JP2007239550A - Compression ratio variable engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To attain both suppression of an NOx emission and improvement of fuel economy by enabling stable cylinder combustion in a lean state that an excess air ratio λ is set at 2 or more. <P>SOLUTION: The internal combustion engine 1 comprising a compression ratio variable mechanism which makes an engine compression ratio variable and a variable valve mechanism 20 which makes at least close timing of an intake valve 11 variable to control an air-intake amount, detects compression pressure of an engine combustion chamber 101 and judges based on an engine operating state whether or not lean combustion is to be performed. The engine is characterized in that when the lean combustion is performed, the excess air ratio is set at a predetermined excess air ratio and in the same state that the excess air ratio is set at the predetermined excess air ratio, the compression ratio and the intake valve closing timing are controlled according to load so that the compression pressure is always equal to or more than the predetermined pressure even when the load is varied. <P>COPYRIGHT: (C)2007,JPO&INPIT

Description

  The present invention relates to a variable compression ratio engine.

In the conventional so-called lean burn engine, in-cylinder combustion does not become unstable, and the air-fuel ratio is made lean within a range that does not exceed the lean air-fuel ratio limit in order to improve fuel efficiency. In an internal combustion engine, it is generally known that a compression ratio should be increased as a means for expanding the lean air-fuel ratio limit. As a technique for increasing the compression ratio of the internal combustion engine, a technique is disclosed in which the compression ratio is made variable by changing the combustion chamber volume in accordance with the output of the air-fuel ratio sensor and the output of the intake pipe pressure sensor. As a result, when the load is low, the compression ratio is increased and the lean air-fuel ratio limit is expanded to enable operation at a leaner air-fuel ratio (see, for example, Patent Document 1).
Japanese Patent Laid-Open No. 60-240837

  However, the above-described conventional compression ratio variable engine controls the compression ratio according to the intake pipe pressure, which is intended to suppress knocking at a high load, and the intake pipe pressure is low. In the low load region, the compression ratio is not controlled according to the intake pipe pressure. On the other hand, when performing lean combustion in such a low load region, the compression ratio control according to the air-fuel ratio is performed, and the compression ratio is controlled to a high compression ratio with the leaning. Therefore, when the load decreases in the lean state, the intake pipe pressure also decreases, but the compression ratio is not changed by this. Therefore, the pressure at the end of the compression stroke (pressure at the start of combustion) (hereinafter referred to as “compression pressure”) also decreases due to the decrease in the intake pipe pressure. When performing lean combustion, the compression pressure at which combustion actually starts is important. When this pressure decreases, stable in-cylinder combustion in a lean state becomes difficult. As a result, there was a problem that the degree of leaning had to be suppressed, leading to deterioration in fuel consumption.

  Moreover, since the NOx (nitrogen oxide) reduction effect by a three-way catalyst cannot be obtained in lean combustion, it is important to reduce NOx discharged directly from the engine. Even if the NOx catalyst is used, since the purification rate is lower than the NOx purification rate of the three-way catalyst in stoichiometric combustion, it is also important to reduce NOx directly discharged from the engine. In lean combustion, it is known that lean combustion with an excess air ratio of approximately 2 will significantly reduce the NOx emission level, but to achieve this, the compression pressure is greatly increased. It is necessary. However, if the actual compression pressure decreases due to the decrease in the intake pipe pressure accompanying the load fluctuation in the low load region, it is difficult to always maintain the excess air ratio in a state of approximately 2 or more. Due to the decrease, there has been a problem that the amount of NOx emission increases.

  The present invention has been made paying attention to such a conventional problem, and at the time of lean combustion, regardless of load fluctuations, the compression pressure is always increased to a predetermined pressure or higher to perform stable lean combustion. With the goal.

  The present invention relates to an internal combustion engine having a variable compression ratio mechanism that makes the engine compression ratio variable and a variable valve mechanism that controls at least the intake valve closing timing to control the intake air amount. Is detected, and it is determined whether or not to perform lean combustion based on the engine operating state. When the lean combustion is performed, the excess air ratio is set to a predetermined lean excess air ratio, and is set to the predetermined lean air excess ratio. The compression ratio and the intake valve closing timing are controlled according to the load so as to be equal to or higher than the pressure.

  According to the present invention, even if there is a load fluctuation during lean combustion, the compression pressure is always maintained at a predetermined pressure or higher. Therefore, even when the excess air ratio is set to, for example, an extremely low NOx emission region of approximately 2 or more, stable lean combustion can be performed.

Hereinafter, embodiments of the present invention will be described with reference to the drawings.
(First embodiment)
FIG. 1 is a view showing a multi-link compression ratio variable engine 1. The engine 1 includes a variable compression ratio mechanism, specifically, a mechanism that changes the compression ratio by changing the piston stroke. The engine 1 having the variable compression ratio mechanism has been previously proposed by the applicant of the present application. However, since the engine 1 is publicly known, for example, in Japanese Patent Application Laid-Open No. 2001-227367, only the outline thereof will be described.

The crankshaft 3 is provided with a crank journal 4 that is rotatably supported by a main bearing (not shown) in the crankcase 2 constituting the engine body in each cylinder. Each crank journal 4 has an axis O ₁ that coincides with the axis (rotation center) of the crankshaft 3 and constitutes a rotating shaft portion of the crankshaft 3.

The crankshaft 3 is eccentric from the axis O ₁ , is provided in each cylinder, a crank arm 5 a that connects the crank pin 5 to the crank journal 4, and the crank pin 5 with respect to the axis O ₁ . And a counterweight 5b that mainly reduces the rotational primary vibration component of the piston motion. The crank arm 5a and the counterweight 5b are integrally formed.

  In this embodiment, the piston 7 slidably fitted to the cylinder 6 formed in each cylinder and the crank pin 5 are connected to a plurality of link members, that is, the upper link 8 (first link) and the lower. The link 9 (second link) is mechanically linked.

The lower link 9 is configured by attaching two main bodies 9 a and 9 b so as to sandwich the crankpin 5. The lower link 9 is connected to the crank pin 5 and the axis O _{2 so as} to be relatively rotatable at the holding portion. The lower link 9, the the one of the lower link body 9a which is almost two equal parts, and the lower end of the upper link 8, the connecting pin 10 inserted through both, relative rotatably connected around axis O _3. On the other hand, the other lower link main body 9b and the upper end side of the control link 11 (third link) are connected to each other around the axis O ₄ by a connecting pin 12 through which both are inserted.

