KR101032288B1 - Spark ignition type internal combustion engine - Google Patents

Abstract

The spark ignition type internal combustion engine includes a variable compression ratio mechanism that can change the mechanical compression ratio, an actual compression action start timing change mechanism that can change the start timing of the actual compression action, and an exhaust valve. In order to obtain the maximum expansion ratio in the engine low load operation, the mechanical compression ratio is maximized and the actual compression ratio is set so that knocking does not occur. The maximum expansion ratio is 20 or more. The closing timing of the exhaust valve at the engine low load operation is substantially the intake top dead center. Thereby, even if the internal combustion engine operates in a state of large expansion ratio, the temperature of the exhaust purification catalyst can be maintained at a relatively high temperature.

Description

Spark-ignition internal combustion engine {SPARK IGNITION TYPE INTERNAL COMBUSTION ENGINE}

The present invention relates to a spark ignition type 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 control the opening timing of the intake valve, and performs a supercharge action by the supercharger during the engine medium load operation and the engine high load operation. In medium load and high load operation, a spark ignition type internal combustion engine is known which increases the mechanical compression ratio and delays the opening of the intake valve as the engine load decreases while the actual compression ratio is kept constant (for example, JP-A-A). See Patent Publication No. 2004-218522).

In general, however, in an internal combustion engine, the greater the expansion ratio, the longer the period during which the reduction force acts on the piston during the expansion stroke, and thus the greater the expansion ratio, the better the thermal efficiency. Therefore, in order to improve the thermal efficiency at the time of engine operation, it is desirable to make the mechanical compression ratio as high as possible to increase the expansion ratio.

However, when the expansion ratio is increased in this way, most of the thermal energy generated in the combustion chamber is converted into kinetic energy, so that the temperature of the exhaust gas is lowered. In addition, with this, the pressure of the exhaust gas in the combustion chamber at the end of the expansion stroke is also lowered, so that the exhaust gas is less likely to be discharged from the combustion chamber. This tendency is particularly remarkable when the expansion ratio is 20 or more.

On the other hand, if the engine exhaust purification catalyst provided in the engine exhaust passage does not rise above a certain temperature, its excellent exhaust purification generally cannot be exerted. For this reason, in most internal combustion engines, the heat of the exhaust gas discharged from the engine main body is used to keep the exhaust purification catalyst at a high temperature.

However, as described above, when the expansion ratio is increased, the temperature of the exhaust gas is lowered, so that the temperature at which the exhaust purification catalyst is raised per unit flow rate is lowered. In addition, when the expansion ratio is increased, the exhaust gas becomes difficult to be discharged from the combustion chamber, so that the flow rate of the exhaust gas flowing into the exhaust purification catalyst becomes small. For this reason, when the internal combustion engine is operated with a large expansion ratio, it becomes difficult to maintain the temperature of the exhaust purification catalyst at a high temperature.

Accordingly, it is an object of the present invention to provide a spark ignition type internal combustion engine capable of maintaining the temperature of the exhaust purification catalyst at a relatively high temperature even when the internal combustion engine is operated with a large expansion ratio.

The present invention provides a spark ignition type internal combustion engine according to the claims as a means for realizing the above object.

In one aspect of the present invention, there is provided a variable compression ratio mechanism capable of changing the mechanical compression ratio, an intake variable valve timing mechanism capable of changing the closing timing of the intake valve, and an exhaust valve,
In the engine low load operation, the variable compression ratio mechanism controls the mechanical compression ratio to maximize the mechanical compression ratio to obtain the maximum expansion ratio, and the intake variable valve timing mechanism controls the closing timing of the intake valve to set the actual compression ratio to prevent knocking. The maximum expansion ratio is 20 or more, and a spark ignition type internal combustion engine is provided in which the closing timing of the exhaust valve at the engine low load operation is substantially an intake top dead center.

According to another aspect of the present invention, there is provided a variable compression ratio mechanism capable of changing the mechanical compression ratio, an intake variable valve timing mechanism capable of changing the closing timing of the intake valve, and an exhaust variable valve timing mechanism capable of changing the closing timing of the exhaust valve; In the engine low load operation, the variable compression ratio mechanism controls the mechanical compression ratio to maximize the mechanical compression ratio to obtain the maximum expansion ratio, and the intake variable valve timing mechanism controls the closing timing of the intake valve to set the actual compression ratio to prevent knocking. The maximum expansion ratio is 20 or more, and a spark ignition type internal combustion engine is provided in which an area where the closing timing of the exhaust valve can be set at the time of engine low load operation is further limited to the intake top dead center side at the time of engine high load operation.

In another aspect of the present invention, in the engine low load operation, the closing timing of the exhaust valve is substantially intake top dead center.

In another aspect of the present invention, the closing timing of the exhaust valve and the opening timing of the intake valve are controlled to minimize the period in which the opening of the intake valve and the opening of the exhaust valve overlap during engine low load operation.

In another aspect of the present invention, the closing timing of the exhaust valve and the opening timing of the intake valve are controlled such that the period in which the opening of the intake valve and the opening of the exhaust valve overlap at the time of engine low load operation is zero.

In another aspect of the invention, in the engine low load operation, the opening timing of the intake valve becomes substantially intake top dead center.

In another aspect of the present invention, the actual compression ratio at the time of engine low load operation becomes the actual compression ratio substantially the same as at the engine medium and high load operation.

In another aspect of the present invention, regardless of the engine load at engine low speed, the actual compression ratio is lowered to within 9-11.

In another aspect of the invention, the higher the engine speed, the higher the actual compression ratio.

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In another aspect of the present invention, the amount of intake air supplied into the combustion chamber is controlled by changing the closing timing of the intake valve.

In another aspect of the present invention, as the engine load is lowered, the closing timing of the intake valve is moved in a direction away from the intake bottom dead center until the limit closing timing for controlling the amount of intake air supplied to the combustion chamber.

In another aspect of the present invention, in the region of the load higher than the engine load when the closing timing of the intake valve reaches the limit closing timing, the amount of intake air supplied into the combustion chamber is not dependent on the throttle valve disposed in the engine intake passage. It is controlled by changing the closing timing of the intake valve.

In another aspect of the present invention, in the region of the load higher than the engine load when the opening timing of the intake valve reaches the limit opening timing, the throttle valve is kept fully open.

In another aspect of the present invention, in the region of the load lower than the engine load when the closing timing of the intake valve reaches the limit closing timing, the throttle valve disposed in the engine intake passage controls the amount of intake air supplied into the combustion chamber. Used.

In another aspect of the present invention, in the region of the load lower than the engine load when the closing timing of the intake valve reaches the limit closing timing, the air-fuel ratio becomes larger as the load is lowered.

