Spark Ignition Type Internal Combustion Engine

Abstract

A spark ignition type internal combustion engine comprises a variable compression ratio mechanism able to change a mechanical compression ratio, an actual compression action start timing changing mechanism able to change a start timing of an actual compression action, and an exhaust valve. At the time of engine low load operation, the mechanical compression ratio is maximized to obtain a maximum expansion ratio, and the actual compression ratio is set so that no knocking occurs. The maximum expansion ratio is 20 or more. The closing timing of the exhaust valve at the time of engine low load operation is made substantially intake top dead center. Due to this, even if operating the internal combustion engine in a state of a large expansion ratio, the temperature of the exhaust purification catalyst can be maintained at a relatively high temperature.

Description

TECHNICAL FIELD

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

BACKGROUND ART

Known in the art is a spark ignition type internal combustion engine provided with a variable compression ratio mechanism able to change a mechanical compression ratio and a variable valve timing mechanism able to control a closing timing of an intake valve, performing a supercharging action by a supercharger at the time of engine medium load operation and engine high load operation and, in the state holding an actual compression ratio fixed at the time of engine medium and high load operation, increasing the mechanical compression ratio and retarding the closing timing of the intake valve as the engine load becomes lower (for example, see Japanese Patent Publication (A) No. 2004-218522).

However, in general, in an internal combustion engine, the larger the expansion ratio, the longer the period in an expansion stroke where a downward force acts on the piston, therefore the larger the expansion ratio, the more the heat efficiency is improved. Therefore, to raise the heat efficiency at the time of engine operation, it is preferable to make the mechanical compression ratio as high as possible and make the expansion ratio α large one.

However, if increasing the expansion ratio in this way, most of the heat energy produced in the combustion chamber is converted to kinetic energy, so the exhaust gas falls in temperature. Further, along with this, the pressure of the exhaust gas in the combustion chamber at the end of the expansion stroke also becomes lower and accordingly the exhaust gas becomes harder to exhaust from the combustion chamber. This tendency appears particularly remarkably when the expansion ratio is made 20 or more.

On the other hand, if the engine exhaust purification catalyst provided in the engine exhaust passage is not raised to a certain temperature or more, generally it cannot exhibit its excellent exhaust purification action. For this reason, in most internal combustion engines, the heat of the exhaust gas exhausted from the engine body is used to maintain the exhaust purification catalyst at a high temperature.

However, as explained above, if increasing the expansion ratio, the exhaust gas falls in temperature, so the temperature by which the exhaust purification catalyst is raised per unit flow rate becomes lower. Further, if increasing the expansion ratio, the exhaust gas becomes harder to be exhausted from the combustion chamber, so the flow rate of the exhaust gas flowing into the exhaust purification catalyst becomes smaller. For this reason, if operating the internal combustion engine in the state of a large expansion ratio, maintaining the exhaust purification catalyst at a high temperature becomes difficult.

DISCLOSURE OF INVENTION

Therefore, an object of the present invention is to provide a spark ignition type internal combustion engine able to maintain an exhaust purification catalyst at a relatively high temperature even when operating the internal combustion engine in the state of a large expansion ratio.

The present invention provides a spark ignition type internal combustion engine described in the claims of the claim section as means for realizing the above object.

In an aspect of the present invention, there is provided a spark ignition type internal combustion engine comprising a variable compression ratio mechanism able to change a mechanical compression ratio, an actual compression action start timing changing mechanism able to change a start timing of an actual compression action, and an exhaust valve, wherein at the time of engine low load operation the mechanical compression ratio is maximized to obtain a maximum expansion ratio and the actual compression ratio is set so that no knocking occurs, wherein the maximum expansion ratio is 20 or more, and wherein the closing timing of the exhaust valve at the time of engine low load operation is made substantially intake top dead center.

In another aspect of the present invention, there is provided a spark ignition type internal combustion engine comprising a variable compression ratio mechanism able to change a mechanical compression ratio, an actual compression action start timing changing mechanism able to change a start timing of an actual compression action, and an exhaust variable valve timing mechanism able to change the closing timing of the exhaust valve, wherein at the time of engine low load operation the mechanical compression ratio is maximized to obtain a maximum expansion ratio and the actual compression ratio is set so that no knocking occurs, wherein the maximum expansion ratio is 20 or more, and wherein a settable region of the closing timing of the exhaust valve at the time of engine low load operation is limited more to an intake top dead center side than that at the time of engine high load operation.

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

In another aspect of the present invention, the engine further comprises an intake variable valve timing mechanism able to change the opening timing of the intake valve, and the closing timing of the exhaust valve and the opening timing of the intake valve are controlled so that at the time of engine low load operation a period where the opening of the intake valve and the opening of the exhaust valve overlap is made minimum.

In another aspect of the present invention, the engine further comprises an intake variable valve timing mechanism able to change the opening timing of the intake valve, and the closing timing of the exhaust valve and the opening timing of the intake valve are controlled so that at the time of engine low load operation the period where the opening of the intake valve and the opening of the exhaust valve overlap, becomes zero.

