 as a fulcrum, whereby the position of the first connecting pin 8 is raised and the piston 2 is pushed upward by the upper link 7. Thereby, LC2 <LC3.

  On the other hand, looking at the eccentric rotation direction αE of the eccentric cam portion 13 in the expanded bottom dead center posture shown in FIGS. 8D and 8H, αE2 in the α2 characteristic shown in FIG. 8D is compared to αE3 in the α3 characteristic shown in FIG. 8H. Similarly, the counterclockwise direction (advance direction) is changed, that is, the center of the eccentric circle has moved relatively upward, whereby the control link 14 pushes the second connecting pin 11 upward to the right, The lower link 10 is rotated clockwise with the crank pin 9 as a fulcrum, whereby the position of the first connecting pin 8 is lowered, and the piston is pulled downward by the upper link 7. Thereby, LE2> LE3.

That is, the difference in the piston position change characteristic between the control phase α3 and the control phase α2 shown in FIG. 7 is generated by the difference in the link posture due to the difference in the eccentric rotation direction of the eccentric cam portion 13 shown in FIG.
[Performance effect of this embodiment]
After the engine is warmed up, the vane rotor 21 of the phase change mechanism 6 is converted to the most advanced position by the control hydraulic pressure from the electromagnetic switching valve 32, and the characteristics of α2, that is, the mechanical compression ratio C2 is small and the mechanical expansion ratio E2 is small. Great characteristics.

  Here, since the mechanical expansion ratio E2 is large, the work performed when the combustion pressure pushes down the piston can be increased, thereby improving the fuel consumption.

  On the other hand, after the engine is warmed up, if the mechanical compression ratio is high, there is a concern that the in-cylinder gas temperature at the compression top dead center becomes excessively high and the cooling loss increases. As described above, since the mechanical compression ratio C2 is reduced, the fuel consumption (thermal efficiency) can be further improved by suppressing the occurrence of such a cooling loss.

  Further, abnormal combustion such as knocking is likely to occur due to this high mechanical compression ratio at a high engine load, but this can also be avoided by lowering the mechanical compression ratio.

  Here, knocking can also be prevented by controlling the mechanical compression ratio to be lowered even with the above-described conventional technique (Japanese Patent Laid-Open No. 2002-276446), but the mechanical expansion ratio is also lowered and the fuel consumption is deteriorated and the torque is lowered. Accompanied by. Furthermore, there is a problem that the catalyst is thermally deteriorated with an increase in exhaust temperature due to a decrease in mechanical expansion ratio.

  In contrast, in the present embodiment, these problems can be avoided because of the high mechanical expansion ratio.

  By the way, in this embodiment, if it is such a piston position change characteristic even at the time of cold, inconvenience arises from the aspect of exhaust emission. That is, since the mechanical expansion ratio E2 is large, the temperature of exhaust gas discharged from the engine main body is lowered by the amount of expansion work, and the warming of the downstream catalyst does not progress, and the exhaust emission conversion performance by the catalyst is lowered ( F).

  Furthermore, since the mechanical compression ratio is low, the in-cylinder gas temperature at the compression top dead center is relatively low during cold operation, and the combustion is poor, so that the emissions discharged from the engine body itself increase.

  Due to the above two (f) and (ki), emission emissions increase from the tail pipe downstream of the catalyst to the atmosphere.

  Therefore, when the engine is cold, a normal piston position change characteristic such as the α3 characteristic is used. As a result, it is possible to obtain an effect such as a reduction in fuel consumption after the warm-up while avoiding an increase in exhaust emission exhausted to the atmosphere during cold operation.

  Note that at the intermediate temperature between the time of cooling and the time after warming up (while warming up), the rotational phase of the vane rotor 21 of the phase change mechanism 6 is controlled to approach α3 as the temperature decreases, and to α2 as the temperature increases. It is controlled so that it approaches.

