JP2008267301A - Internal combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To reduce an actual compression ratio and greatly reduce pumping loss while securing combustion stability by combination of a long stroke piston by a double link type piston-crank mechanism 32 and Miller cycle by a variable valve gear 31. <P>SOLUTION: This engine is provided with: a variable valve gear 31 variably controlling open and close timing of an intake valve; and a variable compression ratio mechanism 32 changing geometrical compression ratio by using the double link type piston-crank mechanism connecting the piston and the crankshaft. Long stroke of the piston is provided without causing increase of base outer shape dimension by constructing the engine to make a largest outer diameter part of a counter weight pass on a side of a pin boss part retaining a piston pin in a vicinity of bottom dead point. Intake valve close timing is advanced in a partial load zone by the variable valve gear 31 to make actual expansion ratio smaller than actual compression ratio. <P>COPYRIGHT: (C)2009,JPO&INPIT

Description

  The present invention relates to an internal combustion engine in which a piston is reciprocated by a piston-crank mechanism and a variable valve operating device is provided on an intake valve side, and more particularly to a technique suitable for improving fuel consumption at a partial load in a gasoline engine.

For example, as described in Patent Document 1, the present applicant has previously proposed a variable valve mechanism capable of continuously expanding / reducing the lift / operation angle of the intake valve, and further, In combination with a mechanism for delaying the phase of the central angle of the valve, a variable valve operating device has been proposed which can obtain a large degree of freedom in lift characteristics. Further, the present applicant uses a multi-link type piston-crank mechanism as a variable compression ratio mechanism of a reciprocating internal combustion engine, and changes a piston top dead center position by moving a part of the link configuration. Various proposals. This type of variable compression ratio mechanism changes the geometric compression ratio (also called the apparent compression ratio or mechanical compression ratio) of an internal combustion engine. It is controlled to a high compression ratio, and at a high load, it is controlled to a low compression ratio to avoid knocking.
JP 2003-314315 A

  As is well known, in the Miller cycle engine, the intake valve closing timing is set to early closing or late closing with respect to the intake bottom dead center, and the actual compression ratio is reduced without lowering the actual expansion ratio. It is well known that it is theoretically possible to substantially reduce or eliminate pump loss by bringing the pressure close to atmospheric pressure. Here, the actual compression ratio is a compression ratio when compression is performed up to the piston top dead center on the assumption that compression starts from the intake valve closing timing, and the actual expansion ratio expands from the piston top dead center and exhausts. The expansion ratio assumes that the expansion ends at the valve opening timing.

  However, when trying to make an actual gasoline engine into a mirror cycle, the temperature near the compression top dead center becomes too low as the actual compression ratio decreases, and there is a problem that good ignitability and combustion stability cannot be secured. is there. In response to this problem, a technology that ensures combustion stability by increasing the geometric compression ratio as the actual compression ratio decreases, and improves fuel efficiency through mirror cycle, in combination with the variable compression ratio mechanism described above. The applicant has previously proposed. However, if the geometric compression ratio is simply increased by the variable compression ratio mechanism, the shape of the combustion chamber is flattened as the compression ratio is increased, and the increase in cooling loss becomes remarkable. If the piston stroke characteristics are close to simple vibration compared to the single link type by the above-mentioned multi-link type piston-crank mechanism, the piston speed near the top dead center will be slowed down. There is a tendency to increase, and the combustion stability cannot be secured satisfactorily due to the temperature drop near the compression top dead center, and further improvement has been desired.

  The present invention has been made paying attention to such a problem, and the piston-crank mechanism is intended to increase the piston stroke without causing an increase in the basic external dimensions of the internal combustion engine. Combined with longer stroke and variable control of intake valve closing timing, it effectively reduces the increase in cooling loss due to a decrease in actual compression ratio (mirror cycle) without causing flattening of the combustion chamber shape. The object is to provide a new internal combustion engine that eliminates the problem and can significantly reduce pump loss and improve fuel efficiency at the time of partial load.

