JP2021021346A - Variable compression ratio system of internal combustion engine - Google Patents

Abstract

To provide a variable compression ratio system of an internal combustion engine which can enhance engine performance such as fuel economy by properly adjusting a mechanical expansion (εE) and a mechanical compression ratio (εC) according to a change of a load.SOLUTION: A variable compression ratio system of an internal combustion engine which can set a mechanical compression ratio (εC) lower than a mechanical expansion ratio (εE) controls the mechanical expansion ratio (εE) to a lower ratio when a load of an internal combustion engine is smaller than a prescribed partial load with respect to the mechanical expansion ratio (εE) when the load of the internal combustion engine is larger than the prescribed partial load. In a region in which the load is larger than the prescribed partial load, the variable compression ratio system improves thermal efficiency by increasing expansion positive work at the high mechanical expansion ratio (εE), and in a region in which the load is smaller than the prescribed partial load, the system improves the thermal efficiency by reducing a pump loss in a latter stage of an expansion stroke by lowering the mechanical expansion ratio (εE).SELECTED DRAWING: Figure 7

Description

The present invention relates to a variable compression ratio system of a four-cycle internal combustion engine, and particularly relates to a variable compression ratio system of an internal combustion engine provided with a variable compression ratio mechanism for changing the positions of top dead center and bottom dead center of a piston. ..

As a conventional variable compression ratio system for an internal combustion engine, various performances of the engine are provided by a variable compression ratio mechanism that variably controls the geometric compression ratio of the internal combustion engine, that is, the mechanical compression ratio (εC) and the mechanical expansion ratio (εE). It has been proposed to improve. For example, the variable compression ratio system of the internal combustion engine described in Non-Patent Document 1 below includes a high expansion general-purpose internal combustion engine system in which the mechanical expansion ratio is set larger than the mechanical compression ratio by a sub-link type piston / crank mechanism. It is shown.

ENGINE TECHNOLOGY REVIEW ・ Vol. 1 No.2 2009/6 "Research on high expansion ratio general-purpose engine"

By the way, the internal combustion engine described in Non-Patent Document 1 described above is a general-purpose internal combustion engine, and although the mechanical compression ratio (εC) and the mechanical expansion ratio (εE) are different values, the mechanical compression ratio (εC) It was not possible to control each of εC) and the mechanical expansion ratio (εE). That is, it is not possible to control the mechanical expansion ratio (εE) and the mechanical compression ratio (εC) in response to changes in the load of the internal combustion engine, and it is difficult to apply it to an internal combustion engine for automobiles with large load fluctuations. Met.

An object of the present invention is a variable compression ratio system for an internal combustion engine capable of appropriately controlling the mechanical expansion ratio (εE) and the mechanical compression ratio (εC) in response to changes in load to improve engine performance such as fuel efficiency. Is to provide.

The present invention is a variable compression ratio system for an internal combustion engine in which the mechanical compression ratio (εC) can be set smaller than the mechanical expansion ratio (εE), and the mechanical expansion ratio (εE) when the load of the internal combustion engine is higher than a predetermined partial load. On the other hand, the mechanical expansion ratio (εE) when the load of the internal combustion engine is lower than the predetermined partial load is controlled to be small.

According to the present invention, in the region where the load is higher than the predetermined partial load, the expansion positive work is increased by the high mechanical expansion ratio (εE) to improve the thermal efficiency, and in the region where the load is lower than the predetermined partial load, the mechanical expansion By reducing the ratio (εE), the pump loss in the latter stage of the expansion stroke, which occurs when the high mechanical expansion ratio is maintained, is reduced to improve the thermal efficiency, so that the load region is higher than the predetermined partial load and the load is lower. Good fuel efficiency can be obtained in both load areas.

It is an overall schematic view of the compression ratio adjustment apparatus which concerns on this invention. It is a side view of the main part which shows the part of the compression ratio adjustment apparatus which concerns on this invention in cross section. It is the front view which removed the front cover of the piston position change mechanism, (A) shows the most retarded angle control state, and (B) shows the most advanced angle state. The operation of the control shaft phase conversion in the variable compression ratio mechanism used in the first embodiment and the second embodiment is shown, and the crankshaft rotation angle (360 °) in which the crankpin near the compression top dead center faces almost directly upward. In (A), the eccentric rotation phase of the control shaft is the control phase α2 (for example, 180 °), (B) is the control phase α1 (for example, 200 °), and (C) is the control phase α3 (for example, 160 °). Each shows a controlled state. It is a characteristic figure which shows the change of the rotation angle of a crankshaft and the height position of a piston with the progress of the combustion cycle in 1st Embodiment and 2nd Embodiment. It is an operation explanatory view of the variable compression ratio mechanism in 1st Embodiment and 2nd Embodiment, and shows the piston position in control phases α1 to α3, (A) is the piston position at exhaust (intake) top dead center, (B) shows the piston position at the intake bottom dead center, (C) shows the piston position at the compression top dead center, and (D) shows the piston position at the expansion bottom dead center. It is explanatory drawing explaining the relationship between the load (torque), the mechanical expansion ratio (εE), and the mechanical compression ratio (εC) in the 1st embodiment and the 2nd embodiment. It is explanatory drawing which shows the PV diagram in 1st Embodiment and PV diagram in the conventional Otto cycle. It is a control chart which shows the control flow which executes the control of 2nd Embodiment.

Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings, but the present invention is not limited to the following embodiments, and various modifications and applications are included in the technical concept of the present invention. Is also included in that range.

First, the first embodiment of the present invention will be described. 1 and 2 show a schematic configuration of a variable compression ratio mechanism. Here, FIG. 1 is a view seen from the right side in FIG.

The internal combustion engine 01 has a piston 2 that reciprocates in the vertical direction along a cylinder bore 03 formed in the cylinder block 02, and a link mechanism 5 described later of the piston pin 3 and the variable compression ratio mechanism 1 by the vertical movement of the piston 2. It is provided with a crankshaft 4 that is rotationally driven via the above. The space separated between the crown surface of the piston 2 in FIG. 1 and the boundary line of the combustion chamber indicated by the one-point chain line is the internal cylinder volume (combustion chamber volume).

Further, an intake valve IV and an exhaust valve EV are provided in the combustion chamber, and are opened and closed by a camshaft (not shown). When these intake valve IV and exhaust valve EV are lifted to the piston 2 side (lower side), they approach the piston crown surface as can be seen from FIG. Here, the lift amount of the intake valve IV is indicated at the position (yi) from the reference position (y = ye = 0) with respect to the piston sliding direction, and the lift amount of the exhaust valve EV is shown from the reference position to the piston sliding direction. On the other hand, it is shown at the position (ye). The position of the piston 2 at this time is Y.

The reference position (y = yes = 0) corresponds to the position where both the intake valve IV and the exhaust valve EV are closed without being lifted. Here, when the piston position Y rises to the (yi) position of the intake valve IV or the (ye) position of the exhaust valve EV at a certain crank angle, interference between the piston crown surface and the intake / exhaust valve will occur. ..

In the following description, the piston 2 is assumed to move up and down in the direction of gravity based on FIG. 1, and the upper side is referred to as "directly above" and the lower side is referred to as "directly below". Actually, the direction of the crank pin and the eccentric direction of the eccentric cam portion, which will be described later, are defined with reference to the sliding direction of the piston 2, and may be read accordingly.

The variable compression ratio mechanism 1 is composed of a link mechanism 5 composed of a plurality of links, a piston position changing mechanism 6 for changing the posture of the link mechanism 5, and the like.

The link mechanism 5 is swingably connected to the upper link 7 via the first connecting pin 8 and the upper link 7 which is the first link connected to the piston 2 via the piston pin 3, and also has a crankshaft. The lower link 10 which is a second link rotatably connected to the crank pin 9 of 4, and the eccentric cam portion of the control shaft 12 which is oscillatingly connected to the lower link 10 via the second connecting pin 11. It is composed of a control link 14 which is a third link rotatably connected to 13.

Further, as shown in FIGS. 1 and 2, a small-diameter first gear gear 15 which is a drive rotating body is fixed to the front end portion of the crankshaft 4, while the driven rotation of the control shaft 12 on the front end portion side. A large-diameter second gear 16 is provided, and the first gear 15 and the second gear 16 mesh with each other, and the rotational force of the crankshaft 4 is transmitted to the control shaft 12 via the piston position changing mechanism 6. It has become so.

The outer diameter of the first gear 15 is about half the outer diameter of the second gear 16, and therefore, the rotation speed of the crankshaft 4 is the first gear 15 and the second gear 16. Due to the difference in outer diameter of the gear, the gear is reduced to half the angular speed and transmitted to the control shaft 12. The phase of the control shaft 12 is changed with respect to the second gear gear 16 by the piston position changing mechanism 6, that is, the relative rotation phase with respect to the crankshaft 4 is changed.