In another aspect of the present invention, the engine further comprises an intake valve opening timing changing mechanism able to change the opening timing of the intake valve and, and at the time of engine low load operation, the opening timing of the intake valve is made substantially intake top dead center.

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

In another aspect of the present invention, at the time of engine low speed, regardless of the engine load, the actual compression ratio falls within a range of 9 to 11.

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

In another aspect of the present invention, the actual compression action start timing changing mechanism is comprised of an intake variable valve timing mechanism able to change the closing timing of the intake valve.

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

In another aspect of the present invention, the closing timing of the intake valve is shifted as the engine load becomes lower to a direction away from intake bottom dead center until a limit closing timing enabling control of the amount of intake air fed into the combustion chamber.

In another aspect of the present invention, in a region of a load higher than the engine load when the closing timing of the intake valve reaches the limit closing timing, the amount of intake air fed into the combustion chamber is controlled without regard to a throttle valve arranged in an engine intake passage by changing the closing timing of the intake valve.

In another aspect of the present invention, in a region of a load higher than the engine load when the closing timing of the intake valve reaches the limit closing timing, the throttle valve is held at the fully opened state.

In another aspect of the present invention, in a region of a load lower than the engine load when the closing timing of the intake valve reaches the limit closing timing, a throttle valve arranged in an engine intake passage is used to control the amount of intake air fed into the combustion chamber.

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

In another aspect of the present invention, in a region of a 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 held at the limit closing timing.

In another aspect of the present invention, the mechanical compression ratio is increased as the engine load becomes lower to the limit mechanical compression ratio.

In another aspect of the present invention, in a region of a load lower than the engine load when the mechanical compression ratio reaches the limit mechanical compression ratio, the mechanical compression ratio is held 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, even if operating the internal combustion engine in the state of a large expansion ratio, the exhaust purification catalyst can be maintained at a relatively high temperature.

BRIEF DESCRIPTION OF THE DRAWINGS

The present invention will be more clearly understood from the description as set forth below with reference to the accompanying drawings, in which:

FIG. 1 is an overview of a spark ignition type internal combustion engine.

FIG. 2 is a disassembled perspective view of a variable compression ratio mechanism.

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

FIG. 4 is a view of a variable valve timing mechanism.

FIGS. 5A and 5B are views showing the amounts of lift of the intake valve and exhaust valve.

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

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

FIGS. 8A and 8B are views for explaining a normal cycle and superhigh expansion ratio cycle.

FIG. 9 is a view showing the change in mechanical compression ratio etc. in accordance with the engine load.

FIGS. 10A, 10B and 10C are views showing the changes in lift of the intake valve and exhaust valve.

FIG. 11 is a view showing a region in which a closing timing of the exhaust valve in accordance with the mechanical compression ratio can be set.

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

FIG. 13 is a flowchart for operational control.

FIGS. 14A, 14B and 14C are views showing the target actual compression ratio etc.

FIGS. 15A and 15B are views showing a map of the closing timing of the exhaust valve etc.

BEST MODE FOR CARRYING OUT THE INVENTION

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

Referring to FIG. 1, 1 indicates a crank case, 2 a cylinder block, 3 a cylinder head, 4 a piston, 5 a combustion chamber, 6 a spark plug arranged at the top center of the combustion chamber 5, 7 an intake valve, 8 an intake port, 9 an exhaust valve, and 10 an exhaust port. The intake port 8 is connected through an intake tube 11 to a surge tank 12, while each intake tube 11 is provided with a fuel injector 13 for injecting fuel toward a corresponding intake port 8. Note that each fuel injector 13 may be arranged at each combustion chamber 5 instead of being attached to each intake tube 11.

The surge tank 12 is connected via an intake duct 14 to an outlet of the compressor 15a of the exhaust turbocharger 15, while an inlet of the compressor 15a is connected through an intake air amount detector 16 using for example a heating wire to an air cleaner 17. The intake duct 14 is provided inside it with a throttle valve 19 driven by an actuator 18.

On the other hand, the exhaust port 10 is connected through the exhaust manifold 20 to the inlet of the exhaust turbine 15b of the exhaust turbocharger 15, while an outlet of the exhaust turbine 15b is connected through an exhaust pipe 21 to a catalytic converter 22 housing an exhaust purification catalyst. The exhaust pipe 21 has an air-fuel ratio sensor 23 arranged in it.

Further, in the embodiment shown in FIG. 1, the connecting part of the crank case 1 and the cylinder block 2 is provided with a variable compression ratio mechanism A able to change the relative positions of the crank case 1 and cylinder block 2 in the cylinder axial direction so as to change the volume of the combustion chamber 5 when the piston 4 is positioned at compression top dead center. Further, it is provided with an intake variable valve timing mechanism B able to control the closing timing of the intake valve 7 to change the start timing of the actual compression action, and able to individually control the opening timing of the intake valve 7. Further, it is provided with an exhaust variable valve timing mechanism C able to individually control the opening timing and closing timing of the exhaust valve 7.