  The present invention provides an internal combustion engine having a piston-crank mechanism that connects a piston that reciprocates in a cylinder and a crankshaft, and a variable valve operating device that can variably control at least the closing timing of an intake valve. In the vicinity of the bottom dead center, the outermost diameter portion of the counterweight passes through the side of the pin boss portion holding the piston pin. In other words, the outermost diameter portion of the counterweight of the crankshaft intersects the extension line in the axial direction of the piston pin in the vicinity of the bottom dead center. In other words, the distance between the piston and the center of the crankshaft at the bottom dead center position is set to be very small, and this does not cause an increase in the basic dimensions of the internal combustion engine, and the bottom dead center from the top dead center. It is possible to secure a larger piston stroke and thus a larger displacement.

  Desirably, at least during partial load, the intake valve closing timing is controlled so that the actual compression ratio determined by the intake valve closing timing is smaller than the actual expansion ratio determined by the exhaust valve opening timing. Pump loss including throttle loss can be remarkably reduced by such a reduction in actual compression ratio, that is, mirror cycle. Here, the actual compression ratio and the actual expansion ratio both depend on the geometric compression ratio as specific values. However, on the comparison of the two sizes, the geometric ratio at that time The compression ratio is not affected. In other words, the relationship that “the actual compression ratio is smaller than the actual expansion ratio” is “(actual compression ratio / actual expansion ratio) <1”. Since they are the same, there is no need to consider them, and the above magnitude relationship is determined by the intake valve closing timing. The relationship “(actual compression ratio / actual expansion ratio) <1” means that the actual compression ratio is positively suppressed. As methods for reducing the actual compression ratio in this way, there are a so-called early closing method in which the intake valve is closed earlier than the intake bottom dead center, and a so-called late closing method in which the intake valve is closed later than the intake bottom dead center.

  Then, by combining the above-described technology for increasing the stroke of the piston with the decrease in the actual compression ratio, the increase in the cooling loss due to the decrease in the actual compression ratio is suppressed without causing the flattening of the combustion chamber shape, It will be possible to significantly improve fuel efficiency.

  Each of the pistons has a skirt portion at a portion on the thrust-anti-thrust side in the circumferential direction. Is the width of the skirt portion along the axial direction of the piston pin substantially equal to the total length of the piston pin? Or shorter.

  In the present invention, preferably, as the piston-crank mechanism, one end of the upper link is connected to the piston via a piston pin, and the other end of the upper link is connected via a first connection pin. A lower link rotatably attached to the crankpin of the crankshaft, and a control having one end connected to the lower link via a second connection pin and the other end supported to be swingable with respect to the internal combustion engine body And a multi-link piston-crank mechanism having a link.

  If the structure further includes a support position changing means for displacing the swing support position of the other end of the control link relative to the internal combustion engine body, the geometric compression ratio is variably controlled by the displacement of the swing support position. A variable compression ratio mechanism can be easily realized.

  In such a multi-link piston-crank mechanism, when the piston is in a position to receive the maximum combustion load, the inclination of the upper link with respect to the cylinder axis is smaller than in the case of the single-link piston-crank mechanism. It is possible to set the link configuration. Thus, the side thrust load acting on the piston is relatively reduced by the posture of the upper link approaching the vertical. Therefore, the piston skirt portion can be reduced in size.

  In the multi-link type piston-crank mechanism, the link configuration is set so that the piston stroke characteristic of the piston with respect to the rotation of the crankshaft is closer to the single vibration characteristic than that of the single-link piston-crank mechanism. It is also easy to do. If the characteristics are close to simple vibrations in this way, the piston acceleration is leveled and the maximum inertial force near the top dead center is reduced, which is advantageous in reducing the size of the piston pin and the pin boss. Moreover, it is advantageous in terms of noise vibration characteristics, and for example, it is possible to avoid the deterioration of the inertial secondary vibration of the piston accompanying the expansion of the piston stroke of the in-line four-cylinder engine.

  According to the present invention, it is possible to achieve a long stroke of the piston without causing an increase in the basic external dimensions of the internal combustion engine, and the actual compression ratio by such a long stroke and variable control of the intake valve closing timing. In combination with a decrease in the temperature, excessive flattening of the combustion chamber shape can be prevented, so an increase in cooling loss due to a decrease in the actual compression ratio (mirror cycle) is suppressed, resulting in a significant pump loss during partial loads. Reduction and fuel efficiency can be improved.