In another aspect of the present invention, in the region of the load lower than the engine load when the closing timing of the intake valve reaches the limit closing timing, the closing timing of the intake valve is maintained at the limit closing timing.

In another aspect of the invention, the mechanical compression ratio is increased up to the limit mechanical compression ratio as the engine load is lowered.

In another aspect of the present invention, in the region of the load lower than the engine load when the mechanical compression ratio reaches the limit mechanical compression ratio, the mechanical compression ratio is maintained at the limit mechanical compression ratio.

According to the present invention, since as much exhaust gas as possible is discharged from the combustion chamber to the exhaust purification catalyst, the exhaust purification catalyst can be maintained at a relatively high temperature even when the internal combustion engine is operated with a large expansion ratio.

The invention will be more clearly understood from the following detailed description with reference to the accompanying drawings.

1 is an overall view of a spark ignition type internal combustion engine.

2 is an exploded perspective view of the variable compression ratio mechanism.

3A and 3B are side cross-sectional views of the internal combustion engine shown.

4 is a view showing a variable valve timing mechanism.

5A and 5B are diagrams showing lift amounts of an intake valve and an exhaust valve.

6A, 6B and 6C are diagrams for explaining the mechanical compression ratio, the actual compression ratio, and the expansion ratio.

7 is a diagram showing the relationship between theoretical thermal efficiency and expansion ratio.

8A and 8B are views for explaining the normal cycle and the ultrahigh expansion ratio cycle.

9 is a view showing changes in mechanical compression ratio and the like depending on engine load.

10A, 10B, and 10C are diagrams showing lift changes of the intake valves and the exhaust valves.

11 is a view showing an area in which the closing timing of the exhaust valve can be set according to the mechanical compression ratio.

12A and 12B are views showing lift changes of the intake valve and the exhaust valve.

13 is a flowchart for operational control.

14A, 14B, and 14C are diagrams showing a target actual compression ratio and the like.

15A and 15B are diagrams showing a map of closing timings of exhaust valves and the like.

<Explanation of symbols for the main parts of the drawings>

1: crank case 2: cylinder block

3: cylinder head 4: piston

5: combustion chamber 7: intake valve

9: exhaust valve A: variable compression ratio mechanism

B: intake variable valve timing mechanism C: exhaust variable valve timing mechanism

1 shows a sectional side view of a spark ignition type internal combustion engine.

Referring to Fig. 1, reference numeral 1 denotes a crankcase, 2 cylinder block, 3 cylinder head, 4 piston, 5 combustion chamber, 6 spark plug disposed at the top center of combustion chamber 5, 7 intake valve. , 8 represents an intake port, 9 represents an exhaust valve, and 10 represents an exhaust port. The intake port 8 is connected to the surge tank 12 via the intake pipe 11, and each intake pipe 11 is provided with a fuel injector 13 for injecting fuel toward the corresponding intake port 8. do. Each fuel injector 13 may be arranged in each combustion chamber 5 instead of being attached to each intake pipe 11.

The surge tank 12 is connected to the outlet of the compressor 15a of the exhaust turbocharger 15 via the intake duct 14, and the inlet of the compressor 15a is, for example, an intake air quantity detector 16 using a hot wire. Is connected to the air cleaner 17. In the intake duct 14 is provided a throttle valve 19 driven by an actuator 18.

On the other hand, the exhaust port 10 is connected to the inlet of the exhaust turbine 15b of the exhaust turbocharger 15 via the exhaust manifold 20, and the outlet of the exhaust turbine 15b is purged of exhaust gas through the exhaust pipe 21. It is connected to a catalytic converter 22 containing a catalyst. The air-fuel ratio sensor 23 is disposed in the exhaust pipe 21.

In addition, in the embodiment shown in FIG. 1, the variable compression ratio which can change the relative position of the crankcase 1 and the cylinder block 2 in the cylinder axial direction to the connection part of the crankcase 1 and the cylinder block 2 is shown. The mechanism A is provided so that the volume of the combustion chamber 5 is changed when the piston 4 is located at compression top dead center. In addition, in order to change the start timing of the actual compression action, the closing timing of the intake valve 7 can be controlled, and the intake variable valve timing mechanism B capable of individually controlling the opening timing of the intake valve 7 is provided. Is provided. Moreover, the exhaust variable valve timing mechanism C which can control the opening timing and closing timing of the exhaust valve 7 separately is provided.

The electronic control unit 30 is composed of a digital computer, which includes a read only memory (ROM) 32, a random access memory (RAM) 33, and a microprocessor (CPU) connected to each other by a bidirectional bus 31. ), An input port 35 and an output port 36 are provided. 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 through the corresponding AD converter 37. The accelerator pedal 40 is also connected to a load sensor 41 that generates an output voltage proportional to the amount of depression of the accelerator pedal 40. The output voltage of the load sensor 41 is input to the input port 35 via the corresponding AD converter 37. In addition, the input port 35 is connected to the crank angle sensor 42 which generates an output pulse every time the crankshaft rotates, for example, 30 degrees. On the other hand, the output port 36 is connected to the spark plug 6, the fuel injector 13, the throttle valve drive actuator 18, the variable compression ratio mechanism A and the intake variable valve timing mechanism B via the drive circuit 38. Is connected to.

2 is an exploded perspective view of the variable compression ratio mechanism A shown in FIG. 1, and FIGS. 3A and 3B are side cross-sectional views of the internal combustion engine shown. Referring to FIG. 2, a plurality of protrusions 50 are formed at the bottom of both side walls of the cylinder block 2 and spaced apart from each other at a specific distance. Each projection 50 is formed with a cam insertion hole 51 of circular cross section. On the other hand, the upper surface of the crankcase 1 is formed with a plurality of protrusions 52 which are separated from each other by a specific distance and which fit between the corresponding protrusions 50. The cam insertion hole 53 of circular cross section is also formed in these protrusion part 52. As shown in FIG.

As shown in FIG. 2, a pair of camshafts 54, 55 are provided. Each cam shaft 54, 55 is fixed with a circular cam 56 rotatably inserted into the cam insertion hole 51 at every position. This circular cam 56 is coaxial with the rotation axis of the cam axes 54, 55. On the other hand, between the circular cams 56, eccentric shafts 57 eccentrically arranged with respect to the rotational axes of the cam shafts 54, 55 extend as shown by hatching in FIGS. 3A and 3B. Another circular cam 58 is attached to each eccentric shaft 57 so as to be eccentrically rotatable. As shown in FIG. 2, this circular cam 58 is disposed between the circular cams 56. This circular cam 58 is rotatably inserted into the corresponding cam insertion hole 53.