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

FIG. 2 is a disassembled perspective view of the variable compression ratio mechanism A shown in FIG. 1, while FIGS. 3A and 3B are side cross-sectional views of the illustrated internal combustion engine. Referring to FIG. 2, at the bottom of the two side walls of the cylinder block 2, a plurality of projecting parts 50 separated from each other by a certain distance are formed. Each projecting part 50 is formed with a circular cross-section cam insertion hole 51. On the other hand, the top surface of the crank case 1 is formed with a plurality of projecting parts 52 separated from each other by a certain distance and fitting between the corresponding projecting parts 50. These projecting parts 52 are also formed with circular cross-section cam insertion holes 53.

As shown in FIG. 2, a pair of cam shafts 54, 55 is provided. Each of the cam shafts 54, 55 has circular cams 56 fixed on it able to be rotatably inserted in the cam insertion holes 51 at every other position. These circular cams 56 are coaxial with the axes of rotation of the cam shafts 54, 55. On the other hand, between the circular cams 56, as shown by the hatching in FIGS. 3A and 3B, extend eccentric shafts 57 arranged eccentrically with respect to the axes of rotation of the cam shafts 54, 55. Each eccentric shaft 57 has other circular cams 58 rotatably attached to it eccentrically. As shown in FIG. 2, these circular cams 58 are arranged between the circular cams 56. These circular cams 58 are rotatably inserted in the corresponding cam insertion holes 53.

When the circular cams 56 fastened to the cam shafts 54, 55 are rotated in opposite directions from each other as shown by the solid line arrows in FIG. 3A from the state shown in FIG. 3A, the eccentric shafts 57 move toward the bottom center, so the circular cams 58 rotate in the opposite directions from the circular cams 56 in the cam insertion holes 53 as shown by the broken line arrows in FIG. 3A. As shown in FIG. 3B, when the eccentric shafts 57 move toward the bottom center, the centers of the circular cams 58 move to below the eccentric shafts 57.

As will be understood from a comparison of FIGS. 3A and 3B, the relative positions of the crank case 1 and cylinder block 2 are determined by the distance between the centers of the circular cams 56 and the centers of the circular cams 58. The larger the distance between the centers of the circular cams 56 and the centers of the circular cams 58, the further the cylinder block 2 from the crank case 1. If the cylinder block 2 moves farther away from the crank case 1, the volume of the combustion chamber 5 when the piston 4 is positioned at compression top dead center increases. Therefore, by making the cam shafts 54, 55 rotate, the volume of the combustion chamber 5 when the piston 4 is positioned at compression top dead center can be changed.

As shown in FIG. 2, to make the cam shafts 54, 55 rotate in opposite directions, the shaft of a drive motor 59 is provided with a pair of worm gears 61, 62 with opposite thread directions. Gears 63, 64 engaging with these worm gears 61, 62 are fastened to ends of the cam shafts 54, 55. In this embodiment, the drive motor 59 may be driven to change the volume of the combustion chamber 5 when the piston 4 is positioned at compression top dead center over a broad range. Note that the variable compression ratio mechanism A shown from FIGS. 1 to 3 shows an example. Any type of variable compression ratio mechanism may be used.

Further, FIG. 4 shows an intake variable valve timing mechanism B attached to the cam shaft 70 for driving the intake valve 7 in FIG. 1. Referring to FIG. 4, the intake variable valve timing mechanism B is comprised of a cam phase changer B1 attached to one end of the cam shaft 70 and changing the phase of the cam of the cam shaft 70 and a cam actuation angle changer B2 arranged between the cam shaft 70 and the valve lifter 24 of the intake valve 7 and changing the actuation angle (working angle) of the cams of the cam shaft 70 to different actuation angles for transmission to the intake valve 7. Note that FIG. 4 is a side sectional view and plan view of the cam actuation angle changer B2.

First, explaining the cam phase changer B1 of the intake variable valve timing mechanism B, this cam phase changer B1 is provided with a timing pulley 71 made to rotate by an engine crank shaft through a timing belt in the arrow direction, a cylindrical housing 72 rotating together with the timing pulley 71, a shaft 73 able to rotate together with a cam shaft 70 and rotate relative to the cylindrical housing 72, a plurality of partitions 74 extending from an inside circumference of the cylindrical housing 72 to an outside circumference of the shaft 73, and vanes 75 extending between the partitions 74 from the outside circumference of the shaft 73 to the inside circumference of the cylindrical housing 72, the two sides of the vanes 75 formed with advancing use hydraulic chambers 76 and retarding use hydraulic chambers 77.

The feed of working oil to the hydraulic chambers 76, 77 is controlled by a working oil feed control valve 78. This working oil feed control valve 78 is provided with hydraulic ports 79, 80 connected to the hydraulic chambers 76, 77, a feed port 82 for working oil discharged from a hydraulic pump 81, a pair of drain ports 83, 84, and a spool valve 85 for controlling connection and disconnection of the ports 79, 80, 82, 83, 84.

To advance the phase of the cams of the cam shaft 70, the spool valve 85 is made to move to downward in FIG. 4, working oil fed from the feed port 82 is fed through the hydraulic port 79 to the advancing use hydraulic chambers 76, and working oil in the retarding use hydraulic chambers 77 is drained 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.