Further, the upper end side of the upper link 8 is externally fitted to a piston pin 13 fixedly provided on the piston 7 so as to be relatively rotatable around the axis O ₅ .

The lower end of control link 11, the large diameter portion 14a of the control shaft 14 which is rotatably supported on the crankcase 2, swingably fitted on its axis O ₆ around and supported. That is, the outer periphery of the small diameter portion 14b of the control shaft 14 has a large diameter portion 14a fixedly provided, the axis O ₆ of the large diameter portion 14a is a place with respect to the axis O ₇ of the small diameter portion 14b Quantitative eccentricity. The control shaft 14 is controlled to rotate according to the operating state of the engine 1 by a compression ratio control actuator 16 via a worm gear 15 and is held at an arbitrary rotation position.

With such a configuration, as the crankshaft 3 rotates, the piston 7 moves up and down in the cylinder 6 via the crankpin 5, the lower link 9, the upper link 8 and the piston pin 13 and is connected to the lower link 9. The control link 11 swings about the swing axis O ₆ on the lower end side as a fulcrum.

  A pent roof type combustion chamber 101 is recessed in the cylinder head covering the top of the cylinder 6. An intake port 102 and an exhaust port 103 are formed so as to open to two inclined surfaces of the combustion chamber 101, respectively. An intake valve 111 opens and closes the tip of the intake port 102, and an exhaust valve 112 opens and closes the tip of the exhaust port 103. Here, the intake port 102 has a bifurcated tip portion, and a pair of intake valves 111 provided in each cylinder open and close the respective tips. Similarly, a pair of exhaust valves 112 are also provided for each cylinder. A spark plug 104 is disposed at the center of the combustion chamber 101 surrounded by these four valves. In addition, a fuel injection valve 105 that injects fuel toward the opening of the intake valve 111 is disposed in the intake port 102, and a swirl control valve (not shown) that adjusts the strength of the flow velocity of the intake air. Is disposed.

  The controller 50 includes a CPU, a ROM, a RAM (not shown), and the like, and is based on operating conditions of the engine 1 detected by various instruments such as a pressure sensor (compression pressure detecting means) that detects a pressure in a cylinder (not shown). The ignition timing, fuel injection amount, compression ratio, etc. of the engine 1 are controlled.

  FIG. 2 is a view for explaining a compression ratio changing method of the multi-link type compression ratio variable engine 1. FIGS. 2A and 2B are views showing the link postures at the high compression ratio position and the low compression ratio position. FIG. 2C is an enlarged view of the vicinity of the control shaft.

When the compression ratio control actuator 16 is controlled by the controller and the control shaft 14 is rotated, the axis O ₇ of the small diameter portion 14 b that becomes the oscillation axis of the control link 11 is rotated around the axis O _{6 of the} large diameter portion 14 a. Rotate to. That is, the engine compression ratio is changed by rotating the control shaft 14 and changing the position of the small diameter portion 14b. For example, as shown in FIGS. 2 (A) and 2 (C), when the axis O _{7 is set} to the position A, the top dead center position becomes high and the compression ratio becomes high.

As shown in FIGS. 2B and 2C, when the axis O _{7 is set} to the position B, the control link 11 is pushed upward, and the position of the axis O ₄ is raised. As a result, the lower link 9 rotates counterclockwise about the axis O ₂ , the axis O ₃ is lowered, and the position of the piston 7 (axis O ₅ ) at the piston top dead center (TDC) is lowered. Therefore, the compression ratio becomes a low compression ratio.

  FIG. 3 is a diagram showing piston stroke characteristics, and FIG. 3A is an enlarged view of a solid line portion surrounded by a square in FIG. 3B.

  The multi-link variable compression ratio engine 1 connects a piston and a crankshaft with a single link (connecting rod), and the piston is higher than a normal engine with a constant compression ratio (hereinafter referred to as “normal engine”). There is a characteristic that the period of staying near the dead center is long.

  This point will be described with reference to FIG. In FIG. 3, the solid line shows the piston stroke characteristics of the multi-link compression ratio variable engine, and in particular, the thick solid line shows the piston stroke characteristics when the compression ratio is high. The thin solid line indicates the piston stroke characteristics when the compression ratio is low. A chain line indicates a piston stroke characteristic of a normal engine. Here, the piston stroke characteristics of the multi-link compression ratio variable engine 1 having the same compression ratio as that of the normal engine are indicated by a thin solid line. From this figure, it can be seen that the variable link ratio compression engine has a longer period in which the piston stays near the top dead center than the normal engine having the same compression ratio.

  Furthermore, when the piston is within a predetermined distance from the top dead center, and the piston is near the top dead center, it is better when the piston is near the top dead center when the compression ratio variable engine has a high compression ratio. The staying period near the top dead center of the piston is longer than when it is near the top dead center at the low compression ratio. That is, in FIG. 3B, L1> L2.

  In this way, the multi-link compression ratio variable engine has a longer period during which the piston stays near the top dead center than the normal engine. Furthermore, the higher the compression ratio, the longer the piston stays near top dead center. The fact that the piston stays in the vicinity of the top dead center means that the high compression state is maintained for a long time during combustion. When the high compression state is maintained for a long time, relatively large combustion energy can be obtained even in the case of ultra lean combustion, so that the combustion characteristics are stabilized.

  Next, the intake valve variable valve mechanism 20 of the engine 1 according to the present invention will be described with reference to FIGS. 4 and 5. FIG. 4 is a perspective view showing the intake valve variable valve mechanism 20 of the engine 1 according to the present invention. FIG. 5 is a drive shaft direction view of the lift / operating angle variable mechanism 30. The intake valve variable valve mechanism 20 previously proposed by the present applicant has been publicly known, for example, in Japanese Patent Laid-Open Nos. 2002-256905 and 11-107725. Only the outline will be described.