When the circular cams 56 fixed to the cam shafts 54 and 55 are rotated in opposite directions from each other as shown by the solid arrows in Fig. 3A from the state as shown in Fig. 3A, the eccentric shaft 57 is the bottom center. The circular cam 58 rotates in the opposite direction from the circular cam 56 as indicated by the broken arrow in FIG. 3A in the cam insertion hole 53. As shown in FIG. 3B, when the eccentric shaft 57 moves toward the bottom center, the center of the circular cam 58 moves below the eccentric shaft 57.

As can be seen by comparing FIGS. 3A and 3B, the relative position of the crankcase 1 and the cylinder block 2 is determined by the distance between the center of the circular cam 56 and the center of the circular cam 58. The cylinder block 2 is far from the crankcase 1 as the distance between the center of the circular cam 56 and the center of the circular cam 58 becomes large. When the cylinder block 2 is moved away from the crankcase 1, the volume of the combustion chamber 5 increases when the piston 4 is located at compression top dead center. Therefore, by rotating the cam shafts 54 and 55, the volume of the combustion chamber 5 can be changed when the piston 4 is located at compression top dead center.

As shown in Fig. 2, in order to rotate the cam shafts 54, 55 in the opposite direction, the shaft of the drive motor 59 is provided with a pair of worm gears 61, 62 having opposite spiral directions. Gears 63 and 64 meshing with these worm gears 61 and 62 are fixed to the ends of the cam shafts 54 and 55. In this embodiment, the drive motor 59 can be driven to change the volume of the combustion chamber 5 to a wide range when the piston 4 is located at compression top dead center. The variable compression ratio mechanism A shown in FIGS. 1-3 shows an example. Any type of variable compression ratio mechanism can be used.

4 shows the intake variable valve timing mechanism B attached to the camshaft 70 for driving the intake valve 7 of FIG. Referring to FIG. 4, the intake variable valve timing mechanism B is mounted on one end of the cam shaft 70 and cam phase change portion B1 for changing the phase of the cam of the cam shaft 70, and the cam shaft 70. ) And a cam operating angle which is disposed between the valve lifter 24 of the intake valve 7 and changes the operating angle (operating angle) of the cam of the cam shaft 70 to a different operating angle and transmits it to the intake valve 7. It is comprised by the change part B2. 4 is a side sectional view and a plan view of the cam operating angle changing unit B2.

First, when the cam phase change part B1 of the intake variable valve timing mechanism B is demonstrated, this cam phase change part B1 is a timing which can make it rotate in a direction of an arrow via a timing belt by the crank shaft of an engine. Pulley 71, cylindrical housing 72 that rotates with timing pulley 71, axis 73 that rotates with camshaft 70 and rotates relative to cylindrical housing 72, of cylindrical housing 72. Between the partition 74 and the plurality of partitions 74 extending from the inner circumferential surface to the outer circumferential surface of the shaft 73 vanes 75 extending from the outer circumferential surface of the shaft 73 to the inner circumferential surface of the cylindrical housing 72 On both sides of the vanes 75, an advance hydraulic chamber 76 and a crust hydraulic chamber 77 are formed.

The hydraulic oil supply to the hydraulic chambers 76, 77 is controlled by the 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 supply port 82 of hydraulic oil discharged from the hydraulic pump 81, and a pair of drain ports 83, 84) and a spool valve 85 for controlling the connection and disconnection between the ports 79, 80, 82, 83, 84 is provided.

In order to advance the phase of the cam of the cam shaft 70, the spool valve 85 moves downward in FIG. 4, so that the hydraulic oil supplied from the supply port 82 passes through the hydraulic port 79 for the advance hydraulic chamber. It is supplied to the 76, and the hydraulic oil in the tectonic hydraulic chamber 77 is discharged from the drain port 84. At this time, the shaft 73 is made to rotate relative to the cylindrical housing 72 in the arrow X direction.

In contrast, in order to perceive the phase of the cam of the camshaft 70, the spool valve 85 moves upward in FIG. 4, so that the hydraulic oil supplied from the supply port 82 is perceived through the hydraulic port 80. The hydraulic oil in the hydraulic hydraulic chamber 76 is supplied from the drain port 83 to the hydraulic oil pressure chamber 77. At this time, the shaft 73 is made to rotate relative to the cylindrical housing 72 in the direction opposite to the arrow X.

When the shaft 73 rotates relative to the cylindrical housing 72, if the spool valve 85 returns to the neutral position shown in FIG. 4, the relative rotational motion of the shaft 73 ends and the shaft 73 is Is maintained at its relative rotational position. Accordingly, the cam phase changing portion B1 can be used to advance or perceive the phase of the cam of the cam shaft 70 by exactly the desired amount. That is, the cam phase change unit B1 can freely advance or retard the opening timing of the intake valve 7.

Next, the cam operating angle changing unit B2 of the intake variable valve timing mechanism B will be described. The cam operating angle changing unit B2 is disposed in parallel with the cam shaft 70 and provided to the actuator 91. Slidably fitted to the control rod 90, which is moved axially by means of the cam rod 70, the cam 92 of the camshaft 70, and is formed on the control rod 90 and slidably fitted to the axially extending spline 93. Pivoting, which meshes with a valve lifter 24 for driving the intermediate cam 94 and the intake valve 7, and which is formed in the control rod 90 and slidably fits into a spirally extending spline 95 ( pivoting cam 96 is provided. The cam 97 is formed on the pivoting cam 96.

When the cam shaft 70 rotates, the cam 92 causes the intermediate cam 94 to always pivot exactly a constant angle. At this time, the pivoting cam 96 also turns only a certain angle. On the other hand, the intermediate cam 94 and the pivoting cam 96 are supported so that they cannot move in the axial direction of the control rod 90, and thus, when the control rod 90 is moved in the axial direction by the actuator 91, The pivoting cam 96 is made to rotate relative to the intermediate cam 94.

When the cam 92 of the cam shaft 70 starts to engage the intermediate cam 94 by the relative rotational position relationship between the intermediate cam 94 and the pivoting cam 96, the cam of the pivoting cam 96 ( 97) If this starts to engage with the valve lifter 24, the opening period and the lift amount of the intake valve 7 become maximum, as indicated by a in FIG. 5B. In contrast, when the actuator 91 is used to cause the pivoting cam 96 to rotate relative to the intermediate cam 94 in the direction of the arrow Y in FIG. 4, the cam 92 of the cam shaft 70 is the intermediate cam 94. ), And the cam 97 of the pivoting cam 96 is engaged with the valve lifter 24 for a while. In this case, as shown by b in FIG. 5B, the opening period and the lift amount of the intake valve 7 become smaller than a.

