JP2008274962A - Spark ignition internal combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To prevent an occurrence of a pumping loss in a high expansion ratio. <P>SOLUTION: This spark ignition internal combustion engine has: a variable compression ratio mechanism A changing a mechanical compression ratio; and a variable valve timing mechanism B individually controlling valve opening timing and valve closing timing of an intake valve. The mechanical compression ratio is maximized so that the maximum expansion ratio is acquired in engine low load operation. The valve opening timing IO of the intake valve 7 is maintained in a target valve opening timing in a noninterfering area of not causing interference with a piston 4 in a substantially intake top dread center in a period of maximizing the mechanical compression ratio. <P>COPYRIGHT: (C)2009,JPO&INPIT

Description

  The present invention relates to a spark ignition internal combustion engine.

It has a variable compression ratio mechanism that can change the mechanical compression ratio and a variable valve timing mechanism that can individually control the opening timing and closing timing of the intake valve, and the intake valve closes as the engine load decreases. A spark ignition internal combustion engine in which the timing is moved away from the intake bottom dead center and the mechanical compression ratio is increased is known (see, for example, Patent Document 1). In this internal combustion engine, at the time of idling operation, the intake valve is opened after the intake top dead center is considerably exceeded, and is closed after a short valve opening period.
JP 2002-285876 A

  Generally speaking, in an internal combustion engine, the lower the engine load, the worse the thermal efficiency. Therefore, in order to improve the thermal efficiency when the vehicle is running, that is, to improve fuel efficiency, improve the thermal efficiency during engine low load operation. It will be necessary. By the way, in the internal combustion engine, the larger the expansion ratio, the longer the period during which the pressing force acts on the piston during the expansion stroke. Therefore, the greater the expansion ratio, the better the thermal efficiency. On the other hand, when the mechanical compression ratio is increased, the expansion ratio is increased. Therefore, in order to improve the thermal efficiency during vehicle travel, it is preferable to increase the mechanical compression ratio during engine low load operation as much as possible so that the maximum expansion ratio can be obtained during engine low load operation. .

  By the way, if the mechanical compression ratio is increased, the combustion chamber volume at the intake top dead center is reduced. Therefore, if the intake valve is opened too early with respect to the intake top dead center, the intake valve interferes with the piston top surface. Is produced. Therefore, when the mechanical compression ratio is increased, the intake valve must be opened in a non-interference area where interference with the piston does not occur so as not to cause such a problem. In this case, when the intake valve opens after the intake top dead center, there is usually no interference between the intake valve and the piston, and the interference between the intake valve and the piston occurs because the intake valve opens before the intake top dead center. Is the time. Therefore, in order to prevent the intake valve and the piston from interfering with each other, it is necessary to open the intake valve within a non-interference area before the intake top dead center or after the intake top dead center.

  In this case, if the intake valve is opened after the intake top dead center, the pressure in the combustion chamber becomes negative until the intake valve is opened, and thus a pumping loss occurs. Accordingly, when the intake valve is opened after a considerable amount of the intake top dead center during idling operation as in the above-described known internal combustion engine, a considerable pumping loss occurs.

As described above, in order to improve the thermal efficiency during vehicle travel, it is preferable to increase the mechanical compression ratio as much as possible so that the maximum expansion ratio can be obtained during engine low load operation. However, if the pumping loss occurs at this time and the thermal efficiency is lowered, the meaning of increasing the mechanical compression ratio is reduced by half.
An object of the present invention is to provide a spark ignition internal combustion engine capable of obtaining a high thermal efficiency by preventing the occurrence of pumping loss while preventing the interference between the intake valve and the piston when the mechanical compression ratio is increased in order to increase the thermal efficiency. Is to provide.

  In order to achieve the above object, the present invention includes a variable compression ratio mechanism capable of changing the mechanical compression ratio, and a variable valve timing mechanism capable of individually controlling the opening timing and closing timing of the intake valve. The lower the engine load is, the more the intake air corresponding to the required load is supplied to the combustion chamber by moving the closing timing of the intake valve away from the intake bottom dead center. In order to obtain an expansion ratio, the mechanical compression ratio is maximized, and at least during the period when the mechanical compression ratio is maximized, the opening timing of the intake valve is almost at the top dead center of the intake, and there is interference with the piston. The target valve opening time in the non-interference area that does not occur is maintained.

  While preventing interference between the intake valve and the piston, the thermal efficiency during traveling of the vehicle can be greatly improved, and thus good fuel consumption can be obtained.

FIG. 1 shows a side sectional view of a spark ignition type internal combustion engine.
Referring to FIG. 1, 1 is a crankcase, 2 is a cylinder block, 3 is a cylinder head, 4 is a piston, 5 is a combustion chamber, 6 is a spark plug disposed at the center of the top surface of the combustion chamber 5, and 7 is intake air. 8 is an intake port, 9 is an exhaust valve, and 10 is an exhaust port. The intake port 8 is connected to a surge tank 12 via an intake branch pipe 11, and a fuel injection valve 13 for injecting fuel into the corresponding intake port 8 is arranged in each intake branch pipe 11. The fuel injection valve 13 may be arranged in each combustion chamber 5 instead of being attached to each intake branch pipe 11.

  The surge tank 12 is connected to the outlet of the compressor 15a of the exhaust turbocharger 15 via an intake duct 14, and the inlet of the compressor 15a is connected to an air cleaner via an intake air amount detector 16 using, for example, heat rays. A throttle valve 19 driven by an actuator 18 is disposed in the intake duct 14.

  On the other hand, the exhaust port 10 is connected to an inlet of an exhaust turbine 15b of an exhaust turbocharger 15 via an exhaust manifold 20, and an outlet of the exhaust turbine 15b is connected to a catalytic converter 22 containing, for example, a three-way catalyst via an exhaust pipe 21. Is done. An air-fuel ratio sensor 23 is disposed in the exhaust pipe 21.