  The intake valve variable valve mechanism 20 includes a lift / operation angle variable mechanism 30 (intake valve variable means) that changes the lift / operation angle of the intake valve 111 and a lift center angle of the intake valve 111 (the intake valve 111 has a maximum lift). And a phase variable mechanism 40 for advancing or retarding the phase of the crank angle position). In FIG. 4, only a pair of intake valves 111 corresponding to one cylinder and its related parts are illustrated in a simplified manner.

  Each cylinder of the engine 1 is provided with a pair of intake valves 111 and a pair of exhaust valves 112 (see FIG. 1). A hollow drive shaft 113 extending in the cylinder row direction is provided above the intake valve 111. The drive shaft 113 is linked to the crankshaft 3 by a belt or chain (not shown) via a driven sprocket 41 provided at one end portion, and rotates around the axis in conjunction with the crankshaft 3.

  A pair of swing cams 120 is attached to the drive shaft 113 so as to be rotatable with respect to the drive shaft 113 for each cylinder. As will be described in detail later, the pair of swing cams 120 swings (moves up and down) around a drive shaft 113 within a predetermined rotation range, so that the valve lifter 119 of the intake valve positioned below the swing shaft 120 is located. Is pressed, and the intake valve 111 is lifted downward. The pair of swing cams 120 are fixed to each other in the same phase by a cylinder or the like.

  In FIG. 5, the swing cam 120 is formed with a support hole 122 a penetratingly formed on the base end side thereof, and is rotatably supported on the outer peripheral surface of the drive shaft 113. A pin hole 123a is formed through the end of the cam nose 123. A base circle surface 124a and a cam surface 124b extending in an arc shape from the base circle surface 124a to the tip edge side of the cam nose are formed on the bottom surface of the rocking cam 120. The base circle surface 124a and the cam surface 124b However, it contacts the valve lifter 119 according to the swing position of the swing cam 120. At this time, an arc-shaped ramp surface having a slightly larger diameter than the base circle surface 124a is provided between the base circle surface 124a and the cam surface 124b so that the contact from the base circle surface 124a to the cam surface 124b is performed smoothly. Provided.

The outer periphery of the drive shaft 113, fixed to a position eccentric from the axis P ₃ of the drive shaft 113 by a predetermined amount, so that the center P ₄ of the cylindrical drive cam 115 is positioned, by press-fitting the driving cam 115 To do. The drive cam 115 is fixed at a position away from the swing cam 120 by a predetermined distance in the axial direction. The base end of the link arm 125 is rotatably fitted to the outer peripheral surface of the drive cam 115. The link arm 125 includes an annular base portion 125a having a relatively large diameter and a protruding end 125b protruding from a part of the base portion 125a. A pin hole 125c is formed through the protruding end 125b.

A control shaft 116 extends obliquely above the drive shaft 113 in the cylinder row direction in parallel with the drive shaft 113 and is rotatably supported. On the outer peripheral surface of the control shaft 116, the control cam 117 is press-fitted or the like so that the center P ₁ of the cylindrical control cam 117 is positioned at a position eccentric from the axis P ₂ of the control shaft 116 by a predetermined amount. Fix it. A rocker arm 118 is rotatably fitted to the outer circumferential surface of the control cam 117. The rocker arm 118 swings about the axis P ₁ of the control cam 117 as a fulcrum.

  The rocker arm 118 extends in the left-right direction perpendicular to the axial direction around the central base end portion 118a supported by the control cam 117, and has one end 118b and the other end 118c at both ends thereof. Further, pin holes 118d and 118e are formed through the one end 118b and the other end 118c, respectively.

  One end portion 118b of the rocker arm 118 and the protruding end 125b of the link arm 125 are connected by a connecting pin 121 through which the rocker arm 118 is inserted so that the rocker arm 118 is positioned upward. And the other end part 118c of the rocker arm 118 and the one end part 126a of the link member 126 are connected by the connecting pin 128 which penetrates both. Further, the other end portion 126b of the link member 126 and the end portion on the cam nose 123 side of the swing cam 120 are connected so that the swing cam 120 is positioned below the rocker arm 118 by a connecting pin 129 that passes through both of them. To do. A snap ring for restricting the axial movement of the link arm 125 and the link member 126 is provided at one end of each of the pins 121, 128, and 129.

When the drive shaft 113 rotates in conjunction with the crankshaft 3 by the lift / operating angle variable mechanism 30 configured as described above, the rocker arm 118 is moved to the center P _{1 of the} control cam 117 via the drive cam 115 and the link arm 125. And the swing cam 120 swings within a predetermined angle range via the link member 126. When the swing cam 120 comes into contact with the upper surface of the valve lifter 119 and pushes down the valve lifter 119, the intake valve 111 is driven to open and close.

Further, the control shaft 116 is controlled to rotate within a predetermined rotation angle range by a lift amount control actuator 110 provided at one end. When the control shaft 116 rotates, the center P ₁ of the control cam 117 serving as a rocking fulcrum of the rocker arm 118 is also rotationally displaced, and the support position of the rocker arm 118 with respect to the engine body changes, and this control shaft 116 is applied. The valve lift characteristics of the intake valves 111 of all the cylinders in the cylinder row, specifically, both the lift amount and the operating angle are continuously changed and controlled. FIG. 6 shows the relationship between the lift amount of the intake valve and the operating angle.

  The lift amount control actuator 110 is controlled by the first hydraulic device 51 based on a control signal from the controller 50 that detects the operating state of the engine 1. In the present invention, the intake air amount control of the engine 1 is performed not by the throttle valve but by the valve lift amount and the valve opening time of the intake valve 111. For this reason, the engine speed is controlled to be a small lift when the engine speed is low and a large lift when the engine speed is high. As will be described in detail later, the “small lift” and “large lift” mean whether the maximum lift amount is small or large.

  The phase variable mechanism 40 includes a phase angle control actuator 42 that relatively rotates the sprocket 41 and the drive shaft 113 within a predetermined angle range. The phase angle control actuator 42 is controlled by the second hydraulic device 52 based on a control signal from the controller 50 that detects the operating state of the engine 1. By the hydraulic pressure control to the phase angle control actuator 42, the sprocket 41 and the drive shaft 113 are relatively rotated, and the lift center angle is retarded.