When the pivoting cam 96 is made to rotate relative to the intermediate cam 94 in the direction of arrow Y in FIG. 4, as shown by c in FIG. 5B, the opening period and the lift amount of the intake valve 7 become smaller. That is, by using the actuator 91 to change the relative rotation positions of the intermediate cam 94 and the pivoting cam 96, the opening period of the intake valve 7 can be freely changed. In this case, however, the lift amount of the intake valve 7 becomes smaller as the opening period of the intake valve 7 becomes shorter.

In this way, the cam phase change portion B1 can be used to freely change the opening timing of the intake valve 7, and the cam operating angle change portion B2 can be used to freely change the opening period of the intake valve 7. Therefore, both the cam phase change unit B1 and the cam operating angle change unit B2, that is, the intake variable valve timing mechanism B, open the intake valve 7 and the open period, that is, the intake valve 7. It can be used to freely change the timing of opening and closing.

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

In addition, the exhaust variable valve timing mechanism C also basically has the same configuration as the intake variable valve timing mechanism B, and the opening timing and opening period of the exhaust valve 9, that is, the opening timing and closing timing of the exhaust valve 9. You can change it freely.

Next, the meanings of the terms used in the present application will be described with reference to FIGS. 6A to 6C. 6A, 6B and 6C are shown to illustrate an engine having a combustion chamber volume of 50 ml and a piston stroke volume of 500 ml. 6A, 6B and 6C, the combustion chamber volume represents the volume of the combustion chamber when the piston is located at compression top dead center.

It is a figure explaining a mechanical compression ratio. The mechanical compression ratio is a value determined mechanically from the stroke volume of the piston during the compression stroke and the combustion chamber volume. This mechanical compression ratio is expressed as (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.

6B is a diagram for explaining an 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. This actual compression ratio is expressed as (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 action is not performed while the intake valve is open. The actual compression action starts after the intake valve is closed. Therefore, the actual compression ratio is expressed as follows using the actual stroke volume. In the example shown in Fig. 6B, the actual compression ratio is (50 ml + 450 ml) / 50 ml = 10.

It is a figure explaining an expansion ratio. The expansion ratio is a value determined from the piston stroke volume and the combustion chamber volume during the expansion stroke. This expansion ratio is expressed as (combustion chamber volume + administrative 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. 7, 8A and 8B. FIG. 7 is a diagram showing a relationship between theoretical thermal efficiency and expansion ratio, and FIGS. 8A and 8B are diagrams showing a comparison between a normal cycle and an ultrahigh expansion ratio cycle selectively used according to the load in the present invention.

FIG. 8A shows a normal cycle in which the intake valve is closed near the bottom dead center and the compression action by the piston is started from substantially near the bottom dead center of compression. Similarly to this example shown in Fig. 8A, the combustion chamber volume is 50 ml and the stroke volume of the piston is 500 ml in the same manner as the example 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 also about 11, and the expansion ratio is also (50 ml + 500 ml) / 50 ml = 11 Becomes In other words, usually in an internal combustion engine, the mechanical compression ratio and the actual compression ratio and expansion ratio become substantially the same.

The solid line in Fig. 7 shows the change in theoretical thermal efficiency when the actual compression ratio and the expansion ratio are substantially equal, that is, in the normal cycle. In this case, it is found that the theoretical thermal efficiency also increases as the expansion ratio increases, that is, as the actual compression ratio increases. Therefore, in the normal cycle, in order to increase the theoretical thermal efficiency, the actual compression ratio should be higher. However, by limiting the occurrence of knocking during engine high load operation, the actual compression ratio can only be increased up to about 12, and therefore, in a normal cycle, the theoretical thermal efficiency cannot be sufficiently high.

On the other hand, under these circumstances, the inventor strictly distinguished the mechanical compression ratio from the actual compression ratio and studied the theoretical thermal efficiency, and found that the expansion ratio is dominant in the theoretical thermal efficiency, and the theoretical thermal efficiency is hardly affected by the actual compression ratio. . In other words, if the actual compression ratio is increased, the explosive force is high but a large energy is required to compress, so even if the actual compression ratio is increased, the theoretical thermal efficiency will hardly be increased.

On the contrary, when the expansion ratio is increased, the period during which the reduction force is applied to the piston during the expansion stroke is longer, and the period during which the piston is giving rotational force to the crankshaft is also longer. Therefore, the greater the expansion ratio, the higher the theoretical thermal efficiency. The broken line in Fig. 7 shows the theoretical thermal efficiency when the actual compression ratio is fixed at 10 and the expansion ratio is increased in that state. In this manner, the theoretical thermal efficiency increase when the expansion ratio is increased while the actual compression ratio is kept at a low value and the theoretical thermal efficiency increase when the actual compression ratio increases with the expansion ratio as shown by the solid line in FIG. Find that there is no big difference.

In this way, if the actual compression ratio is kept at a low value, knocking will not occur, and therefore, if the expansion ratio is increased while keeping the actual compression ratio at a low value, the occurrence of knocking can be prevented and the theoretical thermal efficiency can be greatly increased. 8B shows an example in which the actual compression ratio is kept low and the expansion ratio is increased by using the variable compression ratio mechanism A and the variable valve timing mechanism B. FIG.

Referring to FIG. 8B, in this example, the variable compression ratio mechanism A is used to reduce the combustion chamber volume from 50 ml to 20 ml. On the other hand, the intake variable valve timing mechanism B is used to delay the closing timing of the intake valve until the actual piston stroke volume becomes 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, as described above, the actual compression ratio is about 11 and the expansion ratio is also 11. Compared with this case, in the case shown in Fig. 8B, it is found that only the expansion ratio is increased to 26. This will be referred to as the "ultra high expansion ratio cycle".

As described above, generally speaking, in an internal combustion engine, the lower the engine load, the worse the thermal efficiency, and therefore, in order to improve the thermal efficiency when driving the vehicle, i.e., to improve fuel efficiency, the thermal efficiency is improved during engine low load operation. It is necessary to let. On the other hand, in the ultra high expansion ratio cycle shown in Fig. 8B, the actual piston stroke volume during the compression stroke becomes small, so that the amount of intake air that can be sucked into the combustion chamber 5 becomes small, and therefore this ultra high expansion ratio cycle is achieved when the engine load is relatively low. Only can be used. Therefore, in the present invention, the ultrahigh expansion ratio cycle shown in FIG. 8B is set at the time of engine low load operation, and the normal cycle shown in FIG. 8A is set at the time of engine high load operation. This is a basic feature of the present invention.

9 shows the operation control as a whole in the normal operation of the engine low speed. In the following, operation control as a whole will be described with reference to FIG.