Thereby, the fuel efficiency performance and the exhaust emission performance can be optimally balanced for each temperature. For example, fuel consumption can be improved as much as possible while suppressing exhaust emission.
[Third Embodiment]
9 and 10 show a third embodiment of the present invention, which further changes the conversion angle of the vane rotor 21 and the relative rotational phase of the vane rotor 21 and the control shaft 12 with respect to the first and second embodiments. The control phase α1 and the control phase α4 in FIG. 4 can be converted.

  The conversion angle αT in FIG. 3 is further increased to α4-α1 (for example, 240 ° -137 ° = 103 °). Although the conversion angle may be expanded by reducing the number of vanes 27 to two, in this embodiment, the phase change mechanism is described in Japanese Patent Application Laid-Open Nos. 2012-1997755 and 2012-180816. The electric type is used.

  According to the phase change mechanism described in the above two publications, it is a mechanism that converts the phase between the camshaft and the timing sprocket via the speed reduction mechanism by the rotation of the electric motor. The control shaft 12 is used instead of the shaft, and the second gear gear 16 is used instead of the timing sprocket. By configuring in this way, there is no restriction on the conversion angle from the mechanism layout such as the interference between the vane and the housing, and only the relationship between the stopper convex portion and the stopper concave portion set freely as described in the above two publications. This makes it possible to regulate the rotation of the most retarded angle and the most advanced angle.

  In this embodiment, this sets the most advanced angle phase of the output shaft of the electric phase change mechanism to α1 and the most retarded angle phase to α4.

  Further, as in the first and second embodiments, a biasing means (not shown) for biasing the control shaft 12 in the retarding direction is also provided.

  FIG. 9 shows the piston position change characteristic, the broken line shows the characteristic (maximum retardation angle) at the control phase α4, which is the same characteristic as the α4 characteristic of FIG. 5 of the first embodiment, and the solid line shows the control phase. The characteristic at α1 (the most advanced angle) corresponds to the control phase α1 in FIG.

  As can be seen from FIG. 9, in the piston position change characteristic of the control phase α1, the compression stroke LC1 is sufficiently small and the expansion stroke LE1 is sufficiently large. Therefore, the mechanical compression ratio C1 is sufficiently small, the mechanical expansion ratio E1 is sufficiently large, and the relative ratio D1 (E1 ÷ C1) is a large value sufficiently exceeding 1.

FIGS. 10A to 10D show mechanism posture change diagrams at the control phase α1, and similarly to FIGS. 6 and 8, A is an intake (exhaust) top dead center, B is an intake bottom dead center, C is a compression top dead center, D Indicates each posture at the expansion bottom dead center.
Looking at the eccentric rotation direction αC1 of the eccentric cam portion 13 in the intake bottom dead center posture shown in FIG. 7B, the direction is substantially opposite to the direction of the control link 14. For this reason, the control link 14 and the second connecting pin 11 are pulled down almost to the lower left as much as possible, and the lower link 10 changes the phase almost in the counterclockwise direction around the crank pin 9, and accordingly, the first connecting pin 8 moves almost upward as much as possible, so that the upper link 7 pushes the piston 2 upward almost as much as possible.

As a result, LC1 is sufficiently small and almost maximally small. (LC1 <LC2 <LC3 <LC4)
On the other hand, when the eccentric rotation direction αE1 of the eccentric cam portion 13 is viewed in the expanded bottom dead center posture shown in FIG. 10D, the direction is substantially the same as the direction of the control link 14. For this reason, the control link 14 and the second connecting pin 11 are pushed up to the upper right direction as much as possible, and the lower link 10 changes the maximum limit phase in the clockwise direction around the crank pin 3, and accordingly, the first connecting pin 8. Is moved almost downward as much as possible, so that the piston 2 is pulled down almost fully by the upper link 7. Thereby, LE1 is sufficiently large and almost maximized. (LE1>LE2>LE3> LE4)
That is, the relative ratio D1 (= LE1 / LC1) is also sufficiently large and almost maximized. (D1>D2>D3> D4)
These characteristics are produced by the difference in the link posture depending on the eccentric rotation direction of the eccentric cam portion 13.
[Performance effect of this embodiment]
After the warm-up of the internal combustion engine is completed, the eccentric cam portion 13 is converted to the most advanced position by the electric phase change mechanism, and the characteristic of α1, that is, the mechanical compression ratio C1 is sufficiently small and almost maximum. The mechanical expansion ratio E1 is controlled to a sufficiently large and almost maximum characteristic.