  On the other hand, in the embodiment shown in FIG. 1, the piston 4 is positioned at the compression top dead center by changing the relative position of the crankcase 1 and the cylinder block 2 in the cylinder axial direction at the connecting portion between the crankcase 1 and the cylinder block 2. There is provided a variable compression ratio mechanism A capable of changing the volume of the combustion chamber 5 when performing the operation, and the valve closing timing of the intake valve 7 can be controlled and the intake air can be controlled in order to change the actual start timing of the compression action. A variable valve timing mechanism B capable of individually controlling the valve opening timing of the valve 7 is provided.

  The electronic control unit 30 is composed of a digital computer, and is connected to each other by a bidirectional bus 31. A ROM (read only memory) 32, a RAM (random access memory) 33, a CPU (microprocessor) 34, an input port 35 and an output port 36. It comprises. The output signal of the intake air amount detector 16 and the output signal of the air-fuel ratio sensor 23 are input to the input port 35 via the corresponding AD converters 37, respectively. A load sensor 41 that generates an output voltage proportional to the depression amount L of the accelerator pedal 40 is connected to the accelerator pedal 40, and the output voltage of the load sensor 41 is input to the input port 35 via the corresponding AD converter 37. Is done. Further, a crank angle sensor 42 that generates an output pulse every time the crankshaft rotates, for example, 30 ° is connected to the input port 35. On the other hand, the output port 36 is connected to the spark plug 6, the fuel injection valve 13, the throttle valve driving actuator 18, the variable compression ratio mechanism A, and the variable valve timing mechanism B through corresponding drive circuits 38.

  2 is an exploded perspective view of the variable compression ratio mechanism A shown in FIG. 1, and FIG. 3 is a side sectional view of the internal combustion engine schematically shown. Referring to FIG. 2, a plurality of protrusions 50 spaced from each other are formed below both side walls of the cylinder block 2, and cam insertion holes 51 each having a circular cross section are formed in each protrusion 50. Has been. On the other hand, a plurality of protrusions 52 are formed on the upper wall surface of the crankcase 1 so as to be fitted between the corresponding protrusions 50 spaced apart from each other. Cam insertion holes 53 each having a circular cross section are formed.

  As shown in FIG. 2, a pair of camshafts 54 and 55 are provided, and on each camshaft 54 and 55, a circular cam 56 is rotatably inserted into each cam insertion hole 51. It is fixed. These circular cams 56 are coaxial with the rotational axes of the camshafts 54 and 55. On the other hand, an eccentric shaft 57 arranged eccentrically with respect to the rotation axis of each camshaft 54, 55 extends between the circular cams 56 as shown by hatching in FIG. A cam 58 is eccentrically mounted for rotation. As shown in FIG. 2, the circular cams 58 are disposed between the circular cams 56, and the circular cams 58 are rotatably inserted into the corresponding cam insertion holes 53.

  When the circular cams 56 fixed on the camshafts 54 and 55 are rotated in opposite directions as indicated by solid arrows in FIG. 3A from the state shown in FIG. In order to move toward the lower center, the circular cam 58 rotates in the opposite direction to the circular cam 56 in the cam insertion hole 53 as shown by the broken arrow in FIG. 3A, and is shown in FIG. As described above, when the eccentric shaft 57 moves to the lower center, the center of the circular cam 58 moves below the eccentric shaft 57.

  3A and 3B, the relative positions of the crankcase 1 and the cylinder block 2 are determined by the distance between the center of the circular cam 56 and the center of the circular cam 58. The cylinder block 2 moves away from the crankcase 1 as the distance between the center and the center of the circular cam 58 increases. When the cylinder block 2 moves away from the crankcase 1, the volume of the combustion chamber 5 increases when the piston 4 is positioned at the compression top dead center. Therefore, by rotating the camshafts 54 and 55, the piston 4 is compressed at the top dead center. The volume of the combustion chamber 5 when it is located at can be changed.

  As shown in FIG. 2, in order to rotate the camshafts 54 and 55 in the opposite directions, a pair of worm gears 61 and 62 having opposite spiral directions are attached to the rotation shaft of the drive motor 59, respectively. Gears 63 and 64 meshing with the worm gears 61 and 62 are fixed to the end portions of the camshafts 54 and 55, respectively. In this embodiment, by driving the drive motor 59, the volume of the combustion chamber 5 when the piston 4 is located at the compression top dead center can be changed over a wide range. The variable compression ratio mechanism A shown in FIGS. 1 to 3 shows an example, and any type of variable compression ratio mechanism can be used.

  4 shows a variable valve timing mechanism B provided for the camshaft 70 for driving the intake valve 7 in FIG. As shown in FIG. 4, the variable valve timing mechanism B is attached to one end of the camshaft 70 to change the cam phase of the camshaft 70, and the camshaft 70 and intake valve 7 valve lifters. The cam operating angle changing portion B2 is disposed between the cam shafts 70 and changes the cam operating angle of the camshaft 70 to a different operating angle and transmits it to the intake valve 7. FIG. 4 shows a side sectional view and a plan view of the cam working angle changing portion B2.

  First, the cam phase changing portion B1 of the variable valve timing mechanism B will be described. The cam phase changing portion B1 is rotated by a crankshaft of the engine via a timing belt in a direction indicated by an arrow, a timing pulley 71, A cylindrical housing 72 that rotates together, a rotating shaft 73 that rotates together with the camshaft 70 and that can rotate relative to the cylindrical housing 72, and an outer peripheral surface of the rotating shaft 73 from an inner peripheral surface of the cylindrical housing 72 And a plurality of partition walls 74 extending between the partition walls 74 and the outer peripheral surface of the rotating shaft 73 to the inner peripheral surface of the cylindrical housing 72. An advance hydraulic chamber 76 and a retard hydraulic chamber 77 are formed respectively.

  The hydraulic oil supply control to the hydraulic chambers 76 and 77 is performed by a hydraulic oil supply control valve 78. The hydraulic oil supply control valve 78 includes hydraulic ports 79 and 80 connected to the hydraulic chambers 76 and 77, a hydraulic oil supply port 82 discharged from the hydraulic pump 81, a pair of drain ports 83 and 84, And a spool valve 85 for controlling communication between the ports 79, 80, 82, 83, and 84.