9 shows the change 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 this embodiment, in order for the three-way catalyst in the catalytic converter 22 to simultaneously reduce unburned HC, CO and NOx in the exhaust gas, the average air-fuel ratio in the combustion chamber 5 is usually based on the output signal of the air-fuel ratio sensor 23. Feedback control with stoichiometric air-fuel ratio.

Now, as described above, in the engine high load operation, the normal cycle shown in Fig. 8A is executed. Therefore, as shown in Fig. 9, the mechanical compression ratio is lowered at this time, the expansion ratio is lowered, and the closing timing of the intake valve 7 is advanced as shown by the solid line in Fig. 9. At this time, the amount of intake air increases. At this time, the opening degree of the throttle valve 17 is maintained at full open or substantially full open, so that the pumping loss is zero.

On the other hand, as shown in Fig. 9, as the engine load decreases, the mechanical compression ratio increases, and thus the expansion ratio also increases. At this time, as indicated by the solid line in Fig. 9, the lower the engine load, the later the closing timing of the intake valve 7 is maintained, so that the actual compression ratio is kept substantially constant. Even at this time, the throttle valve 17 is kept in a fully open or substantially full open state, so that the amount of intake air supplied to the combustion chamber 5 is not changed by the throttle valve 17 so as to change the closing timing of the intake valve 7. Is controlled by At this time, the pumping loss is zero.

In this way, when the engine load is lowered from the engine high load operating state, the mechanical compression ratio is increased as the amount of intake air under a substantially constant actual compression ratio is reduced. That is, the volume of the combustion chamber 5 when the piston 4 reaches compression top dead center decreases in proportion to the decrease in the intake air amount. Therefore, the volume of the combustion chamber 5 when the piston 4 reaches compression top dead center changes in proportion to the amount of intake air. At this time, the air-fuel ratio of the combustion chamber 5 becomes a stoichiometric air-fuel ratio, and therefore the volume of the combustion chamber 5 when the piston 4 reaches compression top dead center changes in proportion to the fuel amount.

The lower the engine load, the higher the mechanical compression ratio. When the mechanical compression ratio reaches the limit mechanical compression ratio corresponding to the structural limit of the combustion chamber 5, the mechanical compression ratio is kept at the limit mechanical compression ratio in the load region lower than the engine load L1 when the mechanical compression ratio reaches the limit mechanical compression ratio. do. Therefore, in the engine low load operation, the mechanical compression ratio is maximum, and the expansion ratio is also maximum. In other words, in the present invention, in order to obtain the maximum expansion ratio at the time of engine low load operation, the mechanical compression ratio is maximum. In addition, the actual compression ratio at this time is maintained at the actual compression ratio substantially the same as in the engine heavy load operation.

On the other hand, as indicated by the solid line in FIG. 9, the closing timing of the intake valve 7 is further delayed until the limit closing timing capable of controlling the amount of intake air supplied to the combustion chamber 5 as the engine load decreases. In the load region lower than the engine load L2 when the closing timing of the intake valve 7 reaches the limit closing timing, the closing timing of the intake valve 7 is maintained at the limit closing timing. If the closing timing of the intake valve 7 is maintained at the limit closing timing, the intake air amount will no longer be controlled by changing the closing timing of the intake valve 7. Therefore, the intake air amount needs to be controlled by any other method.

In the embodiment shown in FIG. 9, at this time, that is, in the load region lower than the engine load L2 when the closing timing of the intake valve 7 reaches the limit closing timing, the throttle valve 17 is operated by the combustion chamber 5. Used to control the amount of intake air supplied to the system. However, if the throttle valve 17 is used to control the intake air amount, the pumping loss increases as shown in FIG.

In order to prevent the occurrence of this pumping loss, in the load region lower than the engine load L2 when the closing timing of the intake valve 7 reaches the limit closing timing, the throttle valve 17 is fully opened or substantially opened. With the engine fully open, the air-fuel ratio can be greater as the engine load decreases. At this time, the fuel injector 13 is preferably arranged in the combustion chamber 5 to perform stratified combustion.

As shown in Fig. 9, at the engine low speed, the actual compression ratio remains substantially constant regardless of the engine load. The actual compression ratio at this time is in the range of ± 10 percent, and preferably in the range of ± 5 percent to the actual compression ratio in the engine heavy load operation. In this embodiment, the actual compression ratio at the time of engine low speed is about 10 +/- 1, that is, 9 to 11. However, if the engine speed is high, confusion occurs in the air-fuel mixture in the combustion chamber 5, and knocking is less likely to occur, so in the embodiment according to the present invention, if the engine speed is high, the actual compression ratio is also high.

On the other hand, as described above, the expansion ratio becomes 26 in the ultrahigh expansion ratio cycle shown in Fig. 8B. The higher the expansion ratio, the better. However, 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 addition, in the example shown in Fig. 9, the mechanical compression ratio is continuously changed in accordance with the engine load. However, the mechanical compression ratio may also change step by step depending on the engine load.

On the other hand, as indicated by the broken line in FIG. 9, as the engine load decreases, the amount of intake air can be controlled similarly by the throttle valve by advancing the closing timing of the intake valve 7. Therefore, if both the case shown by the solid line and the case shown by the broken line are represented, in FIG. 9, in the embodiment which concerns on this invention, the closing timing of the intake valve 7 is supplied to a combustion chamber as engine load becomes low. It moves in a direction away from the compression bottom dead center (BDC) until the limit closing time (L ₂ ) to control the intake air volume.

Next, the closing timing of the exhaust valve 9 will be described focusing on the low load operation in which the ultra-high expansion ratio cycle shown in FIG. 8B is executed.

In general, in the low load operation in which the ultra-high expansion ratio cycle is executed, the amount of heat generated by combustion of the air-fuel mixture in the combustion chamber 5 itself is low, and therefore the temperature of the exhaust gas discharged from the combustion chamber 5 tends to be low. In addition, in an internal combustion engine, the larger the expansion ratio, the longer the period during which the reduction force acts on the piston during the expansion stroke, so that most of the thermal energy generated by the combustion of the air-fuel mixture in the combustion chamber is converted into the kinetic energy of the piston. Is converted. In connection with this, the temperature of the combustion gas in a combustion chamber becomes low at the end of an expansion stroke. For this reason, when the ultrahigh expansion ratio cycle shown in FIG. 8B is executed, the temperature of the exhaust gas discharged from the combustion chamber 5 to the exhaust manifold 20 at the time of the exhaust stroke becomes very low. This tendency is particularly marked when the expansion ratio is 20 or more. Between the execution of an ultrahigh expansion ratio cycle having an expansion ratio of 20 or more and a normal cycle having an expansion ratio of about 12, the temperature of the exhaust gas discharged from the combustion chamber 5 differs by about 100 ° C.