As opposed to this, to retard the phase of the cam of the cam shaft 70, the spool valve 85 is made to move upward in FIG. 4, working oil fed from the feed port 82 is fed through the hydraulic port 80 to the retarding use hydraulic chambers 77, and working oil in the advancing use hydraulic chambers 76 is drained from the drain port 83. At this time, the shaft 73 is made to rotate relative to the cylindrical housing 72 in the direction opposite to the arrows X.

When the shaft 73 is made to rotate relative to the cylindrical housing 72, if the spool valve 85 is returned to the neutral position shown in FIG. 4, the operation for relative rotation of the shaft 73 is ended, and the shaft 73 is held at the relative rotational position at that time. Therefore, it is possible to use the cam phase changer B1 so as to advance or retard the phase of the cam of the cam shaft 70 by exactly the desired amount. That is, the cam phase changer B1 can freely advance or retard the opening timing of the intake valve 7.

Next, explaining the cam actuation angle changer B2 of the intake variable valve timing mechanism B, this cam actuation angle changer B2 is provided with a control rod 90 arranged in parallel with the cam shaft 70 and made to move by an actuator 91 in the axial direction, an intermediate cam 94 engaging with a cam 92 of the cam shaft 70 and slidably fitting with a spline 93 formed on the control rod 90 and extending in the axial direction, and a pivoting cam 96 engaging with a valve lifter 24 for driving the intake valve 7 and slidably fitting with a spline 95 extending in a spiral formed on the control rod 90. The pivoting cam 96 is formed with a cam 97.

When the cam shaft 70 rotates, the cam 92 causes the intermediate cam 94 to pivot by exactly a constant angle at all times. At this time, the pivoting cam 96 is also made to pivot by exactly a constant angle. On the other hand, the intermediate cam 94 and pivoting cam 96 are supported not movably in the axial direction of the control rod 90, therefore when the control rod 90 is made to move by the actuator 91 in the axial direction, 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 with the intermediate cam 94 due to the relative rotational positional relationship between the intermediate cam 94 and pivoting cam 96, if the cam 97 of the pivoting cam 96 starts to engage with the valve lifter 24, as shown by a in FIG. 5B, the opening time period and amount of lift of the intake valve 7 become maximum. As opposed to this, when the actuator 91 is used to make the pivoting cam 96 rotate relative to the intermediate cam 94 in the arrow Y-direction of FIG. 4, the cam 92 of the cam shaft 70 engages with the intermediate cam 94, then after a while the cam 97 of the pivoting cam 96 engages with the valve lifter 24. In this case, as shown by b in FIG. 5B, the opening time period and amount of lift 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 arrow Y-direction of FIG. 4, as shown by c in FIG. 5B, the opening time period and amount of lift of the intake valve 7 become further smaller. That is, by using the actuator 91 to change the relative rotational position of the intermediate cam 94 and pivoting cam 96, the opening time period of the intake valve 7 can be freely changed. However, in this case, the amount of the lift of the intake valve 7 becomes smaller the shorter the opening time of the intake valve 7.

The cam phase changer B1 can be used to freely change the opening timing of the intake valve 7 and the cam actuation angle changer B2 can be used to freely change the opening time period of the intake valve 7 in this way, so both the cam phase changer B1 and cam actuation angle changer B2, that is, the intake variable valve timing mechanism B, may be used to freely change the opening timing and opening time period of the intake valve 7, that is, the opening timing and closing timing of the intake valve 7.

Note that the intake variable valve timing mechanism B shown in FIGS. 1 and 4 shows an example. It is also possible to use various types of variable valve timing mechanisms other than the example shown in FIGS. 1 and 4.

Further, the exhaust variable valve timing mechanism C also basically has a configuration similar to the intake variable valve timing mechanism B and can freely change the opening timing and opening time period of the exhaust valve 9, that is, the opening timing and closing timing of the exhaust valve 9.

Next, the meaning of the terms used in the present application will be explained with reference to FIGS. 6A to 6C. Note that FIGS. 6A, 6B and 6C show for explanatory purposes an engine with a volume of the combustion chambers of 50 ml and a stroke volume of the piston of 500 ml. In these FIGS. 6A, 6B and 6C, the combustion chamber volume shows the volume of the combustion chamber when the piston is at compression top dead center.

FIG. 6A explains the mechanical compression ratio. The mechanical compression ratio is a value determined mechanically from the stroke volume of the piston and combustion chamber volume at the time of a compression stroke. 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 becomes (50 ml+500 ml)/50 ml=11.

FIG. 6B explains the actual compression ratio. This actual compression ratio is a value determined from the actual stroke volume of the piston from when the compression action is actually started to when the piston reaches top dead center and the combustion chamber volume. This actual compression ratio is expressed by (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, no compression action is performed while the intake valve is opened. The actual compression action is started after the intake valve closes. 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 becomes (50 ml+450 ml)/50 ml=10.

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

Next, the most basic features of the present invention will be explained with reference to FIGS. 7, 8A and 8B. Note that FIG. 7 shows the relationship between the theoretical thermal efficiency and the expansion ratio, while FIGS. 8A and 8B show a comparison between the normal cycle and superhigh expansion ratio cycle used selectively in accordance with the load in the present invention.