  When the cam phase of the camshaft 70 is to be advanced, the spool valve 85 is moved downward in FIG. 4, and the hydraulic oil supplied from the supply port 82 enters the advance angle hydraulic chamber 76 via the hydraulic port 79. The hydraulic oil in the retarding hydraulic chamber 77 is supplied and discharged from the drain port 84. At this time, the rotation shaft 73 is rotated relative to the cylindrical housing 72 in the arrow X direction.

  On the other hand, when the cam phase of the camshaft 70 is to be retarded, the spool valve 85 is moved upward in FIG. 4 and the hydraulic oil supplied from the supply port 82 is used for retarding via the hydraulic port 80. The hydraulic oil in the advance hydraulic chamber 76 is discharged from the drain port 83 while being supplied to the hydraulic chamber 77. At this time, the rotating shaft 73 is rotated relative to the cylindrical housing 72 in the direction opposite to the arrow X.

  If the spool valve 85 is returned to the neutral position shown in FIG. 4 while the rotation shaft 73 is rotated relative to the cylindrical housing 72, the relative rotation operation of the rotation shaft 73 is stopped, and the rotation shaft 73 is The relative rotation position at that time is held. Therefore, the cam phase changing portion B1 can advance or retard the cam phase of the camshaft 70 by a desired amount. That is, the valve opening timing of the intake valve 7 can be arbitrarily advanced or retarded by the cam phase changing unit B1.

  Next, the cam working angle changing portion B2 of the variable valve timing mechanism B will be described. The cam working angle changing portion B2 is arranged in parallel with the camshaft 70 and is moved in the axial direction by the actuator 91; An intermediate cam 94 engaged with the cam 92 of the camshaft 70 and slidably fitted in an axially extending spline 93 formed on the control rod 90 and a valve lifter for driving the intake valve 7 And a swing cam 96 slidably fitted to a spirally extending spline 95 formed on the control rod 90. A cam 97 is formed on the swing cam 96. Has been.

  When the camshaft 90 is rotated, the intermediate cam 94 is always swung by a certain angle by the cam 92. At this time, the rocking cam 96 is also swung by a certain angle. On the other hand, the intermediate cam 94 and the swing cam 96 are supported so as not to move in the axial direction of the control rod 90. Therefore, when the control rod 90 is moved in the axial direction by the actuator 91, the swing cam 96 is intermediate. It is rotated relative to the cam 94.

  When the cam 97 of the swing cam 86 starts to engage with the valve lifter 24 when the cam 92 of the cam shaft 70 starts to engage with the intermediate cam 94 due to the relative rotational positional relationship between the intermediate cam 94 and the swing cam 96. As shown by a in FIG. 5B, the valve opening period and the lift of the intake valve 7 become the largest. On the other hand, when the swing cam 96 is rotated relative to the intermediate cam 94 in the direction of the arrow Y by the actuator 91, the cam 92 of the camshaft 70 is engaged with the intermediate cam 94 for a while. Thereafter, the cam 97 of the swing cam 96 is engaged with the valve lifter 24. In this case, as indicated by b in FIG. 5B, the valve opening period and the lift amount of the intake valve 7 are smaller than a.

  When the swing cam 96 is further rotated relative to the intermediate cam 94 in the direction of the arrow Y in FIG. 4, the valve opening period and the lift amount of the intake valve 7 are further reduced as indicated by c in FIG. . That is, the valve opening period of the intake valve 7 can be arbitrarily changed by changing the relative rotational position of the intermediate cam 94 and the swing cam 96 by the actuator 91. However, in this case, the lift amount of the intake valve 7 becomes smaller as the opening period of the intake valve 7 becomes shorter.

  Thus, the cam phase changing unit B1 can arbitrarily change the valve opening timing of the intake valve 7, and the cam operating angle changing unit B2 can arbitrarily change the valve opening period of the intake valve 7, so that the cam phase By both the change part B1 and the cam working angle change part B2, that is, by the variable valve timing mechanism B, the valve opening timing and the valve opening period of the intake valve 7, that is, the valve opening timing and the valve closing timing of the intake valve 7 are set. It can be changed arbitrarily.

  The variable valve timing mechanism B shown in FIGS. 1 and 4 shows an example, and various types of variable valve timing mechanisms other than the examples shown in FIGS. 1 and 4 can be used.

  Next, the meanings of terms used in the present application will be described with reference to FIG. 6 (A), (B), and (C) show an engine having a combustion chamber volume of 50 ml and a piston stroke volume of 500 ml for the sake of explanation. , (B), (C), the combustion chamber volume represents the volume of the combustion chamber when the piston is located at the compression top dead center.

  FIG. 6A explains the mechanical compression ratio. The mechanical compression ratio is a value mechanically determined only from the stroke volume of the piston and the combustion chamber volume during the compression stroke, and this mechanical compression ratio is expressed by (combustion chamber volume + stroke volume) / combustion chamber volume. In the example shown in FIG. 6A, this mechanical compression ratio is (50 ml + 500 ml) / 50 ml = 11.

  FIG. 6B describes the actual compression ratio. This actual compression ratio is a value determined from the actual piston stroke volume and the combustion chamber volume from when the compression action is actually started until the piston reaches top dead center, and this actual compression ratio is (combustion chamber volume + actual (Stroke volume) / combustion chamber volume. That is, as shown in FIG. 6B, even if the piston starts to rise in the compression stroke, the compression operation is not performed while the intake valve is open, and the actual compression is performed from the time when the intake valve is closed. The action begins. Therefore, the actual compression ratio is expressed as described above using the actual stroke volume. In the example shown in FIG. 6B, the actual compression ratio is (50 ml + 450 ml) / 50 ml = 10.

  FIG. 6C illustrates the expansion ratio. The expansion ratio is a value determined from the stroke volume of the piston and the combustion chamber volume during the expansion stroke, and this expansion ratio is expressed by (combustion chamber volume + stroke volume) / combustion chamber volume. In the example shown in FIG. 6C, this expansion ratio is (50 ml + 500 ml) / 50 ml = 11.

  Next, the most basic features of the present invention will be described with reference to FIGS. FIG. 7 shows the relationship between the theoretical thermal efficiency and the expansion ratio, and FIG. 8 shows a comparison between a normal cycle and an ultrahigh expansion ratio cycle that are selectively used according to the load in the present invention.