On the other hand, in most internal combustion engines, harmful components (e.g., HC, CO, NO _X, etc.) contained in the exhaust gas are provided by providing an oxidation catalyst, a three-way catalyst, a NO _X storage reduction catalyst or another exhaust purification catalyst in the engine exhaust passage. Removed. Such exhaust purification catalysts cannot effectively purify the harmful components in the exhaust gas unless their temperature is higher than the active temperature. Here, in most internal combustion engines, the temperature of the exhaust gas is considerably higher than the active temperature, so that the exhaust gas enters the exhaust purification catalyst to maintain the temperature of the exhaust purification catalyst above the active temperature.

However, if the ultra high expansion ratio cycle shown in FIG. 8B is executed, the temperature of the exhaust gas discharged from the combustion chamber 5 to the exhaust manifold 20 will be only slightly higher than the active temperature, and the exhaust gas is introduced into the exhaust purification catalyst. Even so, it becomes difficult to maintain the temperature of the exhaust purification catalyst above the active temperature. Therefore, when the ultra high expansion ratio cycle is executed, in order to keep the temperature of the exhaust purification catalyst above the active temperature, it is necessary to introduce as much exhaust gas as possible into the exhaust purification catalyst.

Here, with reference to FIGS. 10A-10C, the relationship between the closing timing of the exhaust valve 9, and the flow volume of the exhaust gas discharged | emitted from the combustion chamber 5 to the exhaust manifold 20 is considered. FIG. 10A shows that when the exhaust valve 9 is substantially closed at intake top dead center, FIG. 10B shows that the exhaust valve 9 is closed before intake top dead center, and FIG. 10C shows that the exhaust valve 9 is after intake top dead center. The lift change of the exhaust valve 9 and the intake valve 7 when it is closed is shown.

As shown in FIG. 10B, when the exhaust valve 9 is closed before the intake top dead center, the volume of the combustion chamber 5 at the time of closing the exhaust valve 9 is the volume of the combustion chamber when the piston is located at the intake top dead center. Greater than (combustion chamber volume). After the exhaust valve 9 is closed, exhaust gas corresponding to the volume of the combustion chamber 5 at the time of closing is left in the combustion chamber 5. For this reason, even after the exhaust valve 9 is closed, a relatively large amount of exhaust gas remains in the combustion chamber 5. Therefore, it is impossible to sufficiently discharge the exhaust gas in the combustion chamber 5 to the exhaust manifold 20, and the flow rate of the exhaust gas flowing into the exhaust purification catalyst becomes small.

On the other hand, as shown in Fig. 10C, when the exhaust valve 9 is closed after the intake top dead center, the exhaust valve 9 is also opened at the intake top dead center, and thus when the piston 4 reaches the intake top dead center, Most of the exhaust gas in the combustion chamber 5 flows out in the exhaust port 10. However, when the exhaust valve 9 is opened even after the intake top dead center, part of the exhaust gas once leaked into the exhaust port 10 with the lowering of the piston 4 will flow back into the combustion chamber 5.

In particular, when the ultra high expansion ratio cycle is executed, in the expansion stroke, the combustion gas in the combustion chamber 5 considerably expands, and thus the pressure of the combustion gas at the end of the expansion stroke becomes relatively low. For this reason, the intensity | strength of the exhaust gas which flows out from the combustion chamber 5 to the exhaust port 10 in an exhaust stroke will become weak. Therefore, if the piston 4 descends after reaching intake top dead center, a part of the exhaust gas flowing out of the exhaust port 10 will easily flow into the combustion chamber 5 again.

In this way, when the exhaust valve 9 is closed after the intake top dead center, the exhaust gas once flowed into the exhaust port 10 returns to the combustion chamber 5 again, thus sufficiently exhausting the exhaust gas in the combustion chamber 5. It will not be possible to exhaust the exhaust manifold 20 and the flow rate of the exhaust gas flowing into the exhaust purification catalyst will be small.

Therefore, in the present embodiment, when the ultra high expansion ratio cycle shown in FIG. 8B is executed, that is, when the mechanical compression ratio is high, in order to prevent the closing timing of the exhaust valve 9 from reaching the intake top dead center too soon or too late, The region in which the closing timing of the exhaust valve 9 can be set is limited to the intake top dead center side.

FIG. 11 is a diagram showing an area in which the closing timing of the exhaust valve 9 can be set according to the mechanical compression ratio.

As shown in Fig. 11, the area in which the exhaust valve 9 can be set becomes an area between the maximum advance amount and the maximum perceptual amount that can be set. As can be seen from the figure, the advance amount by which the closing timing of the exhaust valve 9 can be set becomes smaller (delayed) as the mechanical compression ratio becomes higher, and conversely, the maximum perception by which the closing timing of the exhaust valve 9 can be set. The amount decreases (which leads to) the higher the mechanical compression ratio. For this reason, the area in which the closing timing of the exhaust valve 9 can be set becomes smaller as the mechanical compression ratio is higher, that is, more restrictive as the mechanical compression ratio is higher. For example, as shown in Fig. 11, when the mechanical compression ratio is low, the region in which the closing timing of the exhaust valve 9 can be set is ΔTOC1, and when the mechanical compression ratio is high, the closing timing of the exhaust valve 9 can be set. The area | region which exists is (DELTA) TOC2 ((DELTA) TOC2 <(DELTA) TOC1).

Alternatively, when the ultra high expansion ratio cycle shown in FIG. 8B is executed, that is, when the mechanical compression ratio is high, in order to ensure that the closing timing of the exhaust valve 9 is too early or too late than the intake top dead center, it is shown in FIG. 10A. As can be seen, the closing timing of the exhaust valve 9 can be substantially the intake top dead center.

In this manner, when the mechanical compression ratio is high, the combustion chamber is restricted by limiting the region where the closing timing of the exhaust valve 9 can be set to the intake top dead center side or by making the closing timing of the exhaust valve 9 substantially intake top dead center. The exhaust gas in (5) is sufficiently discharged to the exhaust manifold 20 to increase the flow rate of the exhaust gas flowing into the exhaust purification catalyst.