As shown in FIG. 2, 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 piston position changing mechanism 6 has, for example, the same structure as the hydraulic (vane type) variable valve mechanism described in "Japanese Patent Laid-Open No. 2012-225287" filed by the present applicant, and will be briefly described below. To do.

As shown in FIGS. 2 and 3 (A) and 3 (B), the piston position changing mechanism 6 is housed in the housing 20 to which the second gear gear 16 is fixed and in the housing 20 so as to be relatively rotatable. A vane rotor 21 fixed to one end of a control shaft 12 and a hydraulic circuit 22 for rotating the vane rotor 21 in the forward and reverse directions by hydraulic pressure are provided.

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

The rear cover 24 is integrally provided at the center position of the second gear gear 16, and the outer peripheral portion is fixed to the housing body 20a and the front cover 23 together by four bolts 25. Further, in substantially the center of the rear cover 24, a large-diameter bearing hole 24a bearing on the outer periphery is formed through the cylindrical portion of the vane rotor 21 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 integrally provided at a position approximately 90 ° in the circumferential direction of the outer peripheral surface of the rotor 26. In the rotor 26, the small-diameter cylindrical portion 26a on the front end side is rotatably supported by the central support hole of the front cover 23, while the small-diameter cylindrical portion 26b on the rear end side is rotatably supported by the bearing hole 24a of the rear cover 24. It is supported.

Further, the vane rotor 21 is axially fixed to the front end portion of the control shaft 12 by a fixing bolt 28 inserted into the bolt insertion hole of the rotor 26 from the axial direction. Further, each vane 27 is arranged between the shoes 20b, and a seal member and the seal member that are slidably contacted with the inner peripheral surface of the housing body 20a are placed in an elongated holding groove formed in the axial direction of each outer surface. Leaf springs that press in the direction of the inner peripheral surface of the housing body are fitted and held, respectively. Further, four advance chambers 40 and retard chambers 41 are separated from each other between both sides of each vane 27 and both side surfaces of each shoe 20b.

As shown in FIG. 2, the hydraulic circuit 22 supplies and discharges the hydraulic oil of the hydraulic oil to each advance chamber 40 and the first hydraulic passage 29 for supplying and discharging the hydraulic oil to each retard chamber 41. It has two flood control passages with a second hydraulic passage 30 and a supply passage 31 and a drain passage 32 are connected to both hydraulic passages 29 and 30 via an electromagnetic switching valve 33 for switching passages, respectively. ing. The supply passage 31 is provided with a one-way oil pump 35 that pumps the oil in the oil pan 34, while the downstream end of the drain passage 32 communicates with the oil pan 34.

The first and second hydraulic passages 29 and 30 are formed inside a passage constituent portion provided on the front cover 23 side, and one end of each is inside from the small diameter cylinder portion 26a of the rotor 26 of the passage constituent portion. The rotor 26 is communicated with the rotor 26 via a cylindrical portion 36 inserted and arranged in the support hole, while the other end is connected to the electromagnetic switching valve 33.

The first hydraulic passage 29 includes four branch passages (not shown) communicating with each advance chamber 40, while the second hydraulic passage 30 communicates with each retard chamber 41 with a second oil passage. I have. The electromagnetic switching valve 33 is a 4-port 3-position type, and the internal valve body is configured to relatively switch and control the hydraulic passages 29 and 30 and the supply passage 31 and the drain passage 32. The switching operation is performed by the control signal from the control unit 37.

Then, the vane rotor 21 (control shaft 12) is relative 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 33. Can change the rotation phase (hereinafter referred to as relative rotation phase). Further, in each retard angle chamber 41, four coil springs 42 that constantly urge the vane rotor 21 in the retard angle direction are mounted.

FIGS. 4A to 4C show a case where the relative rotation phase of the second gear 16 and the control shaft 12 is changed. In this figure, the first and second gears 15, 16 and the like are omitted. In the present embodiment, the relative rotation phase can be changed by the relative rotation phase conversion control by the piston position changing mechanism 6 described above, but the vane rotor 21 and the control shaft 12 (eccentric cam portion 13) are used. It can also be done by changing the mounting relationship relatively.

In FIG. 4, the crankshaft 4 is rotated clockwise in FIG. 1 without changing the relative rotation phases of the vane rotor 21 and the control shaft 12 shown in FIG. 2, and the crankpin 9 faces directly upward. The posture at the position (crank angle X = 0 ° near the exhaust (intake) top dead center) and the crank pin 9 facing straight up again (X = 360 ° near the compression top dead center). Is shown. At this time, the piston position (height) is high because it is near the compression top dead center.

At this time, in FIG. 4A, the eccentric direction of the eccentric cam portion 13 is clockwise from the direction directly above, for example, the control phase α is retarded by “α2 = 180” °.

That is, the eccentric direction of the eccentric cam portion 13 is near directly below. Here, the eccentric direction means the direction of the line segment connecting the farthest positions of the outer peripheral surfaces of the eccentric cam portion 13 of the control shaft 12 when viewed from the axial center of the control shaft 12.

Further, in FIG. 4B, the eccentric direction of the eccentric cam portion 13 is retarded clockwise by, for example, 200 ° from the direction directly above, and the control phase α is “α1 = 200 °”. .. This is a state in which the angle is retarded by 20 ° as compared with FIG. 4A (a state in which the angle is retarded by 20 ° as compared with the control phase α2), and is the control phase α having the most retarded angle in the present embodiment.

Further, in FIG. 4C, the control phase α is applied in the second embodiment described later, and the eccentric direction of the eccentric cam portion 13 is clockwise retarded by, for example, 160 ° from the directly above direction. The control phase α is “α3 = 160 °”. This is a state advanced by 20 ° with respect to the control phase α2 in FIG. 4A (a state advanced by 40 ° with respect to the control phase α1), and in the present embodiment, it is referred to as the control phase α having the maximum advance angle. It has become.

That is, the most retarded state (least retarded angle) is (B) in FIG. 4, and the most advanced state (most advanced angle) is (C) in FIG. 4, which is in between. Is (A) in FIG. Since the rotation direction of the eccentric cam portion 13 is the counterclockwise direction in FIGS. 4A to 4C, the counterclockwise direction is the advance angle direction. Further, the phase in the state (B) of FIG. 4 is the control phase (latest angle) α1, the phase in the state (A) is the control phase (intermediate angle) α2, and the phase in the state (C) is the control phase (maximum). Advance angle) Sometimes referred to as α3.

Here, for example, the variable compression ratio mechanism 1 (link mechanism) capable of converting between the control phase (latest angle) α1 shown in FIG. 4B and the control phase (intermediate angle) α2 shown in FIG. 4A. The operation of the piston position changing mechanism) will be described with reference to FIGS. 3A and 3B.

FIG. 3 is a view of FIG. 2 viewed from the left side, and the rotation direction of the second gear gear 16 is clockwise in FIG. FIG. 3A shows the most retarded angle position (corresponding to the control phase α1) of the vane rotor 21 of the piston position changing mechanism 6, and FIG. 3B shows the most advanced angle position (corresponding to the control phase α2). Both sides of the vane 27 (27a) with the maximum widening angle are in contact with one side surface and the other side surface of the adjacent shoes 20b at both the most retarded angle and the most advanced angle positions, and stoppers (retarded angle side stopper, advance angle). It is regulated by the side stopper).

In addition, in FIG. 3B, the control phase α2 is the most advanced angle control phase regulated by the advance angle side stopper in the present embodiment (first embodiment), but is the most in the second embodiment described later. Since the angle is slower than the advance control phase α3, the control phase α2 may be referred to as an intermediate angle control phase as described above.

Here, the vane rotor 21 is mechanically stabilized near the most retarded angle position by the spring force of each coil spring 42, as shown in FIG. 3A. That is, the piston position changing mechanism 6, that is, the default position of the variable compression ratio system is the most retarded position. Then, if the phase conversion angle range αT of the piston position changing mechanism 6 is “αT = α1-α2”, for example, 20 ° (= 200 ° −180 °), the control phase (intermediate angle) α2 and the control phase (maximum) The desired control phase conversion angle range αT can be realized by the conversion angle between α1. Further, when the mounting phase of the vane rotor 21 and the control shaft 12 is set so that the above-mentioned latest retard angle position (default position) coincides with the control phase (latest angle) α1 of FIG. 4, desired phase conversion is performed. (Α1⇔α2) can be realized.

Further, in the second embodiment described later, if the phase conversion angle range αT of the piston position changing mechanism 6 is “αT = α1-α3”, for example, 40 ° (= 200 ° −160 °), the control phase (maximum). The desired control phase conversion angle range αT can be realized by the conversion angle between the advance angle (advance angle) α3 and the control phase (latest angle) α1.

FIG. 5 shows the change characteristic of the piston position. Here, when the crank angle X is "0 °", the crank pin 9 is located directly above, and the intake (exhaust) top dead center of the piston 2 is located in the vicinity thereof. Here, the notation of the intake (exhaust) top dead center means that the top dead center at the end of the exhaust stroke and the top dead center at the start of the subsequent intake stroke are the same, which is both the exhaust top dead center and the intake top dead center. Means.