  FIG. 8A shows a normal cycle when the intake valve closes near the bottom dead center and the compression action by the piston is started from the vicinity of the compression bottom dead center. In the example shown in FIG. 8A, the combustion chamber volume is 50 ml, and the stroke volume of the piston is 500 ml, as in the examples shown in FIGS. 6A, 6B, and 6C. As can be seen from FIG. 8A, in a normal cycle, the mechanical compression ratio is (50 ml + 500 ml) / 50 ml = 11, the actual compression ratio is almost 11, and the expansion ratio is also (50 ml + 500 ml) / 50 ml = 11. That is, in a normal internal combustion engine, the mechanical compression ratio, the actual compression ratio, and the expansion ratio are substantially equal.

  The solid line in FIG. 7 shows the change in the theoretical thermal efficiency when the actual compression ratio and the expansion ratio are substantially equal, that is, in a normal cycle. In this case, it can be seen that the theoretical thermal efficiency increases as the expansion ratio increases, that is, as the actual compression ratio increases. Therefore, in order to increase the theoretical thermal efficiency in a normal cycle, it is only necessary to increase the actual compression ratio. However, the actual compression ratio can only be increased to a maximum of about 12 due to the restriction of the occurrence of knocking at the time of engine high load operation, and thus the theoretical thermal efficiency cannot be sufficiently increased in a normal cycle.

  On the other hand, in this situation, the present inventor has studied to increase the theoretical thermal efficiency by strictly dividing the mechanical compression ratio and the actual compression ratio, and as a result, the theoretical thermal efficiency is governed by the expansion ratio, and the theoretical thermal efficiency The actual compression ratio has been found to have little effect. That is, if the actual compression ratio is increased, the explosive force is increased, but a large amount of energy is required for compression. Thus, even if the actual compression ratio is increased, the theoretical thermal efficiency is hardly increased.

  On the other hand, when the expansion ratio is increased, the period during which the pressing force acts on the piston during the expansion stroke becomes longer, and thus the period during which the piston applies the rotational force to the crankshaft becomes longer. Therefore, the larger the expansion ratio, the higher the theoretical thermal efficiency. The broken line in FIG. 7 indicates the theoretical thermal efficiency when the expansion ratio is increased with the actual compression ratio fixed at 10. Thus, the theoretical thermal efficiency when the expansion ratio is increased while the actual compression ratio is maintained at a low value, and the theoretical thermal efficiency when the actual compression ratio is increased with the expansion ratio as shown by the solid line in FIG. It can be seen that there is no significant difference from the amount of increase.

  Thus, if the actual compression ratio is maintained at a low value, knocking does not occur. Therefore, if the expansion ratio is increased while the actual compression ratio is maintained at a low value, the theoretical thermal efficiency is reduced while preventing the occurrence of knocking. Can be greatly increased. FIG. 8B shows an example where the variable compression ratio mechanism A and variable valve timing mechanism B are used to increase the expansion ratio while maintaining the actual compression ratio at a low value.

  Referring to FIG. 8B, in this example, the variable compression ratio mechanism A reduces the combustion chamber volume from 50 ml to 20 ml. On the other hand, the variable valve timing mechanism B delays the closing timing of the intake valve until the actual piston stroke volume is reduced from 500 ml to 200 ml. As a result, in this example, the actual compression ratio is (20 ml + 200 ml) / 20 ml = 11, and the expansion ratio is (20 ml + 500 ml) / 20 ml = 26. In the normal cycle shown in FIG. 8A, the actual compression ratio is almost 11 and the expansion ratio is 11, as described above. Compared with this case, only the expansion ratio is shown in FIG. 8B. It can be seen that it has been increased to 26. This is why it is called an ultra-high expansion ratio cycle.

  As described above, in general, in an internal combustion engine, the lower the engine load, the worse the thermal efficiency. Therefore, in order to improve the thermal efficiency when the vehicle is running, that is, to improve the fuel efficiency, the thermal efficiency during the engine low load operation is reduced. It is necessary to improve. On the other hand, in the ultra-high expansion ratio cycle shown in FIG. 8B, since the actual piston stroke volume during the compression stroke is reduced, the amount of intake air that can be sucked into the combustion chamber 5 is reduced. The expansion ratio cycle can only be adopted when the engine load is relatively low. Therefore, in the present invention, the super high expansion ratio cycle shown in FIG. 8 (B) is used during low engine load operation, and the normal cycle shown in FIG. 8 (A) is used during high engine load operation. This is the basic feature of the present invention.

FIG. 9 shows the overall operation control during steady operation at a low engine speed. The overall operation control will be described below with reference to FIG.
FIG. 9 shows changes in the mechanical compression ratio, the expansion ratio, the closing timing of the intake valve 7, the actual compression ratio, the intake air amount, the opening degree of the throttle valve 17 and the pumping loss according to the engine load. In the embodiment according to the present invention, the average air-fuel ratio in the normal combustion chamber 5 is the air-fuel ratio sensor 23 so that unburned HC, CO and NO _x in the exhaust gas can be simultaneously reduced by the three-way catalyst in the catalytic converter 22. Feedback control to the stoichiometric air-fuel ratio based on the output signal.

  As described above, the normal cycle shown in FIG. 8 (A) is executed during engine high load operation. Accordingly, as shown in FIG. 9, the mechanical compression ratio is lowered at this time, so the expansion ratio is low, and the closing timing of the intake valve 7 is advanced as shown by the solid line in FIG. At this time, the amount of intake air is large, and at this time, the opening degree of the throttle valve 17 is kept fully open or almost fully open, so that the pumping loss is zero.

  On the other hand, as shown in FIG. 9, when the engine load is lowered, the mechanical compression ratio is increased accordingly, and thus the expansion ratio is also increased. Further, at this time, the valve closing timing of the intake valve 7 is delayed as the engine load becomes lower as shown by the solid line in FIG. 9 so that the actual compression ratio is kept substantially constant. At this time as well, the throttle valve 17 is kept fully open or substantially fully open, and therefore the amount of intake air supplied into the combustion chamber 5 changes the closing timing of the intake valve 7 regardless of the throttle valve 17. Is controlled by that. Also at this time, the pumping loss becomes zero.