  Hereinafter, an embodiment of the present invention will be described in detail with reference to the drawings. FIG. 1 shows an embodiment of a control device for an internal combustion engine according to the present invention. This internal combustion engine includes a variable valve operating device 31 for variably controlling the intake valve opening / closing timing, a compression ratio variable mechanism 32 for variably controlling the geometric compression ratio of the internal combustion engine, and an ignition advance angle for controlling the ignition timing. And a control device 33.

  FIG. 2 is an explanatory diagram showing the configuration of the variable valve operating device 31. The variable valve operating device 31 includes a lift / operating angle variable mechanism 34 for changing the lift / operating angle of the intake valve 41, and the lift. And a phase variable mechanism 35 for advancing or retarding the phase of the central angle (phase with respect to the crankshaft).

  First, the lift / operating angle variable mechanism 34 will be described. The lift / operating angle variable mechanism 34 has been previously proposed by the applicant of the present invention, and is known along with the phase variable mechanism 35 from Japanese Patent Application Laid-Open Nos. 2002-89303 and 2002-89341. Therefore, only the outline will be explained. The lift / operating angle variable mechanism 34 includes a hollow drive shaft 36 rotatably supported by a cam bracket (not shown) above the cylinder head, an eccentric cam 37 fixed to the drive shaft 36 by press-fitting or the like, A control shaft 38 that is rotatably supported by the same cam bracket at a position above the drive shaft 36 and is arranged in parallel with the drive shaft 36, and is swingably supported by an eccentric cam portion 39 of the control shaft 38. A rocker arm 40 and a swing cam 43 that abuts against a tappet 42 disposed at the upper end of each intake valve 41 are provided. The eccentric cam 37 and the rocker arm 40 are linked by a ring-shaped link 44, and the rocker arm 40 and the swing cam 43 are linked by an arm-shaped link 45.

  As will be described later, the drive shaft 36 is driven by a crankshaft of an engine via a timing chain or a timing belt. The eccentric cam 37 has a circular outer peripheral surface, and the center of the outer peripheral surface is offset from the shaft center of the drive shaft 36 by a predetermined amount, and the annular portion 44a of the ring-shaped link 44 rotates on the outer peripheral surface. It is possible to fit. The rocker arm 40 is supported at its substantially central portion by the eccentric cam portion 39, and an extension portion 44 b of the ring-shaped link 44 is linked to one end thereof, and the arm-shaped link is connected to the other end. The upper end of 45 is linked. The eccentric cam portion 39 is eccentric from the axis of the control shaft 38, and therefore the rocking center of the rocker arm 40 changes according to the angular position of the control shaft 38.

  The swing cam 43 is rotatably supported by being fitted to the outer periphery of the drive shaft 36, and the lower end portion of the arm-shaped link 45 is linked to the end portion extending sideways. On the lower surface of the swing cam 43, a base circle surface concentric with the drive shaft 36 and a cam surface extending in a predetermined curve from the base circle surface to the end are continuously provided. These base circle surfaces and cam surfaces are in contact with the upper surface of the tappet 42 in accordance with the swing position of the swing cam 43. That is, the base circle surface is a section where the lift amount becomes 0 as a base circle section, and when the swing cam 43 swings and the cam surface contacts the tappet 42, the base circle section lifts gradually. A slight ramp section is provided between the base circle section and the lift section.

  As shown in FIGS. 1 and 2, the control shaft 38 is configured to rotate within a predetermined angle range by a lift / operation angle control hydraulic actuator 46 provided at one end. The hydraulic pressure supply to the lift / operating angle control hydraulic actuator 46 is controlled by the first hydraulic control unit 47 based on a control signal from the engine control unit 30.