When the crankshaft 4 starts to rotate clockwise in FIG. 1 when the crank angle X is "0 °", the exhaust valve EV is completely closed as shown in the exhaust valve lift curve (ye) in FIG. The intake valve lift curve (yi) of the intake valve IV, which has started to open before "0 °", further increases the lift and sucks fresh air (or air-fuel mixture) from the intake port. Next, the intake bottom dead center is reached near the crank angle X of "180 °", and the lift of the intake valve IV becomes slight near this point. Here, the period from the intake top dead center to the intake bottom dead center is called an intake stroke.

Further, when the crankshaft 4 rotates, the intake valve IV is completely closed, the air-fuel mixture in the cylinder is compressed, and the crank angle X becomes "360 °" (the crank pin 9 is positioned directly above again). ), It becomes the compression top dead center. Here, the period from the intake bottom dead center to the compression top dead center is called a compression stroke.

Spark ignition (or compression ignition) is performed near this compression top dead center, combustion is started, the combustion pressure pushes down the piston 2, and the crank angle X becomes the expansion bottom dead center near "540 °". Become. Here, the period 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 EV starts to open, the combustion gas (exhaust gas) is discharged from the exhaust port as the piston 2 rises again, and the exhaust (intake) top dead center is again near. The crank angle X returns to the position of "720 ° (= 0 °)" (the crank pin 9 is directly above the position). Here, the period from the expansion bottom dead center to the exhaust (intake) top dead center is called the exhaust stroke.

As described above, in the present embodiment, the operation as a four-cycle engine is performed, and the operation is periodic with the crank angle (X) 720 ° as one cycle. In the present embodiment, since the crank angle (X) 720 ° is set as one cycle, the mechanical compression ratio (εC) and the mechanical expansion ratio (εE) can be made different.

For example, as described below, the mechanical expansion ratio (εE) when the load of the internal combustion engine is higher than the predetermined partial load, whereas the mechanical expansion ratio (εE) when the load of the internal combustion engine is lower than the predetermined partial load. By controlling the fuel consumption to a small value, the fuel efficiency can be improved in both the load region higher than the predetermined partial load and the load region lower than the predetermined partial load.

That is, in the load region where the load is higher than the predetermined partial load, the expansion positive work is increased by the high mechanical expansion ratio (εE) to improve the thermal efficiency, and in the load region where the load is lower than the predetermined partial load, the mechanical expansion ratio (εE) By reducing εE), the pump loss in the latter half of the expansion stroke that occurs when the high mechanical expansion ratio is maintained is reduced, and by improving the thermal efficiency, the load region higher than the predetermined partial load and the load lower than the predetermined partial load are achieved. Fuel performance can be improved in both load areas.

In FIG. 5, the thick solid line α2 indicates the piston stroke characteristic (piston crown surface position change characteristic) at the control phase (intermediate angle) α2 of FIG. 4 (A), and the broken line indicates the control phase (between) of FIG. 4 (B). The piston stroke characteristic (piston crown surface position change characteristic) at α1 (latest angle) is shown. The thin solid line α3 shows the piston stroke characteristic (piston crown surface position change characteristic) at the control phase (advance angle) α3 of FIG. 4 (C), which will be described later. In the first embodiment, the control phase is converted between α1 and α2, and in the second embodiment described later, the control phase is converted between α1 and α3.

In both the piston stroke characteristic at the control phase (intermediate angle) α2 and the piston stroke characteristic at the control phase (latest angle) α1, the piston positions (Y02, Y01) at the compression top dead center are substantially the same. (Y0), and the piston position at the expansion bottom dead center and the piston position at the intake bottom dead center are different from each other. Further, in both cases, the piston position at the expansion bottom dead center is set to be lower than the piston position at the intake bottom dead center.

The cylinder internal volume (V) at the compression top dead center is determined by the piston position (Y0) at the compression top dead center in the control phase (intermediate angle) α2 and the control phase (latest angle) α1. It becomes (V0). The cylinder internal volume (V0) is shown in detail, at the compression top dead center, the shape of the combustion chamber surface on the cylinder head side (including the shape of the intake / exhaust valve umbrella surface), the piston crown surface, and the cylinder block inner diameter. It is the volume surrounded by the head gasket etc., that is, the volume of the gas (air-fuel mixture) at the compression top dead center, that is, the in-cylinder volume.

In the characteristics of the control phase (intermediate angle) α2 shown in FIG. 5, the piston position at the intake bottom dead center is the piston position (YC2), and the length from the compression top dead center is the compression stroke (LC2). On the other hand, the piston position at the expansion bottom dead center is the piston position (YE2), and the length from the compression top dead center is the expansion stroke (LE2).

Here, at the intake (exhaust) top dead center, the piston position at the control phase (latest angle) α1 is the piston position (Y'01), and the piston position at the control phase (intermediate angle) α2 is the piston position. (Y'02), and the piston position at the control phase α3 (advance angle) in the second embodiment described later is the piston position (Y'03). The heights of the respective piston positions have a relationship of "Y'01 ≒ Y'02 <Y'03". Further, the intake stroke at that time has a relationship of "LI1> LI2> LI3".

Further, at the intake (exhaust) top dead center after combustion, the piston position at the control phase (latest angle) α1 is the same as the piston position (Y'01) described above, and the control phase (intermediate angle) α2. The piston position is the same as the above-mentioned piston position (Y'02), and the piston position in the control phase α3 (advance angle) is the same as the above-mentioned piston position (Y'03). The height of each piston position has the above-mentioned relationship of "Y'01 ≒ Y'02 <Y'03". Further, the exhaust stroke at that time has a relationship of "LO1 <LO2 <LO3".

Next, the mechanical compression ratio (εC2), which is the mechanical compression ratio (εC) at the control phase (intermediate angle) α2, and the mechanical expansion ratio (εE2), which is the mechanical expansion ratio (εE) at the control phase (intermediate angle) α2. ) Will be explained.

Assuming that the area of the bore (cylinder inner diameter) is "S", the cylinder internal volume (VC2) at the intake bottom dead center is "VC2 = V0 + S x LC2". Therefore, the mechanical compression ratio (εC2) is “εC2 = VC2 ÷ V0 = (V0 + S × LC2) ÷ V0”. On the other hand, the cylinder internal volume (VE2) at the expansion bottom dead center is “VE2 = V0 + S × LE2”. Therefore, the mechanical expansion ratio (εE2) is “εE2 = VE2 ÷ V0 = (V0 + S × LE2) ÷ V0”.

As shown in FIG. 5, since the compression stroke (LC2) and the expansion stroke (LE2) have a relationship of “LC2 <LE2”, the mechanical compression ratio (εC2) and the expansion ratio (εC2) εE2) also has a relationship of “εC2 <εE2”. Considering the absolute values of the mechanical compression ratio (εC2) and the mechanical expansion ratio (εE2), the mechanical compression ratio (εC2) and the mechanical expansion ratio vary depending on the value of the cylinder internal volume (V0) at the compression top dead center. The relationship of "εC2 <εE2" with (εE2) is not broken.

Here, as a result of appropriately selecting the value of the cylinder internal volume (V0) at the compression top dead center, for example, the mechanical compression ratio (εC) is “εC2 = 13.0” and the mechanical expansion ratio (εE) is “εE2 =”. 17.6 ”. The mechanical compression ratio (εC) of "εC2 = 13.0" is a general value, but the mechanical expansion ratio (εE) of "εE2 = 17.6" is a sufficiently high value. Due to this high mechanical expansion ratio (εE2), it is possible to greatly improve thermal efficiency (improve fuel efficiency) above a predetermined partial load (sometimes called a medium load). This will be described later.

Further, in the characteristics of the control phase (latest angle) α1 shown in FIG. 5, the piston position at the intake bottom dead center is the piston position (YC1), and the length from the compression top dead center is the compression stroke (LC1). .. On the other hand, the piston position at the expansion bottom dead center is the piston position (YE1), and the length from the compression top dead center is the expansion stroke (LE1).

Next, the mechanical compression ratio (εC1), which is the mechanical compression ratio (εC) at the control phase (latest angle) α1, and the mechanical expansion ratio (εE), which is the mechanical expansion ratio (εE) at the control phase (latest angle) α1. (ΕE1) will be described.

Since the area of the bore (cylinder inner diameter) is the same "S", the cylinder internal volume (VC1) at the intake bottom dead center is "VC1 = V0 + S x LC1". Therefore, the mechanical compression ratio (εC1) is “εC1 = VC1 ÷ V0 = (V0 + S × LC1) ÷ V0”. On the other hand, the cylinder internal volume (VE1) at the expansion bottom dead center is “VE1 = V0 + S × LE1”. Therefore, the mechanical expansion ratio (εE1) is “εE1 = VE1 ÷ V0 = (V0 + S × LE1) ÷ V0”.