  As described above, when the engine load is reduced from the engine high load operation state, the mechanical compression ratio is increased as the intake air amount is decreased while the actual compression ratio is substantially constant. That is, the volume of the combustion chamber 5 when the piston 4 reaches the compression top dead center is decreased in proportion to the decrease in the intake air amount. Therefore, the volume of the combustion chamber 5 when the piston 4 reaches the compression top dead center changes in proportion to the intake air amount. At this time, since the air-fuel ratio in the combustion chamber 5 is the stoichiometric air-fuel ratio, the volume of the combustion chamber 5 when the piston 4 reaches the compression top dead center changes in proportion to the fuel amount. Become.

When the engine load is further reduced, the mechanical compression ratio is further increased. When the mechanical compression ratio reaches the limit mechanical compression ratio that is the structural limit of the combustion chamber 5, the engine when the mechanical compression ratio reaches the limit mechanical compression ratio is reached. In the region where the load is lower than the load L ₁ , the mechanical compression ratio is maintained at the limit mechanical compression ratio. Therefore, the mechanical compression ratio is maximized and the expansion ratio is maximized during engine low load operation. In other words, in the present invention, the mechanical compression ratio is maximized so that the maximum expansion ratio can be obtained during engine low load operation. Further, at this time, the actual compression ratio is maintained at substantially the same actual compression ratio as that during the engine medium and high load operation.

On the other hand, as shown by a solid line in FIG. 9, the closing timing of the intake valve 7 is delayed to a limit closing timing that can control the amount of intake air supplied into the combustion chamber 5 as the engine load becomes lower. closing timing of the valve 7 is the closing timing of the intake valve 7 in the region of a load lower than the engine load L ₂ when reaching the limit closing timing is held at the limit closing timing. If the closing timing of the intake valve 7 is held at the limit closing timing, the intake air amount can no longer be controlled by the change in the closing timing of the intake valve 7, so it is necessary to control the intake air amount by some other method There is.

In the embodiment shown in FIG. 9, at this time, that is, in a region where the load is lower than the engine load L ₂ when the closing timing of the intake valve 7 reaches the limit closing timing, the throttle valve 17 supplies the fuel into the combustion chamber 5. The amount of intake air to be controlled is controlled. However, when the intake air amount is controlled by the throttle valve 17, the pumping loss increases as shown in FIG.

In order to prevent such a pumping loss, the throttle valve 17 is kept fully open or almost fully open in a region where the load is lower than the engine load L ₂ when the closing timing of the intake valve 7 reaches the limit closing timing. In this state, the air-fuel ratio can be increased as the engine load decreases. At this time, it is preferable to arrange the fuel injection valve 13 in the combustion chamber 5 and perform stratified combustion.

  As shown in FIG. 9, the actual compression ratio is maintained substantially constant regardless of the engine load during low engine speed. The actual compression ratio at this time is approximately within a range of ± 10%, preferably within a range of ± 5% with respect to the actual compression ratio at the time of engine medium and high load operation. In the embodiment according to the present invention, the actual compression ratio at low engine speed is approximately 10 ± 1, that is, between 9 and 11. However, if the engine speed increases, the air-fuel mixture in the combustion chamber 5 is disturbed, so that knocking does not easily occur. Therefore, in the embodiment according to the present invention, the actual compression ratio increases as the engine speed increases.

  On the other hand, as described above, the expansion ratio is 26 in the ultra-high expansion ratio cycle shown in FIG. This expansion ratio is preferably as high as possible, but if it is 20 or more, a considerably high theoretical thermal efficiency can be obtained. Therefore, in the present invention, the variable compression ratio mechanism A is formed so that the expansion ratio is 20 or more.

  In the example shown in FIG. 9, the mechanical compression ratio is continuously changed according to the engine load. However, the mechanical compression ratio can be changed stepwise according to the engine load.

On the other hand, as shown by the broken line in FIG. 9, the intake air amount can be controlled without depending on the throttle valve 17 by advancing the closing timing of the intake valve 7 as the engine load becomes lower. Accordingly, when expressing the case shown in FIG. 9 so as to include both the case shown by the solid line and the case shown by the broken line, in the embodiment according to the present invention, the valve closing timing of the intake valve 7 becomes smaller as the engine load becomes lower. so that is moved in a direction away from intake bottom dead center BDC until the limit closing timing L ₂ enabling control of the amount of intake air fed into the combustion chamber.

Next, the valve opening timing of the intake valve 7 will be described focusing on the low load operation in which the high expansion ratio cycle shown in FIG. 8B is executed.
FIG. 10A shows the lift change of the intake valve 7 when the mechanical compression ratio is the highest during engine low load operation, the lift change of the exhaust valve 9, and the limit at which the intake valve 7 or the exhaust valve 9 interferes with the piston 4. 10A, when the lift curve of the exhaust valve 9 intersects with the piston interference line in FIG. 10A, the exhaust valve 9 interferes with the piston 4, and the lift curve of the intake valve 7 becomes the piston interference line. When crossed, the intake valve 7 interferes with the piston 4.

  Accordingly, in FIG. 10A, the non-interference area θ with respect to the opening timing of the intake valve 7 where the intake valve 7 does not interfere with the piston 4 is substantially after the intake top dead center (TDC). Therefore, in the present invention, the opening timing of the intake valve 7 is set within this non-interference area θ.

  On the other hand, if the intake valve 7 is opened after the intake top dead center, the pressure in the combustion chamber 5 becomes negative until the intake valve 7 is opened, thus causing a pumping loss. This pumping loss increases as the opening timing of the intake valve 7 is delayed from the intake top dead center. Therefore, in the present invention, the target valve opening timing of the intake valve 7 is maintained substantially at the compression top dead center, preferably slightly before the intake top dead center in the non-interference area θ.

  Further, when the engine compression ratio is the highest during engine low load operation, the amount of intake air to be supplied into the combustion chamber 5 is small. Therefore, at this time, the intake valve 7 is closed as shown by the solid line in FIG. The valve timing is considerably delayed, or the closing timing of the intake valve 7 is considerably advanced as shown by the broken line in FIG. In the embodiment according to the present invention, the closing timing of the exhaust valve 9 is substantially fixed at the intake top dead center.