  Here, since the mechanical expansion ratio E1 is almost as large as possible, the work performed by the combustion pressure pushing down the piston can be increased almost as much as possible.

  On the other hand, after completion of such warm-up, there is a concern that the in-cylinder gas temperature at the compression top dead center becomes excessively high and the cooling loss increases, but as in this embodiment, the mechanical compression ratio Since C1 can be reduced almost as much as possible, the occurrence of such a cooling loss can be sufficiently suppressed.

  Abnormal combustion such as knocking at a high engine load can be sufficiently suppressed by the substantially minimum mechanical compression ratio C1, and the fuel efficiency can be sufficiently improved by the almost maximum mechanical expansion ratio E1. Furthermore, the exhaust temperature (high temperature exhaust heat exhaust temperature) can be sufficiently lowered by the substantially maximum mechanical expansion ratio E1, and the thermal deterioration of the catalyst can be sufficiently suppressed.

  By sufficient expansion work and cooling loss reduction as described above, fuel consumption (thermal efficiency) can be sufficiently improved, and at high loads, exhaust temperature can be further sufficiently reduced to prevent catalyst thermal deterioration.

  Here, as for the relative ratio D1 (= E1 ÷ C1), as described above, it is a sufficiently large value exceeding 1, and as this is larger, the mechanical expansion ratio is relatively higher and the mechanical compression is higher. This means that the ratio is relatively low, and can be regarded as an index indicating the height of the above-described effect in fuel efficiency.

  If it is such a piston position change characteristic (approximately the minimum mechanical compression ratio, approximately the maximum mechanical expansion ratio, and approximately the maximum relative ratio) at the time of cold, a great inconvenience occurs from the exhaust emission surface.

  That is, since the mechanical expansion ratio E1 is almost the maximum, the exhaust gas temperature exhausted from the engine body is excessively reduced by a sufficient increase in expansion work, so that the warm air in the downstream catalyst does not progress and the emission conversion performance. Is significantly reduced.

  Further, since the mechanical compression ratio is almost minimum, the in-cylinder gas temperature at the compression top dead center is excessively lowered during cold operation, the combustion is significantly deteriorated, and the emission discharged from the engine body is also significantly increased. As a result, the emission emission from the tail pipe downstream of the catalyst is significantly increased.

  Therefore, when the engine is cold, it is converted into a piston position change characteristic in which the mechanical compression ratio C is large and the mechanical expansion ratio E is small like α4. As a result, similar to the first embodiment, the exhaust emission discharged into the atmosphere is greatly reduced by the effect of improving combustion and exhaust temperature, and a normal general piston position change characteristic (for example, α3) Exhaust emissions can be further reduced than the mechanical compression ratio C = mechanical expansion ratio E. That is, the relative ratio D4 is a value lower than 1, which means that the smaller this is, the smaller the expansion ratio is, and the larger the compression ratio is, and this is regarded as an index indicating the good exhaust emission performance. As described above, this can be done.

As described above, as with the first embodiment, the fuel efficiency after warm-up can be improved to the maximum extent, and exhaust emissions during cold-down can be reduced.
In other words, after warming up, the relative ratio is increased to D1 greater than 1 to increase the fuel efficiency, and when cold, the relative ratio is decreased to D4 smaller than 1 to improve exhaust emissions during cold.