Here, the combustion chamber volume V0 at the compression top dead center is the same combustion chamber volume, assuming that the piston compression top dead center position Y02 at α2 and the piston compression top dead center position Y01 at α1 are substantially the same (Y0). ΕC and εE were calculated as V0, but if there is a significant difference between Y02 and Y01, εC and εE may be calculated using different V0s.

As shown in FIG. 5, since the compression stroke (LC1) and the expansion stroke (LE1) have a relationship of “LC1 <LE1”, the mechanical compression ratio (εC1) and the expansion ratio (εC1) εE1) also has a relationship of “εC1 <εE1”. Considering the absolute values of the mechanical compression ratio (εC1) and the mechanical expansion ratio (εE1), it changes depending on the value of the cylinder internal volume (V0) at the compression top dead center, but the mechanical compression ratio (εC1) and the mechanical expansion ratio The relationship of "εC1 <εE1" with (εE1) is not broken.

Here, as can be seen from FIG. 5, the compression stroke (LC) in the control phase (latest angle) α1 and the control phase (intermediate angle) α2 has a relationship of “LC1> LC2”, and the expansion stroke (LE) is It has a relationship of "LE1 <LE2". Therefore, the mechanical compression ratio (εC1) of the control phase (latest angle) α1 is larger than the mechanical compression ratio (εC2) of the control phase (intermediate angle) α2, and the mechanical expansion ratio (εE1) of the control phase (latest angle) α1. ) Is smaller than the mechanical expansion ratio (εE2) of the control phase (intermediate angle) α2.

As described in the above example, as a result of appropriately selecting the value of the cylinder internal volume (V0) at the compression top dead center, the shape of the combustion chamber and the shape of the piston crown, for example, the mechanical compression ratio (εC) is ". Assuming that εC2 = 13.0 ”and the mechanical expansion ratio (εE) is“ εE2 = 17.6 ”, the cylinders have the same combustion chamber shape and piston crown surface shape, that is, cylinders at almost the same compression top dead center. For example, the mechanical compression ratio (εC) is “εC1 = 13.8” and the mechanical expansion ratio (εE) is “εE1 = 16.3” with respect to the internal volume (V0). As a result, the fuel efficiency performance in a load region lower than the predetermined partial load can be improved. This will be described later.

Here, the reason why the compression stroke (LC) shown in FIG. 5 has a relationship of "LC1> LC2" and the expansion stroke (LE) has a relationship of "LE1 <LE2" is understood from the explanation of the operating posture shown in FIG. can do.

FIG. 6 shows the operating state of each component of the variable compression ratio mechanism 1 that determines the top dead center / bottom dead center position of the piston according to the progress of the combustion cycle. 6 (A) is the intake (exhaust) top dead center, FIG. 6 (B) is the intake bottom dead center, FIG. 6 (C) is the compression top dead center, and FIG. 6 (D) is the expansion bottom dead center. It is shown that the piston position is determined by the eccentric direction of the eccentric cam portion 13 of the variable compression ratio mechanism 1 in the above.

The piston positions that are particularly important in this embodiment are the intake bottom dead center shown in FIG. 6 (B) and the expansion bottom dead center shown in FIG. 6 (D). I will explain it every time.

Looking at the eccentric direction αC of the eccentric cam at the intake bottom dead center shown in FIG. 6B, the eccentricity in the control phase (latest angle) α1 is relative to the eccentric direction αC2 in the control phase (intermediate angle) α2. Direction αC1 is retarded. Therefore, the connecting pin 11 of the control link 14 is pushed up to the upper right relative to the control phase α2, the lower link 10 is rotated clockwise, and the piston 2 is pulled down via the upper link 7, so that the compression stroke ( LC1) is larger than the compression stroke (LC2) at the control phase (intermediate angle) α2.

On the other hand, the eccentric direction αC3 in the control phase (maximum angle) α3 corresponding to the second embodiment described later is advanced in the opposite direction to the eccentric direction αC2 in the control phase (intermediate angle) α2. Therefore, the connecting pin 11 of the control link 14 is pulled down to the lower left relative to the control phase (intermediate angle) α2, the lower link 10 is rotated counterclockwise, and the piston 2 is pushed up via the upper link 7. Therefore, the compression stroke (LC3) is smaller than the compression stroke (LC2) at the control phase (intermediate angle) α2.

Next, looking at the eccentric direction αE of the eccentric cam at the expansion bottom dead center shown in FIG. 6 (D), the control phase (latest angle) α1 with respect to the eccentric direction αE2 at the control phase (intermediate angle) α2. The eccentric direction αE1 in is retarded. Therefore, the connecting pin 11 of the control link 14 is pulled down to the lower left relative to the control phase α2, the lower link 10 is rotated counterclockwise, and the piston 2 is pushed up via the upper link 7, so that the expansion stroke (LE1) is smaller than the expansion stroke (LE2) at the control phase (intermediate angle) α2.

On the other hand, the eccentric direction αE3 in the control phase (maximum angle) α3 corresponding to the second embodiment described later is advanced in the opposite direction to the eccentric direction αE2 in the control phase (intermediate angle) α2. Therefore, the connecting pin 11 of the control link 14 is pushed up to the upper right relative to the control phase (intermediate angle) α2, the lower link 10 is rotated clockwise, and the piston 2 is pulled down via the upper link 7. , The expansion stroke (LE3) becomes larger than the expansion stroke (LE2) at the control phase (intermediate angle) α2.

As described above, according to the link angle characteristic of the variable compression ratio mechanism 1 shown in FIG. 6, the compression stroke (LC) of the control phase (latest angle) α1 and the control phase (intermediate angle) α2 is in the relationship of “LC1> LC2”. In the same way, the expansion stroke (LE) can be set to the relationship of "LE1 <LE2". Therefore, the mechanical compression ratio (εC) can be set to "εC1> εC2", and the mechanical expansion ratio (εE) can be set to "εE1 <εE2". The details of the control phase (advance angle) α3 corresponding to the second embodiment described later will be described later.

In this way, the mechanical compression ratio (εC) can be set to "εC1> εC2" and the mechanical expansion ratio (εE) can be set to "εE1 <εE2" by the link angle characteristic of the variable compression ratio mechanism 1, so that the internal combustion engine can be set. If the mechanical compression ratio (εC) and the mechanical expansion ratio (εE) are controlled according to the load (torque) of the engine, it is possible to improve the performance of the internal combustion engine, particularly the fuel efficiency performance.

Next, specific control regarding the mechanical compression ratio (εC) and the mechanical expansion ratio (εE) with respect to the load (torque) will be described. FIG. 7 shows a load control map of the mechanical compression ratio (εC) and the mechanical expansion ratio (εE) corresponding to the load (torque).

In FIG. 7, the vertical axis indicates the magnitude of the load (torque), and the load (torque) is "no load (T0) (including cranking and stop)" and "low load (T1)". , "Partial load (T2)", "High load (2.5)", and "Full load (T3)". Here, "partial load (T2)" is defined as "predetermined partial load" in this embodiment. Then, for example, when "no load (T0)" to [total load (T3)] is set to 100% load (torque), "low load (T1)" is typically 20%, "partial load". "(T2)" is, for example, 50%, "high load (2.5)" is, for example, 75%, and "total load (T3)" is 100%.

Therefore, "no load (T0)" to "low load (T1)", "low load (T1)" to "partial load (T2)", "partial load (T2)" to "high load (T2.5)" , "High load (T2.5)" to "Full load (T3)", the load (torque) gradually increases. The load (torque) value is an example and is not limited to this. Further, the load (torque) can be typically set by using parameters such as the amount of depression of the accelerator pedal, the opening angle of the throttle valve, and the air flow rate.

In this embodiment, the mechanical compression ratio (εC) and the mechanical expansion ratio (εE) are set as shown by the solid line in FIG. What is characteristic is that the mechanical compression ratio (εC) and the mechanical expansion ratio (εE) are changed with a predetermined “partial load (T2)” as a boundary.

That is, from "no load (T0)" to "low load (T1)", the mechanical compression ratio (εCt0 to εCt1) and the mechanical expansion ratio (εEt0 to εEt1) are set, and "εCt0 = εCt1", It has a relationship of "εEt0 = εEt1".

Further, from "low load (T1)" to "partial load (T2)", the mechanical compression ratio (εCt1 to εCt2) and the mechanical expansion ratio (εEt1 to εEt2) are set, and the load (torque) increases. The mechanical compression ratio (εC) is smaller and the mechanical expansion ratio (εE) is larger.

Further, from "partial load (T2)" to "high load (T2.5)", the mechanical compression ratio (εCt2 to εCt2.5) and the mechanical expansion ratio (εEt2 to εEt2.5) are set. It has a relationship of "εCt2 = εCt2.5" and "εEt2 = εEt2.5".