  The operation of the variable lift / operating angle mechanism 34 will be described. When the drive shaft 36 rotates, the ring-shaped link 44 moves up and down by the cam action of the eccentric cam 37, and the rocker arm 40 swings accordingly. The swing of the rocker arm 40 is transmitted to the swing cam 43 via the arm-shaped link 45, and the swing cam 43 swings. The tappet 42 is pressed by the cam action of the swing cam 43, and the intake valve 41 is lifted. Here, when the angle of the control shaft 38 changes via the lift / operating angle control hydraulic actuator 46, the initial position of the rocker arm 40 changes, and consequently, the initial swing position of the swing cam 43 changes. Since the initial position of the eccentric cam portion 39 can be continuously changed, the valve lift characteristic changes continuously as shown in FIG. That is, the lift and the operating angle can be continuously expanded and contracted simultaneously. In this embodiment, the opening timing and closing timing of the intake valve 41 change substantially symmetrically with the change in the lift and operating angle.

  Next, as shown in FIG. 2, the phase varying mechanism 35 is configured so that the sprocket 48 provided at the front end of the drive shaft 36 and the sprocket 48 and the drive shaft 36 are relatively moved within a predetermined angle range. And a hydraulic actuator 49 for phase control that is rotated to the right. The sprocket 48 is linked to the crankshaft via a timing chain or timing belt (not shown). The hydraulic pressure supply to the phase control hydraulic actuator 49 is controlled by the second hydraulic pressure control unit 50 based on a control signal from the engine control unit 30. By the hydraulic control to the phase control hydraulic actuator 49, the sprocket 48 and the drive shaft 36 are relatively rotated, and the lift center angle is retarded. That is, the lift characteristic curve itself does not change, and the whole advances or retards. This change can also be obtained continuously. The phase variable mechanism 35 is not limited to a hydraulic type, and various configurations such as those using an electromagnetic actuator are possible.

  As described above, according to the variable valve operating apparatus 31 combining the lift / operating angle variable mechanism 34 and the phase variable mechanism 35, both the intake valve opening timing and the intake valve closing timing can be arbitrarily controlled independently. Is possible. As for the control of the lift / working angle variable mechanism 34 and the phase variable mechanism 35, a sensor for detecting the actual lift / working angle or phase may be provided to perform closed loop control, or depending on the operating conditions. It is also possible to simply perform open loop control.

  FIG. 4 shows an example of control of the intake valve opening / closing timing by the variable valve gear 31 under typical operating conditions. Since the variable valve mechanism is not provided on the exhaust valve side, the exhaust valve opening / closing timing is always constant, that is, the exhaust valve closing timing (EVC) is fixed near the top dead center, and the exhaust valve opening timing is fixed. The timing (EVO) is fixed near the bottom dead center, and the actual expansion ratio is always constant. As shown in the figure, when the load is low, a small operating angle is set and the lift center angle is advanced. Therefore, the intake valve closing timing (IVC) has a characteristic (about 90 ° ATDC) that is considerably earlier than the bottom dead center. Such a significant reduction in the actual compression ratio is a setting that can be made in combination with a long stroke, which will be described later, and fuel efficiency can be significantly improved by reducing pump loss. In the middle load range, the variable valve gear 31 is controlled so that the intake valve closing timing approaches the bottom dead center because it is necessary to increase the intake charging efficiency. Therefore, the geometric compression ratio is gradually decreased so as to avoid the occurrence of knocking.

  At the time of high load including full opening, in order to maximize the charging efficiency, the operating angle is sufficiently expanded, the intake valve opening timing is set near the top dead center, and the intake valve closing timing is set near the bottom dead center. Therefore, since the actual compression ratio tends to be high, the geometric compression ratio by the variable compression ratio mechanism is controlled to be lower in order to avoid knocking.

  Here, in the above example, at the time of partial load such as low and medium loads, the operation angle is set to 180 ° or less and the intake valve closing timing is advanced from the intake bottom dead center (so-called early closing). However, the operation angle may be set to 180 ° or more, and the intake valve closing timing may be retarded from the bottom dead center (so-called delayed closing). In this way, in the partial load region, the actual compression ratio becomes smaller than the actual expansion ratio due to so-called mirror cycling that lowers the actual compression ratio. That is, the relationship is “(actual compression ratio / actual expansion ratio) <1”. Here, the actual compression ratio is a compression ratio when compression is performed up to the piston top dead center on the assumption that the compression starts from the intake valve closing timing (IVC), and the actual expansion ratio is expanded from the piston top dead center. Thus, the expansion ratio assumes that expansion ends at the exhaust valve opening timing (EVO).