  Note that, at an intermediate temperature between the time of cooling and after the completion of warming up (during warming up), the output shaft phase (phase of the eccentric cam portion 13) of the electric phase change mechanism is subjected to high conversion angle control, and the temperature becomes lower. Control is performed so as to approach the control phase α4 and closer to the control phase α1 as the temperature becomes higher.

  Thereby, fuel efficiency performance and emission performance can be balanced at a high level for each temperature. In this case, since an electric phase change mechanism capable of high response conversion regardless of temperature is used, there is no conversion delay with respect to the hydraulic phase change mechanism, and a stable effect can be obtained.

  For example, the fuel consumption can be improved to the maximum while the exhaust emission is stably reduced for every change in temperature.

  Furthermore, in this embodiment, even in a transient operation state, various engine performances can be improved by performing high response large conversion angle phase control by an electric phase change mechanism. For example, transient torque during sudden acceleration can be increased. It can be improved.

  As described above, the knock resistance can be improved by reducing the mechanical compression ratio. However, when the mechanical compression ratio is reduced, the suction stroke (≈compression stroke) tends to decrease, and the filling efficiency decreases. Is also possible.

  Therefore, in order to improve the transient torque, it is necessary to increase the filling efficiency as much as possible while suppressing knocking. Therefore, the instantaneous transient torque is maximized at the moment of acceleration transient. Thus, it may be required to appropriately correct and control the mechanical compression ratio.

  Here, when using a turbocharger such as a turbo, which has been increasing in recent years, the supercharging pressure (which tends to cause large knocking) also changes transiently. It may be required to instantaneously control the mechanical compression ratio that can maximize the filling efficiency as much as possible.

  In response to these requirements, in the present embodiment, since the electric phase change mechanism is used as described above, a high response conversion can be performed regardless of the engine oil pressure or the engine temperature, so that a sufficient transient torque improvement effect can be obtained. Can do it.

  In addition, for example, even when the vehicle accelerates to the high load region and then shifts to the low load region, the high-mechanical expansion ratio can be instantaneously achieved by the electric phase change mechanism, so that the fuel efficiency can be obtained quickly.

As described above, various engine performances can be improved by performing the high response phase control of the phase change mechanism with a large conversion angle.
[Fourth Embodiment]
FIG. 11 shows a fourth embodiment in which the link mechanism 5 is changed, and the two first connection pins 8 and the second connection are connected to the control link 14 connected to the eccentric cam portion 13 of the control shaft 12. The point etc. with which the pin 11 is provided differ from 1st-3rd embodiment.

  That is, the link mechanism 5 includes an upper link 7 connected to the piston 2 via a piston pin 3 and a swing shaft connected to the upper link 7 via a first connection pin 8 and a control shaft. The control link 14 is swingably connected to the 12 eccentric cam portions 13, is swingably connected to the control link 14 via the second connection pin 11, and is rotatable to the crankpin 9 of the crankshaft 4. The lower link 10 is connected to the lower link 10.

  Then, the rotation of the crankshaft 4 is transmitted to the second gear gear 16 (control shaft 12) after being reduced to a half angular velocity via the first gear gear 15 as in the first to third embodiments.

  The second gear gear 16 and the control shaft 12 can change the relative rotational phase by the phase variable mechanism 6 similar to those in the first to third embodiments.

  FIG. 11 shows the posture of the piston 2 in the vicinity of the intake bottom dead center, that is, the position where the crank pin 9 is directed downward.

  Here, the eccentric rotation direction of the eccentric cam portion 13 is substantially right above, and the crank pin 9, the second connecting pin 11, and the piston pin 3 are substantially straight and directly upward. As a result, the position of the piston 2 is shifted slightly downward, and the suction stroke and the compression stroke are increased.

  Because of the upward movement of the eccentric cam portion 13, the control link 14 tilts counterclockwise with the second connecting pin 11 as a fulcrum, so that the first connecting pin 8 moves relatively downward, and the upper link 7 This is because the piston 2 is pulled down.