FIG. 8A shows the normal cycle in which the intake valve closes near the bottom dead center and the compression action by the piston is started from near substantially compression bottom dead center. In the example shown in this FIG. 8A as well, in the same way as the examples shown in FIGS. 6A, 6B and 6C, the combustion chamber volume is made 50 ml, and the stroke volume of the piston is made 500 ml. As will be understood 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 also becomes (50 ml+500 ml)/50 ml=11. That is, in an ordinary internal combustion engine, the mechanical compression ratio and actual compression ratio and the expansion ratio become substantially equal.

The solid line in FIG. 7 shows the change in the theoretical thermal efficiency in the case where the actual compression ratio and expansion ratio are substantially equal, that is, in the normal cycle. In this case, it is learned that the larger the expansion ratio, that is, the higher the actual compression ratio, the higher the theoretical thermal efficiency. Therefore, in a normal cycle, to raise the theoretical thermal efficiency, the actual compression ratio should be made higher. However, due to the restrictions on the occurrence of knocking at the time of engine high load operation, the actual compression ratio can only be raised even at the maximum to about 12, accordingly, in a normal cycle, the theoretical thermal efficiency cannot be made sufficiently high.

On the other hand, under this situation, the inventors strictly differentiated between the mechanical compression ratio and actual compression ratio and studied the theoretical thermal efficiency and as a result discovered that in the theoretical thermal efficiency, the expansion ratio is dominant, and the theoretical thermal efficiency is not affected much at all by the actual compression ratio. That is, if raising the actual compression ratio, the explosive force rises, but compression requires a large energy, accordingly even if raising the actual compression ratio, the theoretical thermal efficiency will not rise much at all.

As opposed to this, if increasing the expansion ratio, the longer the period during which a force acts pressing down the piston at the time of the expansion stroke, the longer the time that the piston gives a rotational force to the crankshaft. Therefore, the larger the expansion ratio is made, the higher the theoretical thermal efficiency becomes. The broken line in FIG. 7 shows the theoretical thermal efficiency in the case of fixing the actual compression ratio at 10 and raising the expansion ratio in that state. In this way, it is learned that the amount of rise of the theoretical thermal efficiency when raising the expansion ratio in the state where the actual compression ratio is maintained at a low value and the amount of rise of the theoretical thermal efficiency in the case where the actual compression ratio is increased along with the expansion ratio as shown by the solid line of FIG. 7 will not differ that much.

If the actual compression ratio is maintained at a low value in this way, knocking will not occur, therefore if raising the expansion ratio in the state where the actual compression ratio is maintained at a low value, the occurrence of knocking can be prevented and the theoretical thermal efficiency can be greatly raised. FIG. 8B shows an example of the case when using the variable compression ratio mechanism A and variable valve timing mechanism B to maintain the actual compression ratio at a low value and raise the expansion ratio.

Referring to FIG. 8B, in this example, the variable compression ratio mechanism A is used to lower the combustion chamber volume from 50 ml to 20 ml. On the other hand, the intake variable valve timing mechanism B is used to retard the closing timing of the intake valve until the actual stroke volume of the piston changes from 500 ml to 200 ml. As a result, in this example, the actual compression ratio becomes (20 ml+200 ml)/20 ml=11 and the expansion ratio becomes (20 ml+500 ml)/20 ml=26. In the normal cycle shown in FIG. 8A, as explained above, the actual compression ratio is about 11 and the expansion ratio is 11. Compared with this case, in the case shown in FIG. 8B, it is learned that only the expansion ratio is raised to 26. This will be called the “superhigh expansion ratio cycle” below.

As explained above, generally speaking, in an internal combustion engine, the lower the engine load, the worse the heat efficiency, therefore to improve the heat efficiency at the time of vehicle operation, that is, to improve the fuel efficiency, it becomes necessary to improve the heat efficiency at the time of engine low load operation. On the other hand, in the superhigh expansion ratio cycle shown in FIG. 8B, the actual stroke volume of the piston at the time of the compression stroke is made smaller, so the amount of intake air which can be sucked into the combustion chamber 5 becomes smaller, therefore this superhigh expansion ratio cycle can only be employed when the engine load is relatively low. Therefore, in the present invention, at the time of engine low load operation, the superhigh expansion ratio cycle shown in FIG. 8B is set, while at the time of engine high load operation, the normal cycle shown in FIG. 8A is set. This is the basic feature of the present invention.

FIG. 9 shows the operational control as a whole at the time of steady operation at a low engine speed. Below, the operational control as a whole will be explained with reference to FIG. 9.

FIG. 9 shows the changes in the mechanical compression ratio, expansion ratio, closing timing of the intake valve 7, actual compression ratio, the amount of intake air, opening degree of the throttle valve 17, and pumping loss, along with the engine load. Note that in the present embodiment, to enable the three-way catalyst in the catalytic converter 22 to simultaneously reduce the unburned HC, CO, and NO_x in the exhaust gas, ordinarily the average air-fuel ratio in the combustion chamber 5 is feedback controlled to the stoichiometric air-fuel ratio based on the output signal of the air-fuel ratio sensor 23.