  On the other hand, FIG. 10B shows the lift change of the intake valve 7, the lift change of the exhaust valve 9, and the piston interference line in a certain operation state at the time of medium-high speed medium and high load operation. Since the mechanical compression ratio becomes small at the time of engine medium and high load operation, the piston interference line rises. Therefore, it is not necessary to pay attention to the interference with the piston 4 at this time. In the operating state shown in FIG. 10B, the opening timing of the intake valve 7 is much earlier than the intake top dead center (TDC), and the closing timing of the intake valve 7 is a solid line in FIG. Compared to the case shown by (1), it is advanced, and compared with the case shown by the broken line in FIG.

  By the way, as described above, in the operating state shown in FIG. 10B, the opening timing of the intake valve 7 is considerably before the intake top dead center, while in the operating state shown in FIG. The opening timing of the valve 7 is almost the intake top dead center. Therefore, when the operating state of the engine changes from the operating state shown in FIG. 10B to the operating state shown in FIG. 10A, the opening timing of the intake valve 7 must be delayed, and the operating state of the engine Is changed from the operating state shown in FIG. 10A to the operating state shown in FIG. 10B, the opening timing of the intake valve 7 must be advanced.

  On the other hand, when the operating state of the engine changes from the operating state shown in FIG. 10B to the operating state shown in FIG. 10A, the intake valve 7 is supplied to the combustion chamber 5 by changing the valve closing timing. Similarly, when the amount of intake air to be reduced is reduced and the mechanical compression ratio is increased, and the operating state of the engine changes from the operating state shown in FIG. 10A to the operating state shown in FIG. 7 is changed, the intake air amount to be supplied into the combustion chamber 5 is increased and the mechanical compression ratio is reduced.

  However, in the case where the intake air amount to be supplied into the combustion chamber 5 is reduced and the mechanical compression ratio is increased as described above, while the intake air amount is not sufficiently reduced, that is, while the intake air amount is large. When the compression ratio is increased, the actual compression ratio is increased, and thus knocking occurs. On the other hand, when the intake air amount to be supplied into the combustion chamber 5 is increased and the mechanical compression ratio is decreased as described above, the actual compression ratio is increased if the intake air amount is increased while the mechanical compression ratio is not decreased. And thus knocking will occur.

  In the embodiment according to the present invention, a time lag is provided between the operation of the variable compression ratio mechanism A and the operation of the variable valve timing mechanism B in order to prevent the occurrence of such knocking. Next, this will be described with reference to FIGS. 11 and 12 by taking as an example a case where the lift amount of the intake valve 7 is the lift amount as shown by the solid line in FIG. 10A during engine low load operation. .

FIG. 11 shows changes in the valve opening timing IO of the intake valve 7 when the operating state of the engine changes from the operating state shown in FIG. 10B to the operating state shown in FIG. The change in the valve closing timing IC and the change in the mechanical compression ratio are shown. In addition, in FIG. 11, (B) has shown the time of the driving | running state shown by FIG. 11 (B), (A) has shown the time of the driving | running state shown by FIG. 11 (A).
11 shows that the intake valve 7 is opened by the variable valve timing mechanism B when the operating state of the engine changes from the operating state shown in FIG. 10B to the operating state shown in FIG. The timing IO changing operation and the valve closing timing IC changing operation are started at the same time and completed at the same time.

  Referring to FIG. 11, when the engine operating state shifts from the medium / high speed medium / high load operation shown in FIG. 10 (B) to the low load operation shown in FIG. 10 (A), in the example shown in FIG. After the valve opening timing IO of the intake valve 7 reaches the target valve opening timing in the non-interference area B, the operation of changing the mechanical compression ratio, that is, the increasing operation is started. On the other hand, in the example shown in (II), the valve opening timing IO of the intake valve 7 becomes the target valve opening timing within the non-interference area θ after the change operation of the valve opening timing IO of the intake valve 7 is started. A change operation of the mechanical compression ratio, that is, an increase operation is started before. In the example shown in (III), the change operation of the mechanical compression ratio is started when the change operation of the valve opening timing IO of the intake valve 7 is started, but the change speed of the mechanical compression ratio at this time is slowed down. ing.

  When the operation of changing the mechanical compression ratio shown in (I), (II), and (III) of FIG. 11 is comprehensively expressed, the operation state of the engine is illustrated from the medium to high speed medium and high load operation shown in FIG. 10 (B). The intake valve is set so that the mechanical compression ratio becomes the maximum after the valve opening timing IO of the intake valve 7 reaches the target valve opening timing in the non-interference area θ when the low load operation shown in FIG. The change operation of the mechanical compression ratio is delayed with respect to the change operation of the valve opening timing IO of 7.

  In this way, when the change operation of the mechanical compression ratio is delayed with respect to the change operation of the valve opening timing IO of the intake valve 7, the mechanical compression ratio increases before the intake air amount supplied into the combustion chamber 5 decreases. Thus, knocking can be prevented from occurring.

FIG. 12 shows changes in the valve opening timing IO of the intake valve 7 when the operating state of the engine changes from the operating state shown in FIG. 10A to the operating state shown in FIG. The change in the valve closing timing IC and the change in the mechanical compression ratio are shown. In FIG. 11, (A) shows the operation state shown in FIG. 11 (A), and (B) shows the operation state shown in FIG. 11 (B).
In FIG. 12, as in FIG. 11, when the engine operating state changes from the operating state shown in FIG. 10 (A) to the operating state shown in FIG. 10 (B), intake by the variable valve timing mechanism B is performed. The operation of changing the valve opening timing IO of the valve 7 and the operation of changing the valve closing timing IC are started at the same time and completed simultaneously.

  Referring to FIG. 12, when the engine operating state shifts from the low load operation shown in FIG. 10 (A) to the medium / high speed medium / high load operation shown in FIG. 10 (B), in the example shown in (I) After the mechanical compression ratio has decreased to the target mechanical compression ratio corresponding to the engine operating state, the operation of changing the valve opening timing IO of the intake valve 7 is started. On the other hand, in the example shown in (II), the operation of changing the valve opening timing IO of the intake valve 7 is started while the mechanical compression ratio is lowered to the target mechanical compression ratio corresponding to the engine operating state. In the example shown in (III), the operation of changing the valve opening timing IO of the intake valve 7 is started while the mechanical compression ratio is lowered to the target mechanical compression ratio corresponding to the engine operating state. When the changing operation of the valve opening timing IO of the valve 7 is started, the changing speed of the mechanical compression ratio, that is, the decreasing speed is slowed down.