  FIG. 5 is a configuration explanatory view showing the configuration of the variable compression ratio mechanism 32. The mechanism 32 includes a multi-link type piston-crank mechanism mainly composed of the lower link 4, the upper link 5, and the control link 10. The crankshaft 1 includes a plurality of journal portions 2 and a crankpin 3, and the journal portion 2 is rotatably supported by a main bearing of the cylinder block 18. The crank pin 3 is eccentric from the journal portion 2 by a predetermined amount, and a lower link 4 is rotatably connected thereto. The counterweight 15 extends from the crank web 16 connecting the journal portion 2 and the crankpin 3 to the opposite side of the crankpin 3. The counter weight 15 is provided on both sides of the crank pin 3 so as to face each other, and an outer peripheral portion thereof is formed in an arc shape centering on the journal portion 2.

  The lower link 4 is configured to be split into two left and right members, and the crank pin 3 is fitted in a substantially central connecting hole. The upper link 5 has a lower end side rotatably connected to one end of the lower link 4 by a first connecting pin 6, and an upper end side rotatably connected to a piston 8 by a piston pin 7. The piston 8 receives combustion pressure and reciprocates in the cylinder 19 of the cylinder block 18. The control link 10 that constrains the movement of the lower link 4 is pivotally connected to the other end of the lower link 4 at the upper end side by the second connecting pin 11, and the lower end side becomes a part of the engine body via the control shaft 12. The lower part of the cylinder block 18 is rotatably connected. Specifically, the control shaft 12 is rotatably supported by the engine body and has an eccentric cam portion 12a that is eccentric from the center of rotation, and the lower end portion of the control link 10 is rotatable on the eccentric cam portion 12a. Is fitted.

  As shown in FIG. 1, the rotation position of the control shaft 12 is controlled by a compression ratio control actuator 51 that operates based on a control signal from the engine control unit 30. In the variable compression ratio mechanism using the multi-link type piston-crank mechanism as described above, when the control shaft 12 is rotated by the compression ratio control actuator, the center position of the eccentric cam portion 12a, particularly with respect to the engine main body. The relative position changes. Thereby, the rocking | fluctuation support position of the lower end of the control link 10 changes. When the swing support position of the control link 10 changes, the stroke of the piston 8 changes, and the position of the piston 8 at the piston top dead center (TDC) becomes higher or lower. This makes it possible to change the geometric compression ratio.

  Further, in the above-described multi-link variable compression ratio mechanism, a piston stroke characteristic close to simple vibration can be obtained by appropriately selecting a link dimension. In particular, as shown in FIG. 13, it is possible to obtain characteristics closer to simple vibration compared to the piston stroke characteristics of a general single link type piston-crank mechanism. As a result, the piston acceleration is leveled, and the maximum inertial force near the top dead center of the piston is greatly reduced. According to the piston stroke characteristics close to the simple vibration described above, the speed of the piston 8 near the top dead center is slowed by nearly 20% as compared with that of the single link type piston-crank mechanism.

  Next, the structure of the piston 8 and the upper link 5 will be described. 6 to 9 show the structure of the piston 8 used in the internal combustion engine of the present invention. The piston 8 is integrally cast using an aluminum alloy, and a plurality of, for example, three piston ring grooves 22 are formed on the outer peripheral surface of a piston head 21 having a relatively thick disk shape. A skirt portion 23 is formed on a part of the piston 8 in the circumferential direction that is the thrust-anti-thrust direction so as to extend from the outer circumferential surface along the cylindrical surface. As shown in FIG. 9, the skirt portion 23 has a substantially rectangular projection when viewed from a direction orthogonal to the piston pin 7, and the width along the piston pin axial direction is substantially equal to the entire length of the piston pin 7. It is equal or shorter than the total length of the piston pin 7. That is, the skirt portion 23 is provided in a very small range in the circumferential direction.