Now, as explained above, at the time of engine high load operation, the normal cycle shown in FIG. 8A is executed. Therefore, as shown in FIG. 9, at this time, the mechanical compression ratio is made low, so the expansion ratio becomes low and, as shown by the solid line in FIG. 9, the closing timing of the intake valve 7 (is advanced. Further, at this time, the amount of intake air is large. At this time, the opening degree of the throttle valve 17 is maintained fully opened or substantially fully opened, so the pumping loss becomes zero.

On the other hand, as shown in FIG. 9, along with the reduction in the engine load, the mechanical compression ratio is increased, therefore the expansion ratio is also increased. Further, at this time, the actual compression ratio is held substantially constant by, as shown by the solid line in FIG. 9, retarding the closing timing of the intake valve 7 as the engine load becomes lower. Note that at this time as well, the throttle valve 17 is held at the fully opened or substantially fully opened state, therefore the amount of intake air fed to the combustion chamber 5 is controlled not by the throttle valve 17 but by changing the closing timing of the intake valve 7. At this time as well, the pumping loss becomes zero.

In this way, when the engine load becomes lower from the engine high load operating state, the mechanical compression ratio is increased along with the fall in the amount of intake air under a substantially constant actual compression ratio. That is, the volume of the combustion chamber 5 when the piston 4 reaches compression top dead center is reduced proportionally to the reduction in the amount of intake air. Therefore the volume of the combustion chamber 5 when the piston 4 reaches compression top dead center changes proportionally to the amount of intake air. Note that at this time, the air-fuel ratio in the combustion chamber 5 becomes the stoichiometric air-fuel ratio, so the volume of the combustion chamber 5 when the piston 4 reaches compression top dead center changes proportionally to the amount of fuel.

If the engine load becomes further lower, the mechanical compression ratio is further increased. When the mechanical compression ratio reaches the limit mechanical compression ratio corresponding to the structural limit of the combustion chamber 5, in the region of a load lower than the engine load L1 when the mechanical compression ratio reaches the limit mechanical compression ratio, the mechanical compression ratio is held at the limit mechanical compression ratio. Therefore at the time of engine low load operation, the mechanical compression ratio becomes maximum, and the expansion ratio also becomes maximum. Putting this another way, in the present invention, so as to obtain the maximum expansion ratio at the time of engine low load operation, the mechanical compression ratio is made maximum. Further, at this time, the actual compression ratio is maintained at an actual compression ratio substantially the same as that at the time of engine medium and high load operation.

On the other hand, as shown by the solid line in FIG. 9, the closing timing of the intake valve 7 is retarded further to the limit closing timing enabling control of the amount of intake air fed to the combustion chamber 5 the more the engine load becomes lower. In the region of a load 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 held at the limit closing timing. If the closing timing of the intake valve 7 is held at the limit closing timing, the amount of intake air will no longer be able to be controlled by the change of the closing timing of the intake valve 7. Therefore, the amount of intake air has to be controlled by some other method.

In the embodiment shown in FIG. 9, at this time, that is, in the region of a load 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 used to control the amount of intake air fed to the combustion chamber 5. However, if the throttle valve 17 (is used to control the amount of intake air, as shown in FIG. 9, the pumping loss increases.

Note that to prevent such pumping loss from occurring, in the region of a load lower than the engine load L2 when the closing timing of the intake valve 7 reaches the limit closing timing, in the state holding the throttle valve 17 fully opened or substantially fully opened, the air-fuel ratio may be made larger the lower the engine load. 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 time of engine low speed, regardless of the engine load, the actual compression ratio is held substantially constant. The actual compression ratio at this time is made the range of the actual compression ratio about at the time of engine medium and high load operation ±10 percent, preferably ±5 percent. Note that in the present embodiment, the actual compression ratio at the time of engine low speed is made about 10±1, that is, from 9 to 11. However, if the engine speed becomes higher, the air-fuel mixture in the combustion chamber 5 is disturbed, so knocking becomes harder to occur, therefore in the embodiment according to the present invention, the higher the engine speed, the higher the actual compression ratio.

On the other hand, as explained above, in the superhigh expansion ratio cycle shown in FIG. 8B, the expansion ratio is made 26. The higher this expansion ratio, the better, but if 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 becomes 20 or more.

Further, in the example shown in FIG. 9, the mechanical compression ratio is changed continuously in accordance with the engine load. However, the mechanical compression ratio can also be changed in stages in accordance with the engine load.

On the other hand, as shown by the broken line in FIG. 9, as the engine load becomes lower, by advancing the closing timing of the intake valve 7 as well, it is possible to control the amount of intake air without depending on the throttle valve. Therefore, in FIG. 9, if comprehensively expressing 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 closing timing of the intake valve 7 is shifted, as the engine load becomes lower, in a direction away from compression bottom dead center BDC until the limit closing timing L₂ enabling control of the amount of intake air fed into the combustion chamber.

Next, the closing timing of the exhaust valve 9 will be explained focusing on the low load operation where the superhigh expansion ratio cycle shown in FIG. 8B is executed.