  Next, the operation of the lift / operating angle variable mechanism 30 will be described in detail.

As described above, when the drive shaft 113 rotates in conjunction with the crankshaft, the rocker arm 118 passes the center P ₁ of the control cam 117 via the drive cam 115 and the link arm 125 rotatably fitted to the outer periphery thereof. It swings (moves up and down) as the center. The swing of the rocker arm 118 is transmitted to the swing cam 120 via the link member 126, and the swing cam 120 swings within a predetermined angle range. When the swing cam 120 swings, that is, moves up and down, the valve lifter 119 is pressed and the intake valve 111 is lifted downward.

Here, when the control shaft 116 via the lift control actuator 110 is rotated, the center P ₁ of the control cam 117 serving as a swing fulcrum of the rocker arm 118 also rotationally displaced, the support position of the rocker arm 118 relative to the engine body As a result, the initial swing position of the swing cam 120 changes. Accordingly, the initial contact position between the swing cam 120 and the valve lifter 119 also changes. Since the swing angle of the swing cam 120 per one rotation of the crankshaft is always constant, the maximum lift amount changes as shown in FIG. 7 described below.

  7A and 7B are views of the lift / operating angle variable mechanism 30 in the drive axis direction. FIG. 7A is a diagram showing the position of the swing cam 120 when the intake valve is at zero lift when the swing cam 120 is at the minimum swing or at the maximum swing. FIG. 7B is a view showing the position of the swing cam 120 when the intake valve is fully lifted when the swing cam 120 is at the minimum swing or at the maximum swing. Here, the time of zero lift of the intake valve means that the intake valve 111 does not lift (that is, the lift of the intake valve is zero). Also, when the intake valve is fully lifted, it means that the intake valve 111 has the maximum lift amount.

As shown in FIG. 7A, when the center P ₁ of the control cam 117 is located above the axis P ₂ of the control shaft 116 and the thick portion 117a of the control cam is located above, The rocker arm 118 is positioned upward as a whole, and the end of the swing cam 120 on the side of the connecting pin 129 is relatively lifted upward. That is, the initial position of the swing cam 120 is inclined in a direction in which the cam surface 124b is separated from the valve lifter 119 (see the left side of FIG. 7A). Therefore, when the swing cam 120 swings with the rotation of the drive shaft 113, the base circle surface 124a continues to contact the valve lifter for a long time, and the period during which the cam surface 124b contacts the valve lifter is shortened. For this reason, the maximum lift amount of the intake valve 111 is reduced (see the right side of FIG. 7A). That is, it becomes a small lift. Further, the crank angle section from the opening timing to the closing timing of the intake valve 111, that is, the operating angle of the intake valve 111 is also reduced.

On the other hand, as shown in FIG. 7B, when the center P ₁ of the control cam 117 is located below the axis P ₂ of the control shaft 116 and the thick portion 117a of the control cam is located below. The rocker arm 118 is positioned downward as a whole, and the end of the rocking cam 120 on the side of the connecting pin 129 is relatively pushed down. That is, the initial position of the swing cam 120 is inclined in a direction in which the cam surface 124b approaches the valve lifter 119 (see the left side of FIG. 7B). Therefore, when the swing cam 120 swings with the rotation of the drive shaft 113, the portion that contacts the valve lifter 119 immediately shifts from the base circle surface 124a to the cam surface 124b. For this reason, the maximum lift amount of the intake valve 111 increases (see the right side of FIG. 7B). That is, it becomes a big lift. In addition, the operating angle of the intake valve 111 is increased.

  Since the initial position of the control cam 117 can be continuously changed, the valve lift characteristic of the intake valve 111 is continuously changed accordingly. That is, as shown in FIG. 6, the lift amount and the operating angle of the intake valve 111 can be continuously expanded and reduced simultaneously. Depending on the layout of each part, for example, the opening timing and closing timing of the intake valve 111 change substantially symmetrically as the lift amount and operating angle of the intake valve 111 change.

  FIG. 8 is a diagram illustrating control of operating conditions and intake valve closing timing. As shown in FIG. 8, by combining with the phase variable mechanism 40, the opening timing of the intake valve 111 can be made constant and the closing timing of the intake valve 111 can be set to an arbitrary timing. For example, when it is desired to advance the closing timing of the intake valve 111, such as in a low load range, the central angle is set to an advanced angle with a low lift / small operating angle as shown in FIG. Conversely, when it is desired to delay the intake valve closing timing, such as in a high load range, the center angle is set to be retarded at a high lift / large operating angle, as shown in FIG. In the middle load range, an intermediate lift characteristic is set as shown in FIG.

  FIG. 9 is a diagram showing the relationship between the excess air ratio λ and the NOx emission amount.

  As shown in FIG. 9, the engine has a maximum NOx emission when the air-fuel ratio is somewhat leaner than the stoichiometric air excess ratio λ = 1. Further, the leaner than that, the NOx emission amount is reduced, and when the combustion is performed with the air-fuel ratio being lean to about λ = 2, the NOx emission amount can be reduced to a level almost close to zero. Furthermore, since the engine can reduce fuel consumption when the air-fuel ratio is lean, the fuel consumption can be improved. However, when the engine exceeds the lean air-fuel ratio limit, the amount of fuel contained in the air-fuel mixture is so small that the in-cylinder combustion becomes unstable and cannot be operated.

  Incidentally, it is known that the compression ratio should be increased in order to expand the lean air-fuel ratio limit. This is because if the compression ratio is high, the volume of the combustion chamber at the time of ignition is small, so that it is easy to ignite even with a small amount of fuel.

  FIG. 10 is a graph showing the relationship between the excess air ratio at the lean air-fuel ratio limit and the compression pressure (combustion chamber pressure at the end of the compression stroke (combustion chamber pressure at the start of combustion)). As shown, the excess air ratio at the lean air-fuel ratio limit at which stable in-cylinder combustion is possible is greatly influenced by the compression pressure. That is, by increasing the compression pressure, stable in-cylinder combustion is possible even in a leaner state.