  A pair of pin bosses 24 are formed in the center of the piston 8, that is, in the center of the back surface of the piston head 21 having a disk shape, and the end of the piston pin 7 is rotatable on the pin boss 24. A pin hole 25 to be fitted is formed through. A pair of oil grooves 26 along the axial direction are formed on the inner periphery of the pin hole 25. On the other hand, the upper link 5 is made of steel, and as shown in FIG. 10, a piston pin 7 is press-fitted into one end on the piston 8 side. Further, as shown in FIG. 12, the other end of the upper link 5 connected to the lower link 4 branches into a bifurcated shape and supports both ends of the first connecting pin 6.

  Here, the axial length of the upper piston pin 7 in the upper link 5 and the axial length of the lower first connecting pin 6 are equal to each other. Further, since the load received by the piston pin 7 and the load received by the first connecting pin 6 are basically equal, the piston pin 7 and the first connecting pin 6 can have the same diameter. As shown in FIGS. 10 and 12, the dimension of the piston coupling structure including the pair of pin boss portions 24 and the piston pin 7 in the piston pin axial direction is considerably smaller than the diameter of the piston 8 or the cylinder 19. It has become.

  When the piston 8 is in the vicinity of bottom dead center, the outermost diameter portion of the counterweight 15 of the crankshaft 1 intersects with an extension line extending the piston pin 7 in the axial direction as shown in the figure. ing. In other words, when the piston 8 is in the vicinity of the bottom dead center, the outermost diameter portion of the counterweight 15 passes through the side of the pin boss portion 24 holding the piston pin 7.

  FIG. 11 shows the piston stroke from the top dead center to the bottom dead center when the conventional general single link type piston-crank mechanism 101 and the piston 102 are combined for comparison. As is apparent from a comparison between FIG. 12 and FIG. 12, in the configuration of the above embodiment, the piston stroke from the top dead center to the bottom dead center is greatly increased, and the displacement can be increased. For example, the piston stroke can be expanded by about 20%.

  Further, as apparent from FIG. 12, since the skirt portion 23 is also downsized, the counterweight 15 does not interfere with the skirt portion 23 when passing the side of the pin boss portion 24 as described above. Absent. If the skirt portion 23 is reduced in size as described above, it is difficult to ensure a large rigidity. However, in the multi-link type piston-crank mechanism on which the present invention is based, a side thrust that acts to tilt the piston 8 is used. Since the load is smaller than that in the case of a general single link type piston-crank mechanism, the skirt portion 23 only needs to have a minimum size. Specifically, the maximum combustion pressure acts on the piston 8 in the first half of the expansion stroke. At this time, the upper link 5 is in a posture close to vertical and the inclination of the cylinder 19 with respect to the axis is very small. In particular, the inclination of the cylinder 19 with respect to the axis can be made smaller than the posture of the connecting rod in the case of a single link type piston-crank mechanism. Accordingly, the side thrust load is reduced, and the skirt portion 23 can be downsized.

  Furthermore, as an advantage of the above-mentioned double link type piston-crank mechanism, the piston acceleration is leveled by making the piston-stroke characteristic close to simple vibration, and the maximum inertial force near the top dead center of the piston is greatly reduced. . Therefore, as described above, the pin boss portion 24 that holds the piston pin 7 can be downsized. In the configuration using the single link type piston-crank mechanism 101 shown in FIG. 11, assuming that the crank radius of the crankpin is increased and the piston stroke is made longer, the side thrust load acting on the piston 102 is assumed. As the size of the skirt increases, it is difficult to reduce the size of the skirt, and it is difficult to establish a practical engine.

  The present invention is suitable for an in-line four-cylinder engine. In general, in the case of an in-line four-cylinder engine, the inertial secondary vibration of the piston 8 increases rapidly with the expansion of the piston stroke. Therefore, if the displacement is increased by increasing the stroke, the noise vibration characteristics deteriorate and the quality is significantly impaired. However, the multi-link type piston-crank mechanism used in the present invention has a piston stroke characteristic close to a single vibration, so that such a deterioration of the noise vibration characteristic can be avoided. Moreover, if the piston stroke characteristics are close to simple vibration, the speed of the piston 8 near the top dead center is slower than that of the single link type piston-crank mechanism. Thus, good combustion can be ensured even in a combustion chamber having a large displacement per cylinder.