That is, the exhaust valve 9 is closed at the intake top dead center, and thus, as shown in FIG. 10B, the exhaust valve 9 is closed at the time of closing the exhaust valve 9 as compared with closing the exhaust valve 9 earlier than the intake top dead center. The volume of the combustion chamber 5 is small and therefore the amount of exhaust gas remaining in the combustion chamber 5 after the closing of the exhaust valve 9 can be reduced. In addition, the exhaust valve 9 is closed near the intake top dead center, and as shown in FIG. 10C, compared with closing the exhaust valve 9 later than the intake top dead center, the exhaust gas discharged into the exhaust port 10. The amount of exhaust gas flowing back into the heavy combustion chamber 5 can be reduced. Therefore, when the exhaust valve 9 is closed near the intake top dead center as shown in FIG. 10A, as shown in FIGS. 10B and 10C, the exhaust valve 9 is closed away from the intake top dead center as compared with the case where the exhaust valve 9 is closed. The exhaust gas in the combustion chamber 5 can be sufficiently discharged to the exhaust manifold 20 to increase the flow rate of the exhaust gas flowing into the exhaust purification catalyst. As a result, the exhaust purification catalyst can be kept above the active temperature even in the low load operation in which the ultra-high expansion ratio cycle is executed.

Said "substantially intake top dead center" is within about 10 degrees before and after intake top dead center, preferably within about 5 degrees before and after intake top dead center.

In addition, when the mechanical compression ratio is increased, the combustion chamber volume at the intake top dead center becomes small, so that the exhaust valve 9 interferes with the piston 4 by the closing timing of the exhaust valve 9.

10A to 10C show a piston interference line indicating a limit at which the exhaust valve 9 or the intake valve 7 interferes with the piston 4. When the lift curve of the exhaust valve 9 intersects the piston interference line, the exhaust valve 9 interferes with the piston 4. Here, in FIG. 10C, the lift curve of the exhaust valve 9 intersects the piston interference line. This means that when the exhaust valve 9 is closed later than the intake top dead center, the exhaust valve 9 and the piston 4 interfere with each other depending on the degree of the delay.

In contrast, according to the present embodiment, when the mechanical compression ratio is high, the region where the closing timing of the exhaust valve 9 can be set is limited to the intake top dead center side, and in particular, the closing timing of the exhaust valve 9 can be set. The maximum amount of crust becomes smaller. For this reason, as shown in FIG. 10A, even if the mechanical compression ratio is high, the exhaust valve 9 can be prevented from interfering with the piston 4.

However, when there is a valve overlap where the opening period of the intake valve 7 and the opening period of the exhaust valve 9 overlap, the amount of exhaust gas discharged from the combustion chamber 5 inside to the exhaust manifold 20 also varies during that period. . 12A and 12B, the overlapping period in which the opening period of the intake valve 7 and the opening period of the exhaust valve 9 overlap and the exhaust gas discharged from the combustion chamber 5 to the exhaust manifold 20 are described below. Think about positive relationships. FIG. 12A is a diagram showing a lift change of the exhaust valve 9 and the intake valve 7 when the overlap period is zero, and FIG. 12B.

In general, when the intake valve 7 and the exhaust valve 9 open at the same time, a part of the exhaust gas in the combustion chamber 5 and a part of the exhaust gas once flowed out from the combustion chamber 5 to the exhaust port 10 are intake ports. It will sometimes flow in (8). In this way, if a part of the exhaust gas flows into the intake port 8, the exhaust gas discharged from the combustion chamber 5 to the exhaust manifold 20 will be reduced by that amount.

Therefore, when the overlap period is large, as shown in FIG. 12B, the exhaust gas will often enter the intake port 8 in large quantities. Therefore, the exhaust gas discharged from the combustion chamber 5 to the exhaust manifold 20 will often be less. For this reason, in this case, the flow volume of the exhaust gas which flows into an exhaust purification catalyst will become small.

Therefore, in this embodiment, when the ultra-high expansion ratio cycle shown in FIG. 8B is executed, that is, when the mechanical compression ratio is high, the closing timing of the exhaust valve 9 and the opening timing of the intake valve 7 are overlapped as shown in FIG. 12A. This is controlled to be the minimum of the ranges that can be set. Thus, for example, in an internal combustion engine whose settable overlap period is 10 ° to 60 °, when the mechanical compression ratio is high, the overlap period becomes 10 °, and in an internal combustion engine whose settable overlap period is 0 ° to 50 °, the mechanical compression ratio When is high, the overlap period becomes 0 °.

In this way, by minimizing the overlap period when the mechanical compression ratio is high, the exhaust gas flowing into the intake port 8 is less, and therefore the exhaust gas discharged from the combustion chamber 5 to the exhaust manifold 20 is thus increased. The flow rate of the exhaust gas flowing into the exhaust purification catalyst increases.

The overlap period when the mechanical compression ratio is high is not necessarily minimum if it is shorter than the overlap period when the mechanical compression ratio is low. Therefore, for example, the overlap period when the mechanical compression ratio is high may be 10 ° or less in the settable range even if it is minimum.

As described above, when the mechanical compression ratio is increased, the combustion chamber volume at the intake top dead center becomes small. Therefore, the intake valve 7 interferes with the piston 4 by the opening timing of the intake valve 7.

12A and 12B are diagrams showing a piston interference line indicating a limit at which the exhaust valve 9 or the intake valve 7 interferes with the piston 4. When the lift curve of the intake valve 7 crosses the piston interference line, the intake valve 7 interferes with the piston 4. Here, in FIG. 12B, the lift curve of the intake valve 7 intersects the piston interference line. This means that when the overlap period is enlarged, the intake valve 7 and the piston 4 interfere with each other. That is, in this embodiment, as described above, the closing timing of the exhaust valve 9 becomes substantially intake top dead center. The large overlap period means that the opening timing of the intake valve 7 is greatly advanced. If the opening timing of the intake valve 7 is greatly advanced, the intake valve 7 and the piston 4 will interfere with each other.

In contrast, according to the present embodiment, when the mechanical compression ratio is high, the overlap period is minimized, and therefore the opening timing of the intake valve 7 is substantially at or after intake top dead center. For this reason, as shown in FIG. 12A, even if the mechanical compression ratio becomes high, the intake valve 7 is prevented from interfering with the piston.

Fig. 13 is a diagram showing a control routine for operation control of the spark ignition type internal combustion engine of the present embodiment. Referring to Fig. 13, initially in step 101, the engine load L and the engine speed Ne are obtained. Next, in step 102, the map shown in Fig. 14A is used to calculate the target actual compression ratio. As shown in Fig. 14A, this target actual compression ratio increases as the engine speed Ne increases. Next, in step 103, the map shown in Fig. 14B is used to calculate the mechanical compression ratio CR. In other words, the mechanical compression ratio CR necessary for making the actual compression ratio the target actual compression ratio is stored in the ROM 32 in advance in the form of a map shown in Fig. 14B as a function of the engine load L and the engine speed Ne. This map is used to calculate the mechanical compression ratio CR.