Further, from "high load (T2.5)" to "total load (T3)", the mechanical compression ratio (εCt2.5 to εCt3) and the mechanical expansion ratio (εEt2.5 to εEt3) are set. It has a relationship of "εCt2.5 = εCt3" and "εEt2.5 = εEt3".

Next, the mechanical compression ratio (εC) and the mechanical expansion ratio (εE) at each load (torque) will be described. First, above the "partial load (T2)", the mechanical expansion ratio (εE) is set to a large mechanical expansion ratio (εEt2 = εEt3), for example, "17.6" corresponding to the above-mentioned εE2, and the mechanical compression ratio. (εC) is set to a normal level mechanical compression ratio (εCt2 = εCt3), for example, “13.0” corresponding to the above-mentioned εC2. As a result, it is possible to greatly improve the thermal efficiency (improve the fuel efficiency). The reason will be described with reference to the PV diagram of FIG.

In FIG. 8, the PV diagram at "full load (T3)" is shown in the upper row, and the normal mechanical compression ratio (εC) and the mechanical expansion ratio (εE) are equal (εC = εE) in the left column. The Otto cycle is shown, and the column on the right side shows the high expansion ratio cycle in which the mechanical expansion ratio (εE) according to the present embodiment is larger than the mechanical compression ratio (εC) (εC <εE).

In the Otto cycle on the left side, regular work is performed clockwise from "compression start ⇒ combustion ⇒ expansion ⇒ exhaust start", and the enclosed area becomes the regular work amount (+). On the other hand, a negative work amount (pump loss called intake loss) occurs counterclockwise from "exhaust start ⇒ intake ⇒ compression start", but this intake loss hardly occurs at full load when the throttle valve is fully opened.

On the other hand, looking at the high expansion ratio cycle on the right side, the amount of regular work (+) has increased significantly. That is, for example, the mechanical expansion ratio (εEt3) is increased to “17.6” corresponding to εE2 and the expansion stroke (LE2) is increased, so that the expansion positive work amount (+) itself due to combustion is increased and the thermal efficiency is increased. Will be higher.

In other words, it is possible to increase the amount of engine work in the area displayed as "Increase in work due to high εE (+)". As a result, even if combustion is performed with the same amount of fuel, a large amount of work can be performed, and fuel efficiency can be significantly improved.

Therefore, the high expansion ratio cycle in this embodiment is set to a large mechanical expansion ratio (εEt3 = for example 17.6) and a normal level mechanical compression ratio (εCt3 = for example 13.0), and is a normal level machine. Sufficient improvement in thermal efficiency (improvement in fuel efficiency) for the Otto cycle set to the expansion ratio (εEt3 = for example 13.0) and the normal level mechanical compression ratio (εCt3 = for example 13.0). Can be done.

Next, the PV diagram at the “partial load (T2)” shown in the middle row will be described. In the "partial load (T2)", when the throttle valve is closed a little and the throttle opening is slightly reduced to lower the load (torque), the maximum combustion pressure (P) is lowered.

Then, in the Otto cycle on the left side, the maximum combustion pressure (P) decreases, so that the above-mentioned positive work amount (+) decreases as compared with the case of "full load (T3)". Further, since the throttle valve is closed a little, the intake loss (pump loss in the intake stroke) increases and the negative work amount (−) increases.

On the other hand, even in the high expansion cycle on the right side, the maximum combustion pressure (P) increases when the throttle valve is closed a little and the throttle opening is slightly reduced to reduce the load (torque), as in the Otto cycle. Since it decreases, the above-mentioned regular work amount (+) decreases as compared with the case of "total load (T3)". However, since the expansion stroke (LE) is increased as in the "total load (T3)", the expansion positive work amount (+) itself due to combustion is increased, and the thermal efficiency is increased.

In other words, it is possible to increase the amount of engine work in the area displayed as "Increase in work due to high εE (+)". As a result, it is possible to sufficiently improve the fuel efficiency performance with respect to the normal Otto cycle. The reason why the intake loss increased with respect to the "total load (T3)" is the same as the Otto cycle.

Next, the PV diagram at "low load (T1)" shown in the lower part will be described. In the "low load (T1)", when the throttle valve is further closed to reduce the throttle opening degree and the load (torque) is lowered, the maximum combustion pressure (P) is further reduced.

Then, in the Otto cycle on the left side, the maximum combustion pressure (P) decreases, so that the above-mentioned positive work amount (+) is further reduced as compared with the case of "partial load (T2)". Further, since the throttle valve is further closed, the intake loss (pump loss in the intake stroke) increases and the negative work amount (−) increases.

On the other hand, even in the high expansion cycle on the right side, as in the Otto cycle, if the throttle valve is further closed to reduce the throttle opening and the load (torque) is lowered, the maximum combustion pressure (P) is further increased. Since it decreases, the above-mentioned regular work amount (+) decreases as compared with the case of "partial load (T2)".

It should be noted here that if the mechanical expansion ratio (εE) is left as large as “full load (T3)” or “partial load (T2)”, the expansion stroke (LE) becomes long and the maximum. In the "low load (T1)" state in which the combustion pressure (P) is decreasing, a phenomenon occurs in which the combustion pressure in the combustion chamber decreases to less than atmospheric pressure (so-called negative pressure) in the latter stage of the expansion stroke.

Therefore, the negative pressure generated in the combustion chamber brakes the lowering operation of the piston, so that a peculiar pump loss (P loss) is conversely generated. The triangular region described as "loss due to high εE (P loss) (-)" shown on the upper side of the high expansion ratio cycle on the right side is expressed as the pump loss due to the negative pressure generated in the latter half of the expansion stroke. There is.

In order to cope with this, in the present embodiment, in the "low load (T1)", the mechanical expansion ratio (εE) is made smaller than the "partial load (T2)" and the "full load (T3)", and the expansion stroke (LE) is shortened. In the present embodiment, the mechanical expansion ratio (εE) in the “low load (T1)” is set to “εEt1 = 16.3”, and the mechanical compression ratio (εC) is set to “εCt1 = 13.8”. Even the mechanical expansion ratio (εEt1 = 16.3) at "low load (T1)" is set to a value larger than the mechanical expansion ratio (εEt1 = 13.0) in the normal Otto cycle.

In the "partial load (T2)" of the present embodiment, the high mechanical expansion ratio (εE) is set so that the combustion pressure near the expansion bottom dead center of the piston is near the atmospheric pressure, which is the above-mentioned 17 It is 0.6. By doing so, in the load region above the "partial load (T2)", the thermal efficiency is improved by this high mechanical expansion ratio (εE), and due to the negative pressure in the cylinder in the latter stage of the expansion stroke in the "partial load (T2)". Pump loss can be reduced or substantially eliminated.

As shown in the PV diagram displayed as [the present embodiment] on the lower side of the lower part on the right side of FIG. 8, the mechanical expansion ratio (εE) is set to “partial load (T2)” in “low load (T1)”. Since the expansion stroke (LE) is shortened by making it smaller than the "total load (T3)", the "loss (P loss) due to high εE" shown in the upper part of the lower part of the high expansion ratio cycle on the right side is shown. The amount of negative work in the triangular area described as "(-)" can be eliminated ([this embodiment] on the lower side of the lower row).

Furthermore, even if the mechanical expansion ratio (εE) is slightly smaller than the high mechanical expansion ratio (εE), it is larger than the mechanical expansion ratio (εE) of the normal Otto cycle. In [this embodiment] on the side, it is indicated by "increased work by medium εE (+)"), and the positive work amount (+) due to the increase in the mechanical expansion ratio (εE) compared to the Otto cycle. Can be increased. This is because when shifting from "partial load (T2)" to "low load (T1)" as shown in FIG. 7, the high mechanical expansion ratio (εEt2 = for example, 17.6 of the above-mentioned εE2) is changed to the medium mechanical expansion ratio ( εEt1 = For example, it corresponds to the transition to 16.3) of εE1 described above.

Returning to the load control map of the load (torque), the mechanical compression ratio, and the mechanical expansion ratio of FIG. 7, further explanation will be added. From "partial load (T2)" to "totally negative (T3)", the mechanical expansion ratio (εE) is set to a constant high mechanical expansion ratio (εEt2 = εEt3), for example, "17.6" of the above-mentioned εE2. Further, the mechanical compression ratio (εC) is set to a constant high mechanical compression ratio (εCt2 = εCt3), for example, “13.0” of the above-mentioned εC2. This control characteristic corresponds to the piston position characteristic formed by the control phase α2 shown in FIG. 5 above. As described above, the high mechanical expansion ratio (εEt2 = εEt3) makes it possible to sufficiently improve the thermal efficiency (improve the fuel efficiency).