  The piston 8 illustrated in FIGS. 6 to 9 has two compression rings as a piston ring, whereas the piston 8 illustrated in FIG. 10 has a single compression ring. . Since the weight of the lower half of the piston 8 in which the pin boss portion 24 and the piston pin 7 of the present invention are miniaturized becomes very light, the configuration with a single compression ring stabilizes the behavior of the piston 8 around the piston pin 7. Is advantageous. Further, it is advantageous in terms of expanding the piston stroke, which is the original object of the present invention.

  FIG. 14 shows an example of a geometrical compression ratio control target value corresponding to the intake valve closing timing (IVC). The geometric compression ratio by the variable compression ratio mechanism 32 is basically variably controlled in accordance with the engine load so that the compression ratio is high when the load is low and the compression ratio is low when the load is high. Actually, based on signals from a knock sensor that detects knocking, a temperature sensor (not shown) that detects cylinder wall temperature, etc., the geometric compression ratio has a margin in anticipation of the temperature rise due to heat storage. And the actual compression ratio according to the intake valve closing timing is corrected and controlled.

  Referring to FIG. 14, in order to keep the effective compression ratio (actual compression ratio) substantially constant, it is necessary to increase the geometric compression ratio in an accelerated manner as the intake valve closing timing (IVC) is advanced. Therefore, for example, when the IVC is about 90 ° ATDC as in a low load, the actual compression ratio is halved compared to when the IVC is at the top dead center. (For example, about 20) If the geometric compression ratio is simply increased without increasing the stroke as in the present embodiment, the shape of the combustion chamber becomes extremely flat, and cooling near the compression top dead center is achieved. The effect of increasing the loss and recovering the compression temperature by increasing the compression ratio reaches its peak.

  FIG. 15 shows changes in the heat flux (instantaneous cooling loss) from the combustion gas and the surface temperature of the combustion chamber. As can be seen from this figure, the heat flux tends to increase intensively in a very short period after the top dead center where combustion peaks. In such a very short period after top dead center, the cylinder surface is hardly exposed and does not contribute much to cooling, so the surfaces facing the combustion chamber side of the piston and cylinder head are the main causes of cooling loss. Become. Accordingly, when only the piston stroke amount is increased by, for example, about 20% without changing the bore diameter at a predetermined compression ratio, the S / V (surface area / volume) ratio immediately after the top dead center at which the cooling loss reaches a peak is the exhaust amount. If the surface area that contributes to cooling is not changed while it increases by about 20%, it will substantially decrease by about 17%. Actually, it was confirmed that the cooling loss reduction effect (at the time of the combustion peak) close to this ratio was obtained.

  In contrast to such a high compression ratio due to a long stroke, for example, a high compression ratio by flattening the shape of the combustion chamber does not improve the thermal efficiency as expected. The reason is that the cooling loss increases when the shape of the combustion chamber is made flat as the compression ratio increases. That is, as the compression ratio is increased, the expansion ratio is increased and the exhaust temperature is lowered. As a result, the exhaust energy is reduced and the exhaust loss is reduced. In this case, since the cooling loss also increases, the reduction in the exhaust loss is offset by the increase in the cooling loss, so that improvement in the overall thermal efficiency cannot be expected. In addition, since the cooling loss increases, a temperature decrease accompanying a decrease in the actual compression ratio is promoted, and it is difficult to ensure combustion stability.

  On the other hand, when the compression ratio is increased by increasing the stroke, the cooling loss can be effectively reduced because the compression ratio can be increased without causing flattening of the combustion chamber. In addition to the improvement in temperature, it is possible to effectively suppress the temperature decrease accompanying the decrease in the actual compression ratio. Therefore, by combining such a long stroke and mirror cycle, the actual compression ratio is greatly reduced and the pump loss is greatly reduced while ensuring combustion stability at partial load (low / medium load range). It becomes possible to do.