  Here, the intake / exhaust valve timing is set so that the vicinity of the piston bottom dead center is not the expansion bottom dead center side but the intake bottom dead center side, and the high mechanical compression ratio is increased by increasing the compression stroke. Can be set to

  Next, the expansion bottom dead center will be considered. When the crankshaft 4 is rotated 360 ° (one rotation) in the clockwise direction, the crank pin 9 is again directed directly downward, that is, near the expansion bottom dead center. The eccentric cam portion 13 rotates counterclockwise by 180 ° (half of 360 °) at a half angular velocity via the pair of first and second gear gears 15 and 16. The direction of rotation is now almost directly below.

  As a result, the piston position is now shifted slightly upward. (Expansion stroke reduction). This is because when the eccentric cam portion 13 moves downward, the control link 14 tilts clockwise with the second connecting pin 11 as a fulcrum, so that the first connecting pin 8 moves relatively upward, and the upper link 7 moves to the piston. This is because 2 is pushed up.

  In this way, a low mechanical expansion ratio can be set by reducing the expansion stroke.

  The expansion stroke and the compression stroke referred to here are strokes from the piston position at the compression top dead center, and therefore, as with the α4 characteristic of the first embodiment, the high mechanical compression ratio is low. This is the mechanical expansion ratio.

  As described above, the mechanical compression ratio and the mechanical expansion ratio can be made different even in the link mechanism 5 different from the first to third embodiments.

  Further, if various phase conversions and attachment phase changes are performed by the above-described phase change mechanism, the mechanical compression ratio and the mechanical expansion ratio can be changed appropriately as in the first to third embodiments. .

  The present invention is not limited to the configuration of each of the above-described embodiments. For example, each embodiment shows a single-cylinder internal combustion engine. However, the present invention is applied to multi-cylinders such as 2-cylinder, 3-cylinder, and 4-cylinder. It doesn't matter. In this case, the piston operating characteristics of all the cylinders can be changed by a single or a plurality of phase changing mechanisms, so that all the cylinders can be controlled to a desired mechanical compression ratio and mechanical expansion ratio. Furthermore, it goes without saying that the present invention can be applied not only to a spark ignition engine such as a gasoline engine but also to a compression ignition engine such as a diesel engine.

  In each of the above embodiments, two mechanism types are shown as the link mechanism 5 from the piston 2 to the crankshaft 4. However, the structure of the link mechanism and the like may be appropriately selected without departing from the gist of the present invention. There is no particular limitation.

  Although an example of the pair of first and second gear gears 15 and 16 has been shown as a speed reduction mechanism that decelerates the rotation of the crankshaft 4 to a half angular velocity and transmits it to the eccentric cam portion 13, it is not limited to this. .

  Moreover, in each embodiment, although the rotation direction of the crankshaft 4 and the rotation direction of the eccentric cam part 13 become a reverse direction, it is good also as the same direction. For example, the rotation of the first gear gear 15 on the crankshaft 4 side is decelerated to a half angular velocity via a timing belt (timing chain) and transmitted to the second gear gear 16 on the eccentric cam portion 13 side. Anyway.

In this case, the rotation direction of the crankshaft 4 and the rotation direction of the eccentric cam portion 13 are the same direction, and the piston position change characteristic (vertical axis) with respect to the rotation of the crankshaft 4 (horizontal axis) is reversed left and right. The main point can be realized.
As described above, the configuration is not particularly limited as long as it does not depart from the gist of the present invention.