  When the operation of changing the mechanical compression ratio shown in (I), (II), and (III) of FIG. 12 is comprehensively expressed, the engine operating state is changed from the low load operation shown in FIG. B) When changing to the medium / high speed / medium / high load operation shown in B), the change operation of the valve opening timing IO of the intake valve 7 is started after the change operation of the mechanical compression ratio to lower the mechanical compression ratio is started. Yes.

  Thus, when the change operation of the opening timing of the intake valve is started after the change operation of the mechanical compression ratio is started, the amount of intake air supplied into the combustion chamber 5 when the mechanical compression ratio is high is increased. In this way, knocking can be prevented from occurring.

  FIG. 13 shows a case where the change amount of the engine load is small, and therefore, the valve opening timing IO of the intake valve 7, the valve closing timing IC of the intake valve 7 and the change amount of the mechanical compression ratio are small. At this time, as shown in FIG. 13, the changing operation of the opening timing IO of the intake valve 7, the changing operation of the closing timing IC of the intake valve 7 and the changing operation of the mechanical compression ratio are started at the same time and completed almost simultaneously.

  Referring to FIG. 14, the target valve opening timing IO of the intake valve 7 is stored in advance in the ROM 32 as a function of the engine load L and the engine speed N in the form of a map as shown in FIG. Further, the target valve closing timing IC of the intake valve 7 necessary for supplying the required intake air amount into the combustion chamber 5 is a function map of the engine load L and the engine speed N as shown in FIG. Is stored in the ROM 32 in advance.

On the other hand, FIG. 14D shows the relationship between the target actual compression ratio and the engine load L for each engine speed N ₁ , N ₂ , N ₃ , N ₄ (N ₁ <N ₂ <N ₃ <N ₄ ). ing. At the time of engine low rotation as shown by N ₁ in FIG. 14 (D) as described above is held substantially constant target actual compression ratio regardless of the engine load L, the target actual compression ratio the higher the engine rotation speed It ’s so expensive. Note that the mechanical compression ratio CR necessary for making the actual compression ratio the target actual compression ratio is stored in advance in the ROM 32 in the form of a map as shown in FIG. 14C as a function of the engine load L and the engine speed N. It is remembered.

Next, the operation control routine will be described with reference to FIG.
Referring to FIG. 15, first, at step 100, it is judged if the engine load L is higher than the load L ₂ shown in FIG. When L ≧ L _2, the routine proceeds to step 101 where the valve opening timing IO of the intake valve 7 is calculated from the map shown in FIG. 14A, and the valve closing timing IC of the intake valve 7 is calculated from the map shown in FIG. Calculated. Next, the routine proceeds to step 104. On the other hand, when it is determined in step 100 that L <L ₂ , the routine proceeds to step 102 where the closing timing of the intake valve 7 is set as the limit closing timing, and then in step 103, the intake air amount is reduced by the throttle valve 19. Be controlled. Next, the routine proceeds to step 104.

Step 104, the engine load L is whether low is determined than the load L ₁ shown in FIG. When L ≧ L _1, the routine proceeds to step 105, where the mechanical compression ratio CR is calculated from the map shown in FIG. Next, the routine proceeds to step 107. On the other hand, when it is judged at step 104 that L <L ₁ , the routine proceeds to step 106 where the mechanical compression ratio CR is made the limit mechanical compression ratio. Next, the routine proceeds to step 107.

  In step 107, it is determined whether or not the absolute value | ΔL | of the engine load change amount ΔL is larger than the set value XL. When | ΔL |> XL, the routine proceeds to step 108, where it is judged if the engine load change amount ΔL is negative or not. When ΔL <0, that is, when the engine load has decreased by the set value XL or more, the routine proceeds to step 109, where the variable valve timing mechanism so that the valve opening timing IO and the valve closing timing IC of the intake valve 7 change as shown in FIG. B is driven, and the machine is delayed by the change pattern of any one of (I), (II), and (III) in FIG. 11, that is, with respect to the operation of changing the valve opening timing IO and the valve closing timing IC of the intake valve 7. The variable compression ratio mechanism A is driven so that the compression ratio changes.

  On the other hand, when it is determined in step 108 that ΔL ≧ 0, that is, when the engine load has increased by the set value XL or more, the routine proceeds to step 110 and any one of (I), (II), and (III) in FIG. The variable compression ratio mechanism A is driven so that the mechanical compression ratio changes in accordance with the change pattern, and the valve opening timing IO of the intake valve 7 and the valve closing are delayed with respect to the operation of changing the mechanical compression ratio as shown in FIG. The variable valve timing mechanism B is driven so that the timing IC changes.

  On the other hand, when it is determined in step 107 that | ΔL | ≦ XL, that is, when the change amount ΔL of the engine load is small, the routine proceeds to step 111 where the opening timing IO of the intake valve 7 is changed as shown in FIG. The operation, the changing operation of the valve closing timing IC of the intake valve 7 and the changing operation of the mechanical compression ratio are started at the same time, and the variable compression ratio mechanism A and the variable valve timing mechanism B are driven so as to be completed almost simultaneously.