Next, when the load (torque) is lower than this "partial load (T2)", as shown in FIG. 7, it is suppressed that the in-cylinder pressure (combustion pressure) is lower than the atmospheric pressure near the bottom dead center of the piston in the expansion stroke. Therefore, the mechanical expansion ratio (εE) will be reduced. For example, this "partial load (T2)" is a load (torque) at which the in-cylinder pressure (combustion pressure) at the expansion bottom dead center is near the atmospheric pressure, which is half of the "total load (T3)" (T3 / 2). It corresponds to the load (torque) in the vicinity or on the slightly lower load side. As a specific guideline for "partial load (T2)", it is about 55% to 25% of the total load, and within this range, the in-cylinder pressure (combustion pressure) near the bottom dead center of the piston in the expansion stroke is easy. Can approach atmospheric pressure.

Then, when the load (torque) of the internal combustion engine drops to "low load (T1)", the mechanical expansion ratio (εE) is set to the medium mechanical expansion ratio (εEt1 = for example, 16.3 of εE1 described above). As described above, the pump loss due to the negative pressure in the latter stage of the expansion stroke is eliminated, reduced or suppressed, and the mechanical expansion ratio (εE) is increased as much as possible (the medium mechanical expansion ratio is usually in the Otto cycle). It is possible to obtain a positive work load (+) due to (set to a value larger than the mechanical expansion ratio of), and it is possible to improve fuel efficiency.

Here, as shown in FIG. 7, in the “low load (T1)”, the mechanical compression ratio (εCt1) is slightly larger than, for example, “13.8” of the above-mentioned εC1. As a result, the fuel efficiency performance at "low load (T1)" can be further improved.

That is, in the "low load (T1)", the in-cylinder temperature of the air-fuel mixture at the compression top dead center of the compression stroke tends to be lower than that in the "partial load (T2)", and there is a problem that combustion deteriorates. On the other hand, by increasing the mechanical compression ratio (εC) to "13.8", the in-cylinder temperature of the air-fuel mixture at the compression top dead center is increased to improve combustion, resulting in "low load (T1)". It is possible to improve the combustion at "low load (T1)" and further improve the fuel efficiency performance at "low load (T1)".

As shown in the PV diagram on the lower right side of FIG. 8, the "intake loss (P loss) (-)" in the "low load (T1)" is due to the throttle opening of the throttle valve being throttled. Pump loss. Therefore, if a variable valve timing mechanism is used on the intake side to control the closing timing of the intake valve in a direction away from bottom dead center, the load (torque) that does not depend on the throttle valve can be reduced (reduction of the intake air amount). ) Can be done. Therefore, the loss in this region can be reduced, and the fuel efficiency can be further improved.

Returning to FIG. 7 again, even if the load (torque) is further reduced from the “low load (T1)” to reach the “no load (T0)” including the engine stop, the mechanical compression ratio (εC) is particularly increased. Since it is maintained at 13.8 near the mechanical compression ratio (εC1), the relationship between the mechanical compression ratio (εC) and the mechanical expansion ratio (εE) is maintained as “εC <εE”. As can be understood from the explanation of FIG. 7, in the present embodiment, the maximum mechanical compression ratio (εC) max in the variable range of the mechanical compression ratio (εC) and the minimum mechanical expansion ratio (εC) in the variable range of the mechanical expansion ratio (εE) The relationship of εE) min is set to "(εC) max <(εE) min".

As a result, a quiet and smooth start can be performed by the effect of reducing the pressure inside the cylinder at the first compression top dead center during the start cranking, the so-called decompression effect. Here, this mechanical compression ratio (εC1) corresponds to the control phase (latest angle) α1, which is the above-mentioned default position (mechanically stable position) of the variable compression ratio mechanism 1, that is, the piston position changing mechanism 6. Therefore, the mechanical compression ratio (εC1) is originally maintained even when the engine is stopped, and the variable compression ratio mechanism 1, that is, the piston position changing mechanism 6 is used in a series of operation processes leading to “engine stop ⇒ cranking ⇒ no load operation”. It is possible to perform a reliable, stable and smooth quiet start without having to control.

Furthermore, since "(εC) max <(εE) min" is set, even if an abnormality should occur in the control system, "(εC) <(εE)" is surely set. The compression effect can be maintained and startability can be ensured.

As described above, in the present embodiment, the load region (here, “high load (T2.5)”, “total load (T2)”) in which the load (torque) is higher than the predetermined partial load (here, “partial load (T2)”) In T3) ”), by setting a high mechanical expansion ratio (εE), the amount of regular work (+) is increased to improve thermal efficiency (improvement of fuel efficiency).

Further, in the load region where the load (torque) is lower than the "partial load (T2)" (here, the region where the load (T1) is reached to the "no load (T0)" via the "low load (T1)"), the mechanical expansion ratio (εE) By reducing the above-mentioned "pump loss in the latter stage of the expansion stroke" that occurs when the high mechanical expansion ratio (εE) is maintained, it is possible to reduce.

As described above, according to the present embodiment, the fuel efficiency can be improved in both the load region higher than the predetermined partial load and the load region lower than the predetermined partial load.

Next, a second embodiment of the present invention will be described. In this second embodiment, the conversion range of the control phase is expanded from the region of α1 to α2 (the first embodiment) to the region of α1 to α3. The control phase (maximum advance angle) α3 shown in FIG. 5 is controlled by the “full load (T3)”, which is different from the first embodiment.

In the second embodiment, in FIG. 7, when the predetermined “high load (T2.5)” is exceeded, the mechanical expansion ratio (εE) is the “partial load (T2)” mechanical expansion ratio (εEt2 = εE2). As shown by the broken line, it is set to the maximum mechanical expansion ratio (εEt3 = εE3) at “total load (T3)”, for example, “17.9”. In this case, from "high load (T2.5)" to "total load (T3)", it is set to continuously increase from the mechanical expansion ratio (εE2) to the mechanical expansion ratio (εE3).

Similarly, the mechanical compression ratio (εC) further decreases from the mechanical compression ratio (εCt2 = εC2) at the “partial load (T2)”, and the minimum mechanical compression ratio (εCt3 =) at the “total load (T3)”. εC3 = for example 12.2). As shown by the broken line, it is set to "12.2", which is the minimum mechanical compression ratio (εC3) in "total load (T3)". In this case, from "high load (T2.5)" to "total load (T3)", it is set to be continuously reduced from the mechanical compression ratio (εC2) to the mechanical compression ratio (εC3).

With such a setting, it is possible to further improve the fuel efficiency performance from "high load (T2.5)" to "total load (T3)". That is, since the mechanical expansion ratio (εE) is further larger than that of the first embodiment, the “increased positive work (+) due to high εE” as shown in the PV diagram on the right side of the upper part of FIG. 8 is further increased. As a result, the improvement of thermal efficiency (improvement of fuel efficiency) can be further improved. Further, by setting the mechanical expansion ratio (εE3) to a higher value, it is possible to suppress an increase in the exhaust gas temperature, which is a problem in the “high load (T2.5)” to “total load (T3)”.

Further, knocking is likely to occur from "high load (T2.5)" to "full load (T3)", but in order to suppress this knocking, the ignition timing is obliged to be retarded, resulting in deterioration of fuel efficiency. And the generated torque was likely to decrease.

On the other hand, in the system of the present embodiment, the mechanical compression ratio (εC) is set small to "12.2", so that the combustion gas temperature at the compression top dead center is lowered and knocking is less likely to occur. It is possible to advance the ignition timing. As a result, the fuel efficiency performance can be further improved, and the absolute value of the generated torque can be increased.

Furthermore, since the exhaust temperature can be further lowered, heat damage due to high-temperature exhaust gas (catalytic heat deterioration, heat resistance problem of the exhaust system), which is a problem in "high load (T2.5)" to "full load (T3)", can be solved. It can be further suppressed.

Further, although the load control map of the load (torque), the mechanical compression ratio, and the mechanical expansion ratio shown in FIG. 7, it is possible to have different maps for each engine speed instead of a single map. Then, more precise control becomes possible.

Next, the description of the variable compression ratio mechanism 1 will be supplemented. In the first embodiment, the control phase (latest angle) α1 is converted to the control phase (intermediate angle) α2 to obtain the mechanical compression ratio (εC1, εC2) and the mechanical expansion ratio (εE1, εE2). The conversion angle range αT (= α1-α2) of the control phase α was, for example, 20 °, which was relatively small.

On the other hand, in the present embodiment, the control phase (latest angle) α1 is converted to the control phase (maximum advance angle) α3 to obtain the mechanical compression ratio (εC1 to εC3) and the mechanical expansion ratio (εE1 to εE3). It can be obtained, and the conversion angle range αT (= α1-α3) of the control phase α is, for example, 40 °, which can be increased to about twice that of the first embodiment. Here, the control phase (intermediate angle) α2 corresponding to the “partial load (T2)” in FIG. 7 is near the intermediate position between the control phase (latest angle) α1 and the control phase (advance angle) α3. ing.