  Here, in the present invention, the engine load (intake air amount) is controlled not at the throttle valve opening / closing control but at the intake valve closing timing IVC. In a general engine, the amount of intake air is adjusted by controlling the opening of a throttle valve to control the engine load. Therefore, the intake valve closing timing IVC is fixed near the bottom dead center (BDC). At this time, if the intake valve closing timing IVC is advanced from the vicinity of BDC to the top dead center (TDC) side, the intake valve 111 is closed while the piston 7 is being lowered. Accordingly, since the actual compression ratio is reduced, the engine output is also reduced. That is, the amount of decrease in the actual compression ratio increases as the advance amount increases, and the engine output also decreases.

  By utilizing such characteristics, the engine 1 according to the present invention, in principle, uses the variable phase mechanism 40 to set the intake valve closing timing IVC in the vicinity of BDC when the load is high, and closes the intake valve as the load decreases. Control is performed so that the amount of advancement of the timing IVC is increased. Thus, the engine load is controlled by controlling the intake valve closing timing IVC without using a slot valve.

  Thus, in the present invention, in order to increase the lean air-fuel ratio limit, the in-cylinder combustion is realized by increasing the compression pressure. At this time, even if the intake valve closing timing IVC changes, the compression pressure is always kept at a predetermined value. Control to be above the pressure. Thus, by always controlling the compression pressure to a predetermined value or higher, stable lean combustion can be performed even when the excess air ratio is set to an extremely low NOx emission region of approximately 2 or more.

  In general engines, at the time of full opening acceleration (at full load), the compression ratio is about 10 (about 12 at the maximum) from the viewpoint of knock prevention, and the intake valve closing timing is behind the bottom dead center (most volume) It is set to the point where efficiency increases. And the compression pressure at this time becomes a value of about 2 (MPa). However, as shown in FIG. 10, when performing lean combustion with an excess air ratio λ = 2 or more, depending on engine characteristics (bore, stroke, gas flow, etc.), compression at the time of full-open acceleration considering at least knocking A pressure of about 3 (MPa) higher than the pressure is required.

  Therefore, in the present embodiment, for example, compared with the compression pressure when the engine rotation speed increases and reaches 3000 rpm during high load operation such as during full-open acceleration, during low load operation (engine rotation speed is lower than the compression pressure). It is necessary to set a high compression pressure (for example, when traveling at a substantially constant speed of 3000 rpm). Hereinafter, the control method will be described.

  FIG. 11 is a control flowchart of the compression ratio and intake valve closing timing of the engine 1 according to this embodiment. In addition, each flow shown below is repeatedly performed for every predetermined single time.

  First, in step S101, the controller 50 detects the engine load and the rotational speed for determining the driving state of the vehicle.

  Next, in step S102, based on the engine load and rotational speed detected in step S101, it is determined whether or not lean combustion with an excess air ratio λ = 2 or more is permitted. In the present embodiment, lean combustion with an excess air ratio λ = 2 or more is permitted when the engine load is less than half of the full load (hereinafter referred to as “low load region”) and the engine rotation speed is 3000 rpm or more.

If lean combustion is permitted, the process proceeds to step S103, and the target excess air ratio λ _Target is read. The target excess air ratio λ _Target is read with reference to a predetermined table shown in FIG. 12 based on the engine load. FIG. 12 is a table in which the engine load is set to λ = 2 in the vicinity of half the load at the full load, and λ is set to lean as the load decreases.

When the target excess air ratio λ _Target is read, the process proceeds to step S104, and the target compression pressure Pc _Target is read. The target compression pressure Pc _Target is read with reference to a predetermined table shown in FIG. 13 based on the target excess air ratio λ _Target . FIG. 13 is a table set so that the target compression pressure Pc _Target increases as the target excess air ratio λ _Target becomes leaner.

When the target compression pressure Pc _Target is read, the process proceeds to step S105, and the target intake valve closing timing IVC _Target is read. In the present embodiment, the target intake valve closing timing IVC _Target is a fixed value IVC _Lean . IVC _Lean is, for example, in the vicinity of BDC, and has high volumetric efficiency so that the excess air ratio can be set to about λ = 2 even when the engine load is close to half of the full load without performing throttling by the throttle. It is set as follows. Since the intake air amount changes depending on the operating state, IVC _Lean may be changed according to the engine speed.

When the target intake valve closing timing IVC _Target is read, the process proceeds to step S106, and the target compression ratio ε _Target is calculated. The target compression ratio ε _Target is calculated with reference to a predetermined map shown in FIG. 14 based on the target compression pressure Pc _Target and the target intake valve closing timing IVC _Target . FIG. 14 is a map in which the target compression ratio ε _Target is set higher as the target compression pressure Pc _Target is higher and the target intake valve closing timing IVC _Target is near TDC.

Note that it is effective to correct the target compression ratio ε _Target when the pressure at the start of compression changes. For example, at high altitudes, the atmospheric pressure changes and the air density decreases, so either the atmospheric pressure or the intake pipe pressure is detected, and the target compression ratio ε _Target is corrected to a high value based on the detection result. It is effective. Thus, stable lean combustion can be realized even under conditions where the intake pipe pressure changes.

  In step S107, the opening degree of the swirl control valve (means for making the intake air flow rate variable) is set. The opening of the swirl control valve is set, for example, from the engine speed or the accelerator opening. This makes it possible to further stabilize the combustion in the lean state by making the flow rate of the intake air flowing into the combustion chamber variable and enhancing the gas flow.

  In step S108, the excess air ratio λ, the intake valve closing timing IVC, and the compression ratio ε are controlled so that the target values obtained in steps S103, S105, and S106 are obtained.

On the other hand, if lean combustion is not permitted in step S102, the process proceeds to step S109. In step S109, the target excess air ratio λ _Target is set to 1, and stoichiometric combustion is basically performed.
When the target excess air ratio λ _Target is read, the process proceeds to step S110, and the target intake valve closing timing IVC _Target is read. As described above, the engine load is controlled not at the throttle valve opening / closing control but at the intake valve closing timing IVC. The target intake valve closing timing IVC _Target is determined in advance as shown in FIG. 15 according to the load. Calculate with reference to the table. FIG. 15 is a table set so that the target intake valve closing timing IVC _Target is advanced from the retarded side to the TDC side from BDC as the load decreases from the time of full load. The volumetric efficiency is highest when the target intake valve closing timing IVC _Target is on the retarded side with _respect to BDC, and the volumetric efficiency decreases with advancement to the TDC side.