  However, when the stroke is increased, there is an essential problem that the output at high speed does not extend. This is because the intake flow velocity in the vicinity of the intake valve reaches the sonic velocity early due to the longer stroke. Increasing the size of the intake / exhaust valve is difficult because it is limited by the bore diameter of the cylinder. Also for such a problem, the combination with the variable valve operating device 31, particularly the variable lift control of the intake valve by the variable lift / operating angle mechanism 34, is very effective. That is, by making a large lift at high speed, the critical flow rate at which the flow in the vicinity of the intake valve becomes sonic can be greatly increased. Therefore, according to the combination of the long stroke and the variable lift control of the intake valve, the output performance during high-speed rotation can be greatly improved.

  FIG. 16 shows an example of the heat insulating piston. In addition to the above-mentioned long stroke, if the crown 8 of the piston 8 constituting a part of the wall surface of the combustion chamber is coated with a heat insulating material 8A made of a nonmetallic material such as ceramic, the effect of reducing the cooling loss is further improved. growing. However, when the heat insulating material 8A such as ceramic is coated, the temperature becomes high at a high load and the like, and knocking is likely to occur. Combination with variable control of the compression ratio is necessary.

BRIEF DESCRIPTION OF THE DRAWINGS The structure explanatory drawing of the internal combustion engine which concerns on one Example of this invention. The perspective view which shows simply the variable valve apparatus in this Example. The characteristic view which shows the characteristic change of the lift and working angle by a lift and working angle variable mechanism. The characteristic view which shows the example of the variable control of the actual compression ratio by the said variable valve apparatus. Structure explanatory drawing which shows the variable compression ratio mechanism in this execution example. Sectional drawing of the piston along the surface orthogonal to a crankshaft. Sectional drawing of the piston along a crankshaft axial direction. The perspective view which notches and shows a part of piston. The side view of a piston. Explanatory drawing which shows the positional relationship of the piston and counterweight in a bottom dead center. Explanatory drawing of the piston stroke in the conventional piston-crank mechanism. Explanatory drawing of the piston stroke in an Example. The characteristic view which shows the piston-stroke characteristic of an Example. The characteristic view which shows an example of the control target value of the geometric compression ratio corresponding to intake valve closing timing. Explanatory diagram showing combustion chamber surface temperature and change in instantaneous heat flux with respect to crank angle The cross-sectional view which shows an example of the piston which coated the crown surface of the piston which comprises a part of wall surface of a combustion chamber with the nonmetallic material.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 ... Crankshaft 4 ... Lower link 5 ... Upper link 7 ... Piston pin 8 ... Piston 10 ... Control link 15 ... Counterweight 23 ... Skirt part 24 ... Pin boss part 31 ... Variable valve gear 32 ... Variable compression ratio mechanism (double link) Type piston-crank mechanism)

Claims (7)

A piston-crank mechanism that connects a piston that reciprocates in a cylinder and a crankshaft, and a variable valve device that can variably control at least the closing timing of the intake valve;
An internal combustion engine configured so that the outermost diameter portion of the counterweight passes through a side of a pin boss portion holding the piston pin when the piston is near bottom dead center.   2. The internal combustion engine according to claim 1, wherein a stroke amount of the piston is made larger than a diameter of the piston.   The intake valve closing timing is controlled so that an actual compression ratio determined by the intake valve closing timing is smaller than an actual expansion ratio determined by the exhaust valve opening timing at least during partial load. 2. The internal combustion engine according to 2.   The piston-crank mechanism has an upper link whose one end is connected to the piston via a piston pin, and the other end of the upper link is connected via a first connecting pin, and is rotatable to the crank pin of the crankshaft. And a control link having one end connected to the lower link via a second connecting pin and the other end supported to be swingable with respect to the internal combustion engine body. The internal combustion engine according to any one of claims 1 to 3, wherein the internal combustion engine is a piston-crank mechanism.   The control link further includes a support position varying means for displacing the swing support position of the other end of the control link with respect to the internal combustion engine body, and the geometric compression ratio is variably controlled by the displacement of the swing support position. The internal combustion engine according to claim 4.   6. The internal combustion engine according to claim 5, wherein the geometric compression ratio is variably controlled in accordance with the engine load so that the compression ratio is high when the load is low and the compression ratio is low when the load is high.   The internal combustion engine according to any one of claims 1 to 6, wherein a part of the wall surface of the combustion chamber is made of a nonmetallic material.

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