1 is an overall view of a spark ignition internal combustion engine. It is a disassembled perspective view of a variable compression ratio mechanism. 1 is a schematic side sectional view of an internal combustion engine. It is a figure which shows a variable valve timing mechanism. It is a figure which shows the lift amount of an intake valve and an exhaust valve. It is a figure for demonstrating a mechanical compression ratio, an actual compression ratio, and an expansion ratio. It is a figure which shows the relationship between theoretical thermal efficiency and an expansion ratio. It is a figure for demonstrating a normal cycle and a super-high expansion ratio cycle. It is a figure which shows changes, such as a mechanical compression ratio according to an engine load. It is a figure which shows the lift amount of an intake valve and an exhaust valve. It is a figure which shows the valve opening timing IO of intake valve, valve closing timing IC, and the change of mechanical compression ratio. It is a figure which shows the valve opening timing IO of intake valve, valve closing timing IC, and the change of mechanical compression ratio. It is a figure which shows the valve opening timing IO of intake valve, valve closing timing IC, and the change of mechanical compression ratio. It is a figure which shows the map etc. of the valve opening time IO of an intake valve. It is a flowchart for performing operation control.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Crankcase 2 Cylinder block 3 Cylinder head 4 Piston 5 Combustion chamber 7 Intake valve A Variable compression ratio mechanism B Variable valve timing mechanism

Claims (21)

  It has a variable compression ratio mechanism that can change the mechanical compression ratio and a variable valve timing mechanism that can individually control the opening timing and closing timing of the intake valve, and the intake valve closes as the engine load decreases. By moving the timing away from the intake bottom dead center, the amount of intake air corresponding to the required load is supplied into the combustion chamber, and the mechanical compression ratio is maximized so that the maximum expansion ratio can be obtained during low engine load operation. At the same time, at least during the period when the mechanical compression ratio is maximized, the valve opening timing of the intake valve is maintained at the target valve opening timing within the non-interference area that is almost the top dead center of the intake air and does not interfere with the piston. A spark ignition internal combustion engine.   The spark ignition internal combustion engine according to claim 1, wherein the maximum expansion ratio is 20 or more.   The spark ignition type internal combustion engine according to claim 1, wherein an actual compression ratio at the time of engine low load operation is substantially the same as that at the time of engine medium and high load operation.   4. The spark ignition type internal combustion engine according to claim 3, wherein the actual compression ratio is within a range of about ± 10% with respect to the actual compression ratio at the time of medium and high load operation regardless of the engine load at a low engine speed.   The spark ignition internal combustion engine according to claim 3, wherein the actual compression ratio is increased as the engine speed increases.   When the engine operating state shifts from medium to high speed medium to high load operation to low load operation where the mechanical compression ratio is maximized, mechanical compression is performed after the intake valve opening timing reaches the target valve opening timing within the non-interference area. The spark ignition internal combustion engine according to claim 1, wherein the change operation of the mechanical compression ratio is delayed with respect to the change operation of the valve opening timing of the intake valve so that the ratio becomes maximum.   When the engine operating state shifts from medium to high speed medium to high load operation to low load operation where the mechanical compression ratio is maximized, mechanical compression is performed after the intake valve opening timing reaches the target valve opening timing within the non-interference area. 6. The spark ignition internal combustion engine according to claim 5, wherein a ratio changing operation is started.   When the engine operating state shifts from medium to high speed medium to high load operation to low load operation at which the mechanical compression ratio is maximized, the intake valve opening timing is changed after the intake valve opening timing is changed. 6. The spark ignition internal combustion engine according to claim 5, wherein an operation of changing the mechanical compression ratio is started before the target valve opening timing in the non-interference area is reached.   6. The spark ignition type internal combustion engine according to claim 5, wherein when the engine operating state shifts from a medium-high-speed medium-high load operation to a low-load operation in which the mechanical compression ratio is maximized, the changing speed of the mechanical compression ratio is slowed down.   When the engine operating state transitions from low-load operation, where the mechanical compression ratio is maximized, to medium-high speed, medium-high load operation, the intake valve opens after the mechanical compression ratio change operation that lowers the mechanical compression ratio is started. The spark ignition type internal combustion engine according to claim 1, wherein a timing changing operation is started.   When the engine operating state shifts from low-load operation, where the mechanical compression ratio is maximized, to medium-high speed, medium-high load operation, the intake valve is opened after the mechanical compression ratio drops to the target mechanical compression ratio corresponding to the engine operating state. The spark ignition internal combustion engine according to claim 9, wherein the valve timing changing operation is started.   When the engine operating state shifts from low-load operation, where the mechanical compression ratio is maximized, to medium-high-speed, medium-high load operation, the intake valve of the intake valve is reduced while the mechanical compression ratio decreases to the target mechanical compression ratio according to the engine operating state. The spark ignition internal combustion engine according to claim 9, wherein the valve opening timing changing operation is started.   When the engine operating state transitions from low-load operation, where the mechanical compression ratio is maximized, to medium-high speed, medium-high load operation, the change speed of the mechanical compression ratio changes when the valve opening timing change operation is started. The spark ignition internal combustion engine according to claim 9, which is slowed down.   The closing timing of the intake valve is moved in a direction away from the intake bottom dead center until the limit closing timing at which the intake air amount supplied into the combustion chamber can be controlled as the engine load decreases. Spark ignition internal combustion engine.   In the region where the load of the intake valve is higher than the engine load when the closing timing of the intake valve reaches the above limit closing timing, the amount of intake air supplied into the combustion chamber does not depend on the throttle valve disposed in the engine intake passage. 14. The spark ignition internal combustion engine according to claim 13, which is controlled by changing a closing timing of the intake valve.   The spark ignition internal combustion engine according to claim 14, wherein the throttle valve is held in a fully open state in a region where the load is higher than the engine load when the closing timing of the intake valve reaches the limit closing timing.   In the region where the load is lower than the engine load when the intake valve closing timing reaches the above limit closing timing, the amount of intake air supplied into the combustion chamber is controlled by the throttle valve disposed in the engine intake passage. The spark ignition internal combustion engine according to claim 13.   14. The spark ignition type internal combustion engine according to claim 13, wherein the air-fuel ratio is increased as the load decreases in a region where the load is lower than the engine load when the closing timing of the intake valve reaches the limit closing timing.   The spark ignition according to claim 13, wherein the closing timing of the intake valve is maintained at the limit closing timing in a region where the load is lower than the engine load when the closing timing of the intake valve reaches the limit closing timing. Internal combustion engine.   2. The spark ignition internal combustion engine according to claim 1, wherein the mechanical compression ratio is increased to a limit mechanical compression ratio as the engine load decreases.   20. The spark ignition internal combustion engine according to claim 19, wherein the mechanical compression ratio is maintained at the limit mechanical compression ratio in a region where the load is lower than the engine load when the mechanical compression ratio reaches the limit mechanical compression ratio.

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