In general, at the time of low load operation where a superhigh expansion ratio cycle is executed, the amount of heat generated due to combustion of the air-fuel mixture in the combustion chamber 5 itself is small, so the temperature of the exhaust gas exhausted from the combustion chamber 5 easily becomes low. In addition to this, in an internal combustion engine, the larger the expansion ratio, the longer the period during which a force pushing down the piston acts at the time of an expansion stroke, so most of the heat energy produced by combustion of the air-fuel mixture in the combustion chamber is converted to kinetic energy of the piston. Along with this, the temperature of the combustion gas in the combustion chamber at the end of the expansion stroke becomes lower. For this reason, when the superhigh expansion ratio cycle shown in FIG. 8B is executed, at the time of an exhaust stroke, the temperature of the exhaust gas exhausted from the combustion chamber 5 to the exhaust manifold 20 becomes extremely low. This tendency appears particularly remarkably when the expansion ratio is made 20 or more. Between execution of a superhigh expansion ratio cycle where the expansion ratio is made 20 or more and a normal cycle where the expansion ratio is made 12 or so, the temperature of the exhaust gas exhausted from the combustion chamber 5 differs by about 100° C.

On the other hand, in most internal combustion engines, the harmful ingredients contained in exhaust gas (for example, HC, CO, NO_x, etc.) are removed by providing inside the engine exhaust passage a three-way catalyst, NO_x storing and reducing catalyst, or other exhaust purification catalyst. Such an exhaust purification catalyst cannot effectively remove the harmful ingredients in the exhaust gas unless its temperature becomes the activation temperature or more. Here, in most internal combustion engines, the temperature of the exhaust gas is considerably higher than the activation temperature, so the exhaust gas is made to flow into the exhaust purification catalyst to maintain the temperature of the exhaust purification catalyst at the activation temperature or more.

However, if the superhigh expansion ratio cycle shown in FIG. 8B is executed, the temperature of the exhaust gas exhausted from the combustion chamber 5 to the exhaust manifold 20 will only become slightly higher than the activation temperature, so even if making the exhaust gas flow into the exhaust purification catalyst, it becomes difficult to maintain the temperature of the exhaust purification catalyst at the activation temperature or more. Therefore, when the superhigh expansion ratio cycle is executed, to maintain the temperature of the exhaust purification catalyst at the activation temperature or more, it is necessary to make as much exhaust gas as possible flow into the exhaust purification catalyst.

Here, referring to FIGS. 10A to 10C, let us consider the relationship between the closing timing of the exhaust valve 9 and the flow rate of exhaust gas exhausted from the combustion chamber 5 to the exhaust manifold 20. FIG. 10A shows the changes in lift of the exhaust valve 9 and the intake valve 7 in the case where the exhaust valve 9 is closed at substantially intake top dead center, FIG. 10B shows the same in case where the exhaust valve 9 is closed before intake top dead center, while FIG. 10C shows the same in the case where the exhaust valve 9 is closed after intake top dead center.

As shown in FIG. 10B, when closing the exhaust valve 9 before intake top dead center, the volume of the combustion chamber 5 when closing the exhaust valve 9 is larger than the volume of the combustion chamber when the piston is positioned at intake top dead center (combustion chamber volume). After the exhaust valve 9 closes, exhaust gas corresponding to the volume of the combustion chamber 5 at the time of closing remains in the combustion chamber 5. For this reason, even after the exhaust valve 9 closes, a relatively large amount of exhaust gas remains in the combustion chamber 5. Therefore, it is not possible to sufficiently exhaust the exhaust gas in the combustion chamber 5 to the exhaust manifold 20 and the flow rate of exhaust gas into the exhaust purification catalyst becomes small.

On the other hand, as shown in FIG. 10C, when closing the exhaust valve 9 after intake top dead center, the exhaust valve 9 is open even at intake top dead center, so when the piston 4 reaches intake top dead center, almost all of the exhaust gas in the combustion chamber 5 flows out into the exhaust port 10. However, if the exhaust valve 9 is open even after intake top dead center, part of the exhaust gas flowing out once into the exhaust port 10 will end up again flowing into the combustion chamber 5 along with the descent of the piston 4.

In particular, when the superhigh expansion ratio cycle is executed, at the time of the expansion stroke, the combustion gas in the combustion chamber 5 considerably expands, so the pressure of the combustion gas at the end of the expansion stroke will be relatively low. For this reason, the strength of the exhaust gas flowing out from the combustion chamber 5 to the exhaust port 10 at the exhaust stroke will be weak. Therefore, if the piston 4 descends after reaching intake top dead center, part of the exhaust gas flowing out into the exhaust port 10 will again easily flow into the combustion chamber 5.

In this way, when closing the exhaust valve 9 after intake top dead center, the exhaust gas flowing out once into the exhaust port 10 will again return to the inside of the combustion chamber 5, so the exhaust gas in the combustion chamber 5 will not be able to be sufficiently exhausted to the exhaust manifold 20 and the flow rate of exhaust gas flowing into the exhaust purification catalyst will be small.

Therefore, in the present embodiment, when the superhigh expansion ratio cycle shown in FIG. 8B is executed, that is, when the mechanical compression ratio is high, to prevent the closing timing of the exhaust valve 9 from being too much earlier or too much later than intake top dead-center, the region where the closing timing of the exhaust valve 9 can be set is limited to the intake top dead center side.