In step S111, the target compression ratio ε _Target is read. In the present embodiment, in the case of stoichiometric combustion, the target compression ratio ε _Target is fixed, and is set to about 10 so as not to cause knocking even when fully open.

  Then, the process proceeds to step S108, and the excess air ratio λ, the intake valve closing timing IVC, and the compression ratio ε are controlled so that the target values obtained in steps S109, S110, and S111 are obtained.

FIG. 16 shows the operating states of the excess air ratio λ, the compression pressure Pc, the compression ratio ε, the volume efficiency η _v and the intake valve closing timing IVC when controlled as described above.

  According to the present embodiment described above, in the lean state in which the excess air ratio λ is set to 2 or more, it is higher than the upper limit of the compression ratio that is restricted from the viewpoint of preventing knocking at the time of full opening acceleration regardless of load fluctuations. Control is performed so that combustion is performed at a compression ratio (compression pressure). Therefore, stable in-cylinder combustion can be performed, and further, the cooling loss can be significantly reduced by lowering the combustion gas temperature. Therefore, it is possible to simultaneously suppress the NOx emission amount and improve the fuel consumption.

  Further, by controlling so that the engine load becomes leaner as the engine load becomes lower, the fuel efficiency at the time of low load can be further improved. At the same time, by increasing the compression ratio as the load decreases, stable lean in-cylinder combustion is possible even at low loads. Further, the intake amount at this time is kept constant by fixing the intake valve closing timing. As a result, even if the load becomes low, the compression pressure is not extremely increased (see FIG. 14 and the like). Therefore, the temperature pressure at the time of combustion can be increased without extremely increasing the compression ratio, and the combustion stability can be improved.

  Further, by enhancing the gas flow in the combustion chamber by the swirl control valve, the combustion stability during lean combustion can be further improved.

(Second embodiment)
FIG. 17 is a control flowchart of the compression ratio and intake valve closing timing of the engine 1 according to the second embodiment. This embodiment is different from the first embodiment in that the target excess air ratio is a fixed value and the target intake valve closing timing IVC _Target at the time of lean is read according to the load. Hereinafter, the difference will be mainly described.

  In steps S201 and S202, the same control as in steps S101 and S102 of the control flow according to the first embodiment of FIG. 11 is executed. First, in step S201, the controller 50 determines the engine load for determining the driving state of the vehicle. And detect rotational speed.

  Next, in step S202, based on the engine load and rotational speed detected in step S101, it is determined whether or not lean combustion with an excess air ratio λ = 2 or more is permitted. Also in this embodiment, when the engine rotation speed is 3000 rpm or more and the engine load is in a low load region, lean combustion is permitted for an excess air ratio λ = 2 or more.

If lean combustion is permitted, the process proceeds to step S203, and the target excess air ratio λ _Target is set to 2.

In step S204, the target compression pressure Pc _Target is read. The target compression pressure Pc _Target is read by referring to the predetermined table shown in FIG. 13 described above based on the target excess air ratio λ _Target . In the present embodiment, since the target excess air ratio λ _Target is a fixed value, the target compression pressure Pc _Target may also be a fixed value.

When the target compression pressure Pc _Target is read, the process proceeds to step S205, and the target intake valve closing timing IVC _Target at the time of lean is read. The target intake valve closing timing IVC _Target is read based on the load with reference to a predetermined table shown in FIG. FIG. 18 is a table in which the intake valve closing timing IVC is set so that the volumetric efficiency during lean is higher than the volumetric efficiency during stoichiometry.

When the lean target intake valve closing timing IVC _Target is read, the process proceeds to step S206. In step S206, the same control as in step S106 of the control flow according to the first embodiment of FIG. 11 is executed, and based on the target compression pressure Pc _Target and the target intake valve closing timing IVC _Target , the target compression ratio ε _Target Is calculated. The target compression ratio ε _Target is calculated with reference to the predetermined map shown in FIG.

  In step S207, the excess air ratio λ, the intake valve closing timing IVC, and the compression ratio ε are controlled so that the target values obtained in steps S203, S205, and S206 are obtained.

  On the other hand, the control executed in steps S208, S209, and S210 when the lean combustion is not permitted in step S202 executes the same control as in the first embodiment, and thus the description thereof is omitted here. .

FIG. 19 shows the operating states of the excess air ratio λ, the compression pressure Pc, the compression ratio ε, the volume efficiency η _v, and the intake valve closing timing IVC when controlled as described above.

  According to the present embodiment described above, in the lean state in which the excess air ratio λ is set to 2, the compression ratio (compression pressure) is higher than the upper limit of the compression ratio that is restricted from the viewpoint of preventing knocking at the time of full-open acceleration. Control to perform combustion. Therefore, even in an engine where it is difficult to set the excess air ratio λ to 2 or more, stable in-cylinder combustion can be performed. Further, since the combustion gas temperature is lowered, the cooling loss can be greatly reduced, so that it is possible to simultaneously suppress the NOx emission amount and improve the fuel consumption.

  In addition, as the engine load becomes lower, the intake valve closing timing IVC is advanced so that the intake air amount decreases, and the compression ratio is increased to prevent the lean air-fuel ratio limit from being lowered due to the reduction of the intake air amount. , Lean combustion can be performed.

(Third embodiment)
FIG. 20 is a control flowchart of the compression ratio and intake valve closing timing of the engine 1 according to the third embodiment. This embodiment is in common with the above embodiment in that the lean combustion is permitted when the engine load is in the low load region, but the above point is that the lean combustion is prohibited in the engine extremely low load region described later. It is different from the embodiment. Hereinafter, the difference will be mainly described.

  In step S301, the same control as in step S101 of the control flow according to the first embodiment of FIG. 11 is executed. First, in step S301, the controller 50 determines the engine load and the rotational speed for determining the driving state of the vehicle. To detect.

  Next, in step S302, based on the engine load and rotational speed read in step S101, it is determined whether or not lean combustion with an excess air ratio λ = 2 or higher is permitted. In this embodiment, in principle, when the engine speed is 3000 rpm or more and the engine load is in a low load region, lean combustion with an excess air ratio λ = 2 or more is permitted, but the engine load is 1/4 of the full load. Lean combustion is prohibited in the following extremely low load regions.