As shown in FIG. 4, in the first embodiment, the phase conversion is from the control phase (latest angle) α1 to the control phase (intermediate angle) α2, but in the present embodiment, the control phase (latest angle) ) Phase conversion from α1 to the control phase (advance angle) α3 becomes possible, and the conversion range is expanded.

The piston position change characteristic at the control phase (maximum advance angle) α3 is shown by the thin solid line in FIG. In FIG. 5, the piston position (YE3) at the expansion bottom dead center is further lower than the piston position (YE2) of the control phase (intermediate angle) α2, and the expansion stroke (LE3) is the control phase (intermediate angle) α2. It is even larger than the expansion stroke (LE2). As a result, the mechanical expansion ratio (εE3) is set to, for example, "17.9", and is larger than the mechanical expansion ratio (εE2) of the control phase (intermediate angle) α2, for example, "17.6".

On the other hand, the piston position (YC3) at the intake bottom dead center is higher than the piston position (YC2) of the control phase (intermediate angle) α2, and the compression stroke (LC3) is the compression stroke of the control phase (intermediate angle) α2. It will be even smaller than (LC2). As a result, the mechanical compression ratio (εC3) is set to, for example, “12.2”, which is even smaller than the mechanical compression ratio (εC2) of the control phase (intermediate angle) α2, for example, “13.0”.

Here, the reason why the compression stroke (LC) shown in FIG. 5 has a relationship of “LC2> LC3” and the expansion stroke (LE) has a relationship of “LE2 <LE3” is understood from the explanation of the operating posture shown in FIG. can do.

Looking at the eccentric direction αC of the eccentric cam at the intake bottom dead center shown in FIG. 6B, the eccentricity in the control phase (advance angle) α3 is relative to the eccentric direction αC2 in the control phase (intermediate angle) α2. The direction αC3 is further advanced by 20 °. Therefore, the connecting pin 11 of the control link 14 is pulled down to the lower left relative to the control phase α2, the lower link 10 is rotated counterclockwise, the upper link 7 is pushed upward, and the piston 2 is further pushed up. , The compression stroke (LC3) is smaller than the compression stroke (LC2) at the control phase (intermediate angle) α2. As a result, the mechanical compression ratio (εC3) is further reduced.

Next, looking at the eccentric direction αE of the eccentric cam at the expansion bottom dead center shown in FIG. 6 (D), the control phase (advance angle) α3 is relative to the eccentric direction αE2 at the control phase (intermediate angle) α2. The eccentric direction αE3 at is further advanced by 20 °. Therefore, the connecting pin 11 of the control link 14 is pushed up to the upper right relative to the control phase (intermediate angle) α2, the lower link 10 is rotated clockwise, and the piston 2 is pulled down via the upper link 7. , The expansion stroke (LE3) becomes larger than the expansion stroke (LE2) at the control phase (intermediate angle) α2. As a result, the mechanical expansion ratio (εE3) becomes even larger. As described above, the compression of the control phase (intermediate angle) α2 and the control phase (advance angle) α3 by the link angle characteristic of the variable compression ratio mechanism 1 shown in FIG. The stroke (LC) can be set in the relationship of "LC2>LC3", and the expansion stroke (LE) can be set in the relationship of "LE2 <LE3" in the same manner. Therefore, the mechanical compression ratio (εC) can be set to “εC2> εC3”, and the mechanical expansion ratio (εE) can be set to “εE2 <εE3”.

In this way, the mechanical compression ratio (εC) can be set to "εC2> εC3" and the mechanical expansion ratio (εE) can be set to "εE2 <εE3" by the link angle characteristic of the variable compression ratio mechanism 1. The amount of regular work (+) at "full load (T3)" can be significantly increased. That is, the mechanical expansion ratio (εEt3) at T3 is increased to “17.9” of εE3, and the expansion stroke (LE3) is further increased, so that the expansion positive work amount (+) itself due to combustion becomes large. , Thermal efficiency is high.

Further, in the present embodiment system, the mechanical compression ratio (εCt3) at T3 is set small to “12.2” of εC3, so that the combustion gas temperature at the compression top dead center drops and knocking occurs. Since it becomes difficult, it is possible to advance the ignition timing. As a result, the fuel efficiency performance can be further improved, and the absolute value of the generated torque can be increased.

Furthermore, the exhaust gas temperature can be further lowered, causing heat damage (catalytic heat deterioration, heat resistance problem of the exhaust system) caused by high-temperature exhaust gas, which is a problem in "high load (T2.5)" to "full load (T3)". It can be further suppressed.

Next, using the compression ratio adjusting mechanism described above, specific control corresponding to the operating state will be described with reference to FIG. 9 based on the second embodiment.

Here, in FIG. 7 described above, the εCεE change characteristic as shown by the solid line is obtained from no load (T0) to high load (T2.5), and the broken line is shown from high load (T2.5) to full load (T3). The second embodiment, which is a characteristic, will be described as a basis. The same applies to the case where the first embodiment in which the high load (T2.5) to the full load (T3) is a solid line is used as a basis, but here, the second embodiment will be described as a basis.

FIG. 9 shows a specific control flowchart thereof. The following control flow is a control flow after the power switch of the automobile is turned on and the microcomputer of the control device is started up.

≪Step S10≫
In step S10, the operating state of the internal combustion engine is read, and at least a signal for driving the starter motor is captured. In addition to this, the temperature of the internal combustion engine and the like are read. When the reading of the operating conditions is completed, the process proceeds to step S11.

<< Step S11 >>
In step S11, it is determined whether or not the starting condition of the internal combustion engine is satisfied. The determination of the starting condition can be performed by monitoring the input state of the starter switch, the rise of the drive signal of the starter motor, and the like. If it is determined that the starting condition is satisfied, the process proceeds to step S12, and if it is determined that the internal combustion engine has already been started, the process proceeds to step S15.

<< Step S12 >>, << Step S13 >>, << Step S14 >>
Since the start condition is satisfied in step S11, the start control is executed in steps S12 to S14.

In step S12, the target control phase α1 suitable for starting the internal combustion engine is controlled. However, the target control position α1 is mechanically stable at the control phase (latest angle) α1 which is the default position (latest angle position) by the variable compression ratio mechanism 1. Therefore, in the process from the stopped state to the start control through cranking, the control phase (latest angle) α1 is stable, so that a reliable and stable start can be performed.

Then, in this state, cranking control and start control are executed in steps S13 and S14. Cranking control executes cranking by a starter motor, and start control executes fuel injection control and ignition control to start an internal combustion engine.

Here, as described above, in the control phase (latest angle) α1, the relationship between the mechanical compression ratio (εC1) and the mechanical expansion ratio (εE1) is “εC1 <εE1”, so that due to the decompression effect. The starting vibration is small, smooth cranking, and good start-up of combustion pressure can be performed.

≪Step S15≫
Since the starting condition is not satisfied in step S11 and the internal combustion engine is operating, the target load (torque) is calculated in step S15. The target load (torque) is calculated from the amount of depression of the accelerator pedal operated by the driver, the throttle valve opening, the intake air flow rate, the vehicle speed, and the like. When the target load (torque) is obtained, the process proceeds to step S16.

≪Step S16≫
In step S16, the target control phase αt is obtained from the load control map as shown in FIG. 7 based on the target load (torque). The target control phase αt is associated with the target load (torque), and when the target load (torque) is obtained, the target control phase αt is searched on the map.

As described above, in "no load (T0)" to "low load (T1)", the control phase (latest angle) α1 is set, and in "low load (T1)" to "partial load (T2)", It is set between the control phase (latest angle) α1 and the control phase (intermediate angle) α2. Further, in the "partial load (T2)", the control phase (intermediate angle) α2 is set.

Next, in "partial load (T2)" to "full load (T3)", the control phase (intermediate angle) α2 or the control phase (intermediate angle) α2 to the control phase (advance angle) α3 is set. To. It should be noted that the control phase αt (= α2) can be set to be constant between the “partial load (T2)” and the “full load (T3)” (corresponding to the first embodiment).

Then, the mechanical compression ratio (εC) and the mechanical expansion ratio (εE) are determined corresponding to the respective target control phases αt. When the target control phase αt is obtained, the process proceeds to step S17.

≪Step S17≫
In step S17, the operating parameters when the internal combustion engine is operating are read. The operating parameters include the engine temperature of the internal combustion engine itself, the outside air temperature of the intake air sucked into the internal combustion engine, the outside air humidity, the in-cylinder pressure (combustion pressure), and the like. The engine temperature is obtained by the water temperature sensor, the temperature and humidity of the intake air are obtained by the temperature sensor and the humidity sensor, and the in-cylinder pressure can be obtained from the combustion pressure sensor or a predetermined mapped in-cylinder pressure map. When the reading of the operating state is completed, the process proceeds to step S18.

≪Step S18≫
In step S18, it is determined whether or not the correction of the target control phase αt is necessary. If the correction is not necessary, the process proceeds to step S19, and if the correction is necessary, the process proceeds to step S20.