DESCRIPTION OF SYMBOLS 01 ... Internal combustion engine 02 ... Cylinder block 03 ... Bore 1 ... Piston position variable mechanism 2 ... Piston 3 ... Piston pin 4 ... Crankshaft 5 ... Link mechanism 6 ... Phase change mechanism 7 ... Upper link (1st link)
8 ... 1st connecting pin 9 ... Crank pin 10 ... Lower link (2nd link)
DESCRIPTION OF SYMBOLS 11 ... 2nd connection pin 12 ... Control shaft 13 ... Eccentric cam part 14 ... Control link (3rd link)
15 ... 1st gear gear (drive rotating body)
16 ... Second gear gear (driven rotor)

Claims (9)

A control device for an internal combustion engine comprising a piston position variable mechanism that varies a piston stroke position in a compression stroke and an expansion stroke of a four-cycle internal combustion engine,
The cylinder internal volume at the compression top dead center of the piston is VO, the cylinder internal volume at the piston intake bottom dead center is VC, and the cylinder internal volume at the expansion bottom dead center of the piston is VE,
When a value obtained by dividing the VC by the VO is defined as a mechanical compression ratio C, and a value obtained by dividing the VE by the VO is defined as a mechanical expansion ratio E,
The piston position variable mechanism is configured to change the mechanical compression ratio C and the mechanical expansion ratio E based on the stroke position of the piston, and to change both the C and E differently. A control device for an internal combustion engine. The control apparatus for an internal combustion engine according to claim 1,
When a value (E / C) obtained by dividing the mechanical expansion ratio E by the mechanical compression ratio C is defined as a relative ratio D,
A control apparatus for an internal combustion engine, characterized in that the relative ratio D can be changed. The control apparatus for an internal combustion engine according to claim 2,
The piston position variable mechanism includes a first link connected to the piston via a piston pin, a second link rotatably connected to a crank pin provided on a crankshaft, and a control provided with an eccentric cam. A shaft and a third link rotatably connected to the control shaft;
The first link, the second link, and the third link are linked together, and the control shaft is configured to rotate at an angular velocity half that of the crankshaft,
A control device for an internal combustion engine, comprising: a phase changing mechanism that makes it possible to change a relative rotational phase between the crankshaft and the control shaft. The control apparatus for an internal combustion engine according to claim 2,
A control apparatus for an internal combustion engine, characterized in that the relative ratio D can be changed from a region larger than 1 to a region substantially smaller than 1 to 1. The control apparatus for an internal combustion engine according to claim 2,
The control apparatus for an internal combustion engine, characterized in that the relative ratio D can be changed from a region smaller than 1 to a region substantially larger than 1 to 1. The control apparatus for an internal combustion engine according to claim 2,
A control apparatus for an internal combustion engine, characterized in that when a driving force is not applied to the piston position variable mechanism, the piston position variable mechanism is mechanically stabilized near the minimum value of the relative ratio D. A control device for an internal combustion engine comprising a piston position variable mechanism that varies a piston stroke position in a compression stroke and an expansion stroke of a four-cycle internal combustion engine,
The piston position variable mechanism is provided on a control shaft, a first link connected to the piston via a piston pin, a second link rotatably connected to a crank pin provided on the crankshaft, and the control shaft. A third link rotatably connected to the eccentric cam,
The first link, the second link, and the third link are linked together, and the control shaft is configured to rotate at an angular velocity half that of the crankshaft,
A control device for an internal combustion engine, comprising: a phase changing mechanism that makes it possible to change a relative rotational phase between the crankshaft and the control shaft. A control device for an internal combustion engine comprising a piston position variable mechanism that varies a piston stroke position in a compression stroke and an expansion stroke of a four-cycle internal combustion engine,
The piston position variable mechanism is connected to the piston via a piston pin, and is pivotably connected to the first link via a first connection pin and rotates to the crank pin of the crankshaft. A second link connected to the second link via a second connection pin, and a third link rotatably connected to an eccentric cam provided on the control shaft. With
Configuring the control shaft to rotate at half the angular speed of the crankshaft;
A control device for an internal combustion engine, comprising: a phase changing mechanism that makes it possible to change a relative rotational phase between the crankshaft and the control shaft. The control apparatus for an internal combustion engine according to claim 7 or 8,
The control shaft is rotationally driven by a driven rotator that is rotationally driven by a drive rotator provided on the crankshaft, and
The control apparatus for an internal combustion engine, wherein the phase changing mechanism is arranged coaxially with the control shaft.

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