If lean combustion is permitted, the process proceeds to step S303, where the target excess air ratio λ _Target is set to 2.

Next, in step S304, the target compression pressure Pc _Target is read. In the present embodiment, the target compression pressure Pc _Target is read with reference to a predetermined table shown in FIG. 21 based on the load. The table in FIG. 21 is set so that the target compression pressure Pc _Target increases as the load decreases regardless of the target excess air ratio λ _Target . This is because even in lean combustion with the same excess air ratio λ = 2, the temperature of the combustion chamber decreases as the load decreases, so that stable in-cylinder combustion is realized by increasing the compression pressure Pc. .

When the target compression pressure Pc _Target is read in step S304, the process proceeds to step S305. In addition, since the control currently performed by step S305-step S307 is performing the control similar to step S205-S207 of the control flow concerning 2nd embodiment of FIG. 17, description is abbreviate | omitted here.

On the other hand, if lean combustion is not permitted in step S302, the process proceeds to step S308. In step S308, the target excess air ratio λ _Target is set to 1, and the stoichiometric combustion is basically performed.

When the target excess air ratio λ _Target is read, the process proceeds to step S110, and the target intake valve closing timing IVC _Target is read. The engine load is controlled not at the throttle valve opening / closing control but at the intake valve closing timing IVC. The target intake valve closing timing IVC _{Target at} this time is determined in advance according to the load shown in FIG. Calculate with reference to the table. The control executed in steps S308 and S309 is the same as the control executed in steps S109 and S110 of the control flow according to the first embodiment in FIG.

  Next, in step S310, the engine load is detected again.

In step S311, it is determined whether or not the load detected in step S310 is within the low load region. If the engine load detected in step S310 is higher than the extremely low load region, the process proceeds to step S312, and the target compression ratio ε _Target is fixed to about 10 so as not to cause knocking even when fully open. On the other hand, when the engine load is in the extremely low load region, the process proceeds to step S313. In this case, the intake valve closing timing is approaching the TDC (see Fig. 15 and the like), the effective compression ratio is lowered, because some cases cylinder combustion becomes unstable, the target compression ratio epsilon _Target first 15 degree Set to. If the process proceeds to step S313, the target compression ratio ε _Target is calculated in step S306. The target compression ratio ε _Target is calculated with reference to the predetermined map shown in FIG.

  Then, the process proceeds to step S307, and the excess air ratio λ, the intake valve closing timing IVC, and the compression ratio ε are controlled so that the target values obtained in steps S308, S309, S312, and S306 are obtained.

FIG. 22 shows the operating states of the excess air ratio λ, the compression pressure Pc, the compression ratio ε, the volume efficiency η _v, and the intake valve closing timing IVC when controlled as described above.

  According to the present embodiment described above, stoichiometric combustion is performed in a region where lean combustion at an extremely low load is difficult. Therefore, at an extremely low load, the catalyst performance is not deteriorated due to a decrease in exhaust temperature due to lean combustion. In addition, since the combustion stability can be ensured by maintaining the compression pressure at a predetermined pressure or higher at all times, the intake valve closing timing IVC is set so that the intake amount decreases as the load decreases at the time of stoichiometric combustion at an extremely low load. It is possible to advance. That is, it is possible to realize a mirror cycle with reduced pump loss while ensuring combustion stability.

(Fourth embodiment)
In this embodiment, the control of the compression ratio of the engine 1 and the intake valve closing timing described in the above embodiments can be performed, but the link geometry setting of the multi-link compression ratio variable engine 1 is changed. The above embodiments are different. The differences will be described below.

  FIG. 23 is an enlarged view of the vicinity of a control shaft of a multi-link compression ratio variable engine. FIG. 23A is an enlarged view of the vicinity of the control shaft according to each of the above embodiments, and FIG. 23B is an enlarged view of the vicinity of the control shaft according to the present embodiment.

  FIG. 24 is a diagram showing the piston stroke characteristics according to the present embodiment, and FIG. 24A is an enlarged view of a solid line portion surrounded by a square in FIG.

  As shown in FIG. 23, in the present embodiment, the control direction and size of the control shaft 14 are changed as compared with the above embodiments. That is, in each of the above-described embodiments, the compression ratio becomes higher as the control shaft 14 is rotated to the right. In the present embodiment, the diameter of the control shaft 14 is larger, and the distance from the center of the control shaft 14 to the axis O7 of the small diameter portion 14b is longer. With such dimensions, as shown in FIG. 24, the piston behavior can be set more characteristically, that is, the behavior near the top dead center can be set extremely, and the piston ascending speed can be made faster than the descending speed. can do. That is, the degree of L31 <L32 increases.

  In addition, the higher the compression ratio is, the faster the piston rises. That is, L31 <L41. Moreover, piston lowering becomes slower as the compression ratio becomes higher. That is, L32> L42.

  According to this embodiment described above, at a high compression ratio, the piston rises rapidly and the period until reaching the top dead center is short. Therefore, pre-ignition (self-ignition) can be prevented. Further, since the piston descends slowly, the stay time near the top dead center of the piston is long, and combustion stability can be secured. On the other hand, since the lowering speed of the piston is high at a low compression ratio, knocking at the late stage of combustion can be prevented.

  The present invention is not limited to the above-described embodiment, and it is obvious that various modifications can be made within the scope of the technical idea.

For example, in the above embodiment, the lean combustion is representatively described as the lean combustion, but the target compression ratio ε _Target is calculated based on the target compression pressure Pc _Target even in the lean combustion by EGR (exhaust gas recirculation). Is valid. Thereby, stable combustion can be realized even when a large amount of EGR is performed.