The necessity of this correction can be determined by various methods, but the estimated in-cylinder pressure and the actual in-cylinder pressure at the time of the mechanical compression ratio (εC) and the mechanical expansion ratio (εE) determined by the current target control phase αt. The necessity of correction may be determined by comparing the above, or the necessity of correction may be determined based on the engine temperature and the outside air temperature. Since the in-cylinder pressure (combustion pressure) fluctuates depending on the engine temperature, outside air temperature, humidity, etc., it can be determined that correction is necessary when the engine temperature, outside air temperature, humidity, etc. exceed a predetermined threshold value.

≪Step S19≫
In step S18, it is determined that the correction of the target control phase αt is not necessary. Therefore, in step S19, the variable compression ratio mechanism 1 is driven so as to have the target control phase αt obtained in step S16. After that, it goes out to the return and waits for the next start timing.

≪Step S20≫
Since it is determined that the target control phase αt needs to be corrected in step S18, the target control phase αt obtained in step S16 is corrected in step S20.

As described above, the maximum combustion pressure changes depending on the engine temperature, the outside air temperature, the humidity, etc., and the in-cylinder pressure (combustion pressure) at the expansion bottom dead center changes. If the in-cylinder pressure at the bottom dead center of expansion is less than the atmospheric pressure, a pump loss in the latter stage of the expansion stroke will occur. For this purpose, the mechanical expansion ratio (εE) is corrected in the direction of decreasing, that is, the target control phase αt is corrected in the direction of retarding.

That is, if the engine temperature and the outside air temperature are too high or too low, the combustion pressure at the expansion bottom dead center tends to decrease, and the higher the humidity, the lower the combustion pressure at the expansion bottom dead center. According to this, the target control phase αt may be corrected in the direction of retarding.

Alternatively, an in-cylinder pressure (combustion pressure) sensor may be provided in the combustion chamber to read the in-cylinder pressure itself at the expansion bottom dead center. In this case, the accuracy of detecting the in-cylinder pressure at the expansion bottom dead center is improved, so that the accuracy of the control target phase αt correction is improved. If the in-cylinder pressure at the bottom dead center of expansion is less than atmospheric pressure, the mechanical expansion ratio (εE) is corrected to a smaller direction, that is, the control target phase αt is retarded to suppress pump loss in the latter stage of the expansion stroke and achieve fuel efficiency. Performance can be improved.

On the contrary, if the in-cylinder pressure at the expansion bottom dead center exceeds the atmospheric pressure, the target control phase αt is corrected in the direction of advancing, that is, the mechanical expansion ratio (εE) is increased, and the expansion stroke is further expanded. , Expansion positive work can be increased to improve fuel efficiency. When the target control phase αt is corrected in step S20, the process proceeds to step S21.

Although not shown in the control flow, in addition to this mechanical expansion ratio (εE) correction, the combustion pressure at the expansion bottom dead center can also be corrected by the opening timing of the exhaust valve. That is, if the opening time of the exhaust valve is changed in the advance angle direction from the expansion bottom dead center, the in-cylinder pressure at the time of opening the valve can be substantially matched with the atmospheric pressure on the exhaust port side. Therefore, it is possible to prevent the in-cylinder pressure at the bottom dead center of expansion from becoming less than the atmospheric pressure, and it is possible to further suppress the pump loss in the latter stage of the expansion stroke with higher accuracy and improve the fuel efficiency.

≪Step S21≫
Since the correction of the target control phase αt is completed in step S20, in step S21, the piston position conversion mechanism 6 of the variable compression ratio mechanism 1 is driven so as to have the target control phase αt corrected in step S20. After that, it goes out to the return and waits for the next start timing.

As described above, in the present invention, in the variable compression ratio system of the internal combustion engine in which the mechanical compression ratio (εC) can be set smaller than the mechanical expansion ratio (εE), when the load of the internal combustion engine is higher than the predetermined partial load. The mechanical expansion ratio (εE) when the load of the internal combustion engine is lower than the predetermined partial load is controlled to be smaller than the mechanical expansion ratio (εE) of.

According to this, in the load region where the load is higher than the predetermined partial load, the expansion positive work is increased by the high mechanical expansion ratio (εE) to improve the thermal efficiency, and in the load region where the load is lower than the predetermined partial load, the machine By reducing the expansion ratio (εE), the pump loss (negative work) in the latter stage of the expansion stroke that occurs when the high mechanical expansion ratio is maintained is reduced, and by improving the thermal efficiency, a predetermined partial load is applied. The fuel efficiency can be improved in both the higher load region and the lower load region.

In the above-described embodiment, the hydraulic piston position changing mechanism using a vane as the piston position changing mechanism is shown, but the piston position changing mechanism is not limited to the hydraulic type, and for example, an electric type as shown in "Japanese Patent Laid-Open No. 2016-125343". A piston position changing mechanism can also be used. Further, as the link mechanism for changing the piston position, in addition to the mechanism shown in the embodiment, for example, a mechanism as shown in FIG. 11 of "Japanese Patent Laid-Open No. 2016-205173" can be used.

In short, as long as the link mechanism and the phase changing mechanism satisfy the same functions as those in the above-described embodiment, the difference in the specific form does not matter. Further, the applicable internal combustion engine may be a spark ignition type or a compression self-ignition type.

The present invention is not limited to the above-described embodiment, and includes various modifications. For example, the above-described embodiment has been described in detail in order to explain the present invention in an easy-to-understand manner, and is not necessarily limited to the one including all the described configurations. Further, it is possible to replace a part of the configuration of one embodiment with the configuration of another embodiment, and it is also possible to add the configuration of another embodiment to the configuration of one embodiment. Further, it is possible to add / delete / replace a part of the configuration of each embodiment with another configuration.

01 ... Internal combustion engine, 02 ... Cylinder block, 03 ... Bore, 1 ... Variable compression ratio mechanism, 2 ... Piston, 3 ... Piston pin, 4 ... Crankshaft, 5 ... Link mechanism, 6 ... Piston position change mechanism, 7 ... Upper Link (1st link), 8 ... 1st connecting pin, 9 ... Crank pin, 10 ... Lower link (2nd link), 11 ... 2nd connecting pin, 12 ... Control shaft, 13 ... Eccentric cam part, 14 ... Control Link (third link), 15 ... 1st gear gear (driving rotating body), 16 ... 2nd gear gear (driven rotating body).

Claims (9)

In a variable compression ratio system of an internal combustion engine in which the mechanical compression ratio (εC) can be set smaller than the mechanical expansion ratio (εE).
The mechanical expansion ratio (εE) when the load of the internal combustion engine is lower than the predetermined partial load is set smaller than the mechanical expansion ratio (εE) when the load of the internal combustion engine is a predetermined partial load. A variable compression ratio system for internal combustion engines. In the variable compression ratio system of the internal combustion engine according to claim 1.
When the load of the internal combustion engine is lower than the predetermined partial load, the mechanical compression ratio (εC) is set larger than the mechanical compression ratio (εC) when the load of the internal combustion engine is higher than the predetermined partial load. A variable compression ratio system for internal combustion engines. In the variable compression ratio system of the internal combustion engine according to claim 1.
The predetermined partial load is a variable compression ratio system of an internal combustion engine, characterized in that the in-cylinder pressure at the bottom dead center of expansion of the piston is in the vicinity of atmospheric pressure. In the variable compression ratio system of the internal combustion engine according to claim 1.
The predetermined partial load is a variable compression ratio system of an internal combustion engine, characterized in that the load region is about half of the load in the state of full load or lower. In the variable compression ratio system of the internal combustion engine according to claim 1.
A variable compression ratio system for an internal combustion engine, characterized in that the mechanical compression ratio (εC) is reduced and the mechanical expansion ratio (εE) is increased when shifting to a high load region where the load is larger than the predetermined partial load. In the variable compression ratio system of the internal combustion engine according to claim 1.
The characteristic is that the relationship between the mechanical compression ratio (εC) and the mechanical expansion ratio (εE) is set to "(εC) <(εE)" when the internal combustion engine is stopped, cranked, and started. Variable compression ratio system for internal combustion engines. In the variable compression ratio system of the internal combustion engine according to claim 1.
The relationship between the maximum mechanical compression ratio (εC) max in the variable range of the mechanical compression ratio (εC) and the minimum mechanical expansion ratio (εE) min in the variable range of the mechanical expansion ratio (εE) is “(εC) max < A variable compression ratio system for an internal combustion engine, characterized in that it is set to "(εE) min". In the variable compression ratio system of the internal combustion engine according to claim 1.
A variable compression ratio system for an internal combustion engine, characterized in that the mechanical compression ratio (εC) and the mechanical expansion ratio (εE) are set in a relationship of “(εC) <(εE)” at a default position. In the variable compression ratio system of the internal combustion engine according to claim 1.
A variable compression ratio system for an internal combustion engine, characterized in that the mechanical expansion ratio (εE) is corrected based on the operating parameters of the internal combustion engine.

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