In addition, the closing timing IC of the intake valve 7 necessary to supply the required intake air amount to the combustion chamber 5 is ROM in advance in the form of a map shown in FIG. 14C as a function of the engine load L and the engine speed Ne. Stored at 32. In step 104, this map is used to calculate the closing timing (IC) of the intake valve 7.

Next, in step 105, it is determined whether the engine load L is smaller than the predetermined value L ₃ . Here, the predetermined value of temperature (L ₃₎ is, for example, when the smaller the engine load, and the engine load and the equivalent value, lowering of the temperature of the exhaust gas is the temperature of the exhaust purifying catalyst to be reduced to below the activation temperature May involve degradation. When it is determined in step 105 that the engine load L is smaller than the predetermined value L ₃ , the flow proceeds to step 106. In step 106, the closing timing EC of the exhaust valve 9 becomes substantially intake top dead center. Then, in step 107, the overlap period DELTA OL is minimized and the routine proceeds to step 110.

On the other hand, when the engine load is determined to be equal to or greater than the predetermined value L ₃ in step 105, the routine proceeds to step 108. In step 108, the map shown in FIG. 15A is used to calculate the closing timing EC of the exhaust valve 9, and then in step 109, the map shown in FIG. 15B is used to calculate the overlap period ΔOL. That is, the closing timing EC and the overlapping period ΔOL of the exhaust valve 9 are previously stored in the ROM 32 in the form of a map shown in FIGS. 15A and 15B as a function of the engine load L and the engine speed Ne. Stored. This map is used to calculate the closing timing EC and the overlapping period ΔOL of the exhaust valve 9. The routine then proceeds to step 110.

In step 110, the mechanical compression ratio becomes the mechanical compression ratio CR by controlling the variable compression ratio mechanism A, and the closing timing of the intake valve 7 is controlled by controlling the intake variable valve timing mechanism B. In addition, an overlap period becomes an overlap period (ΔOL). In addition, the closing timing of the exhaust valve 9 becomes the closing timing EC by controlling the exhaust variable valve timing mechanism C. FIG.

Although the present invention has been described with reference to specific embodiments selected for the purpose of illustration, it should be apparent that numerous changes may be made to the embodiments by those skilled in the art without departing from the basic principles and contents of the invention.

Claims (19)

A variable compression ratio mechanism capable of changing the mechanical compression ratio, an intake variable valve timing mechanism capable of changing the closing timing of the intake valve, and an exhaust valve, In a spark ignition type internal combustion engine in which the amount of intake air supplied into the combustion chamber is controlled mainly by changing the closing timing of the intake valve, and the mechanical compression ratio is increased at the time of engine low load operation as compared to at the time of engine high load operation. In engine low load operation, the expansion ratio is 20 or more, A spark ignition type internal combustion engine at the engine low load operation, the closing timing of the exhaust valve is substantially an intake top dead center. A variable compression ratio mechanism capable of changing the mechanical compression ratio, an intake variable valve timing mechanism capable of changing the closing timing of the intake valve, and an exhaust variable valve timing mechanism capable of changing the closing timing of the exhaust valve, In a spark ignition type internal combustion engine in which the amount of intake air supplied into the combustion chamber is controlled mainly by changing the closing timing of the intake valve, and the mechanical compression ratio is increased at the time of engine low load operation as compared to at the time of engine high load operation. In engine low load operation, the expansion ratio is 20 or more, A spark ignition type internal combustion engine in which the closing timing of the exhaust valve can be set at the time of engine low load operation is limited to the intake top dead center side at the time of engine high load operation. 3. The spark ignition type internal combustion engine as set forth in claim 2, wherein in the engine low load operation, the closing timing of the exhaust valve is substantially intake top dead center. The spark ignition type internal combustion engine according to claim 2, wherein the closing timing of the exhaust valve and the opening timing of the intake valve are controlled to minimize the period in which the opening of the intake valve and the opening of the exhaust valve overlap during engine low load operation. The spark ignition type internal combustion engine according to claim 2, wherein the closing timing of the exhaust valve and the opening timing of the intake valve are controlled so that the period in which the opening of the intake valve and the opening of the exhaust valve overlap at the time of engine low load operation is zero. The spark ignition type internal combustion engine according to claim 1 or 2, wherein the opening timing of the intake valve at the engine low load operation is substantially an intake top dead center. 3. The spark ignition type internal combustion engine according to claim 1 or 2, wherein the actual compression ratio at the engine low load operation is substantially the same actual compression ratio as at the engine medium load operation. 8. The spark ignition type internal combustion engine as set forth in claim 7, wherein the actual compression ratio is lowered from 9 to 11 regardless of engine load at engine low speed. 9. The spark ignition type internal combustion engine according to claim 8, wherein the higher the engine speed, the higher the actual compression ratio. delete The spark ignition type internal combustion engine according to claim 1 or 2, wherein the amount of intake air supplied to the combustion chamber is controlled by changing the closing timing of the intake valve. 12. The spark ignition type internal combustion engine according to claim 11, wherein the closing timing of the intake valve is moved away from the dead center under intake until a limit closing timing for controlling the amount of intake air supplied to the combustion chamber as the engine load decreases. 13. The intake air according to claim 12, wherein the amount of intake air supplied to the combustion chamber is independent of the throttle valve disposed in the engine intake passage in a load region in which the closing timing of the intake valve is higher than the engine load when the closing timing is reached. Flame-ignition internal combustion engine controlled by changing the closing timing of the valve. The spark ignition type internal combustion engine according to claim 13, wherein the throttle valve is kept in a fully open state in a load region higher than the engine load when the closing timing of the intake valve reaches the limit closing timing. 13. The throttle valve disposed in the engine intake passage is used to control the amount of intake air supplied to the combustion chamber in a load region lower than the engine load when the closing timing of the intake valve reaches the limit closing timing. Spark-ignition internal combustion engine. The spark ignition type internal combustion engine according to claim 12, wherein the air-fuel ratio increases as the load decreases in a load region lower than the engine load when the closing timing of the intake valve reaches the limit closing timing. 13. The spark ignition type internal combustion engine according to claim 12, wherein the closing timing of the intake valve is maintained at the limit closing timing in a load region lower than the engine load when the closing timing of the intake valve reaches the threshold closing timing. 3. The spark ignition type internal combustion engine according to claim 1 or 2, wherein the mechanical compression ratio is increased to a limit mechanical compression ratio as the engine load is lowered. 19. The spark ignition type internal combustion engine as set forth in claim 18, wherein in the load region where the mechanical compression ratio is lower than the engine load when the limit mechanical compression ratio is reached, the mechanical compression ratio is maintained at the limit mechanical compression ratio.

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