It is a figure which shows a multiple link type compression ratio variable engine. It is a figure explaining the compression ratio change method by a multiple link type compression ratio variable engine. It is a figure which shows a piston stroke characteristic. It is a perspective view which shows an intake valve variable valve mechanism. It is a drive-axis direction view of a lift / operation angle variable mechanism. It is a figure which shows the relationship between the lift amount of an intake valve, and an operating angle. It is a drive-axis direction view of a lift / operation angle variable mechanism. It is a figure which shows control of an operating condition and an intake valve closing timing. It is the figure which showed the relationship between excess air ratio (lambda) and NOx discharge | emission amount. It is the figure which showed the relationship between the excess air ratio of a lean air fuel ratio limit, and compression pressure. It is a control flowchart of the compression ratio and intake valve closing timing of the multiple link type compression ratio variable engine according to the first embodiment. 6 is a table for setting a target excess air ratio λ _{Target according} to the present invention. It is a table which sets the target compression pressure Pc _Target concerning 1st and 2nd embodiment. 6 is a map for setting a target compression ratio ε _{Target according} to the present invention. It is a table which sets the target intake valve closing timing IVC _Target concerning 1st, 3rd embodiment. It is a figure which shows each operation state of excess air ratio (lambda) with respect to the load concerning 1st embodiment, compression pressure Pc, compression ratio (epsilon), volumetric efficiency (eta) _v, and intake valve closing timing IVC. It is a control flowchart of the compression ratio and intake valve closing timing of the multi-link compression ratio variable engine according to the second embodiment. 10 is a table for setting a target intake valve closing timing IVC _{Target according} to the second embodiment. It is a figure which shows each operation state of excess air ratio (lambda) with respect to the load concerning 2nd embodiment, compression pressure Pc, compression ratio (epsilon), volumetric efficiency (eta) _v, and intake valve closing timing IVC. It is a control flowchart of the compression ratio and intake valve closing timing of the multiple link type compression ratio variable engine according to the third embodiment. It is a table which sets target compression pressure PcTarget _{concerning a} third embodiment. It is a figure which shows each operation state of the excess air ratio (lambda) with respect to the load concerning 3rd embodiment, compression pressure Pc, compression ratio (epsilon), volumetric efficiency (eta) _v, and intake valve closing timing IVC. It is a figure which shows the enlarged view of the control shaft vicinity of a multiple link type compression ratio variable engine. It is a figure which shows the piston stroke characteristic concerning 4th embodiment.

Explanation of symbols

1 Multi-link compression ratio variable engine 3 Crankshaft 7 Piston 8 Upper link (first link)
9 Lower link (second link)
11 Control link (third link)
13 piston pin 14 control shaft 14b small diameter part (eccentric shaft part)
20 Intake valve variable valve mechanism 101 Combustion chamber 111 Intake valve

Claims (14)

In an internal combustion engine comprising a compression ratio variable mechanism that makes the engine compression ratio variable, and a variable valve mechanism that controls the intake air amount by varying at least the intake valve closing timing,
Compression pressure detection means for detecting the compression pressure of the engine combustion chamber;
Lean combustion determination means for determining whether to perform lean combustion based on the engine operating state;
When performing the lean combustion, the excess air ratio setting means for setting the excess air ratio to a predetermined lean excess air ratio; and
Similarly, the lean ratio that controls the compression ratio and the intake valve closing timing according to the load so that the compression pressure is always equal to or higher than the predetermined pressure even when the load fluctuates with the predetermined lean excess air ratio set. Combustion control means;
An internal combustion engine comprising: The internal combustion engine according to claim 1, wherein the compression pressure is a pressure at the end of a compression stroke of the engine combustion chamber. The lean burn determining means permits lean burn if the engine load is in a low load operation region,
The internal combustion engine according to claim 1 or 2, wherein the predetermined pressure is higher than the compression pressure at full load at the same rotational speed. The internal combustion engine according to any one of claims 1 to 3, wherein the excess air ratio setting means sets the excess air ratio to lean as the engine load becomes lower. The lean combustion control means increases the compression ratio so that the compression pressure becomes higher than the predetermined pressure as the engine load becomes lower, and controls the intake valve closing timing so that the intake amount becomes constant. The internal combustion engine according to any one of claims 1 to 4, characterized in that: The lean combustion control means controls the intake valve closing timing so that the intake amount decreases as the engine load becomes lower, and increases the compression ratio so that the compression pressure does not fall below the predetermined pressure. An internal combustion engine according to any one of claims 1 to 4. The lean combustion control means includes
If the engine load is in an extremely low load operation region, lean combustion is prohibited and stoichiometric combustion is performed, and the intake valve closing timing is controlled so that the intake amount decreases as the engine load becomes lower, and the compression pressure The internal combustion engine according to any one of claims 1 to 4, wherein the compression ratio is controlled so that is higher than the predetermined pressure. Means for varying the flow velocity of the intake air flowing into the combustion chamber;
8. The internal combustion engine according to claim 1, wherein when the lean combustion is performed, an intake air flow rate flowing into the combustion chamber is increased as compared with a case where other combustion is performed. . The internal combustion engine according to any one of claims 1 to 8, wherein the lean combustion is lean combustion with an excess air ratio of approximately 2 or more. At least pressure detecting means for detecting either atmospheric pressure or intake pipe pressure;
The internal combustion engine according to any one of claims 1 to 9, wherein the lean combustion means corrects a compression ratio value as a control target based on the pressure detected by the pressure detection means. . The internal combustion engine according to any one of claims 1 to 10, wherein the lean combustion is lean combustion by EGR that introduces exhaust gas into an air-fuel mixture. The compression ratio variable mechanism is
A first link having one end coupled to the piston via a piston pin;
A second link connected at one end to the other end of the first link and rotatably mounted on the crankshaft;
A third link having one end connected to the other end of the second link;
A control shaft that has an eccentric shaft portion that is eccentric with respect to the rotation center shaft, and that freely connects the other end of the third link to the eccentric shaft portion;
Based on the engine operating state, by rotating the control shaft and moving the eccentric shaft portion up and down, the top dead center position of the piston is changed to change the compression ratio,
The internal combustion engine according to any one of claims 1 to 11, characterized in that 13. The link configuration of the first link, the second link, and the third link is set so that the staying period near the top dead center becomes longer as the compression ratio becomes higher in the stroke characteristics of the piston. The internal combustion engine described. The stroke characteristic of the piston is such that the time for moving the predetermined distance from the top dead center is increased with respect to the time for moving the predetermined distance to reach the top dead center. The internal combustion engine according to claim 12, wherein link configurations of the first link, the second link, and the third link are set.

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