FIG. 11 is a view showing a region in which the closing timing of the exhaust valve 9 in accordance with the mechanical compression ratio can be set.

As shown in FIG. 11, the region in which the exhaust valve 9 can be set becomes the region between the settable maximum amount of advance and maximum amount of retardation. As will be understood from the figure, the amount of advance by which the closing timing of the exhaust valve 9 can be set is made smaller (later) the higher the mechanical compression ratio, while conversely the maximum amount of retardation by which the closing timing of the exhaust valve 9 can be set is made smaller (earlier) the higher the mechanical compression ratio. For this reason, the region in which the closing timing of the exhaust valve 9 can be set becomes smaller the higher the mechanical compression ratio, that is, is more restricted the higher the mechanical compression ratio. 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, while when the mechanical compression ratio is high, the region in which the closing timing of the exhaust valve 9 can be set is made ΔTOC2 (ΔTOC2<ΔTOC1).

Alternatively, when the superhigh expansion ratio cycle shown in FIG. 8B is executed, that is, when the mechanical compression ratio is high, to reliably prevent the closing timing of the exhaust valve 9 from becoming too advanced from or too retarded from the intake top dead center, as shown in FIG. 10A, the closing timing of the exhaust valve 9 may be made substantially intake top dead center.

In this way, when the mechanical compression ratio is high, by limiting the region in which the closing timing of the exhaust valve 9 can be set to the intake top dead center side or making the closing timing of the exhaust valve 9 substantially intake top dead center, it is possible to sufficiently exhaust the exhaust gas in the combustion chamber 5 to the exhaust manifold 20 and make the flow rate of exhaust gas flowing into the exhaust purification large.

That is, the exhaust valve 9 is made to close near intake top dead center, so as shown in FIG. 10B, compared with closing the exhaust valve 9 in advance of intake top dead center, the volume of the combustion chamber 5 at the time of closing of the exhaust valve 9 is small and therefore it is possible to reduce the amount of exhaust gas remaining in the combustion chamber 5 after closing the exhaust valve 9. Further, the exhaust valve 9 is made to close near intake top dead center, so as shown in FIG. 10C, compared with when closing the exhaust valve 9 retarded from intake top dead center, the amount of exhaust gas flowing into the combustion chamber 5 in the exhaust gas flowing out into the exhaust port 10 can be reduced. For this reason, as shown in FIG. 10A, when the exhaust valve 9 is made to close near intake top dead center, as shown in FIGS. 10B and 10C, compared with when making the exhaust valve 9 close away from intake top dead center, the exhaust gas in the combustion chamber 5 can be sufficiently exhausted into the exhaust manifold 20 and the flow rate of the exhaust gas flowing into the exhaust purification catalyst can be increased. As a result, even at the time of low load operation where the superhigh expansion ratio cycle is executed, it is possible to maintain the exhaust purification catalyst at the activation temperature or more.

Note that the “substantially intake top dead center” indicates within 10° before and after intake top dead center, preferably within 5° before and after intake top dead center.

Further, if raising the mechanical compression ratio, the combustion chamber volume at intake top dead center becomes smaller and accordingly depending on the closing timing of the exhaust valve 9, the exhaust valve 9 will end up interfering with the piston 4.

FIGS. 10A to 10C show the piston interference line showing the limit where the exhaust valve 9 or intake valve 7 interferes with the piston 4. When the lift curve of the exhaust valve 9 interferes with 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 with the piston interference line. This means that when closing the exhaust valve 9 retarded from intake top dead center, while depending also on the extent of retardation, the exhaust valve 9 and piston 4 will end up interfering.

As opposed to this, according to the present embodiment, when the mechanical compression ratio is high, the region in which the closing timing of the exhaust valve 9 can be set is limited to the intake top dead center side, in particular the amount of maximum retardation by which the closing timing of the exhaust valve 9 can be set is made smaller. For this reason, as shown in FIG. 10A, even if the mechanical compression ratio becomes higher, the exhaust valve 9 can be prevented from interfering with the piston 4.

However, when there is a valve overlap where the opening time period of the intake valve 7 and the opening time period of the exhaust valve 9 overlap, the amount of exhaust gas exhausted from inside of the combustion chamber 5 to the exhaust manifold 20 changes even during that period. Below, referring to FIGS. 12A and 12B, consider the relationship between the overlap period where the opening time period of the intake valve 7 and the opening time period of the exhaust valve 9 overlap and the amount of exhaust gas exhausted from the combustion chamber 5 to the exhaust manifold 20. FIG. 12A shows the case where the overlap period is zero, while FIG. 12B shows the changes in lifts of the exhaust valve 9 and the intake valve 7 when the overlap period is large.

In general, when the intake valve 7 and exhaust valve 9 are simultaneously opened, part of the exhaust gas in the combustion chamber 5 and part of the exhaust gas once flowing out from the combustion chamber 5 to the exhaust port 10 will sometimes flow into the intake port 8. In this way, when part of the exhaust gas flows into the intake port 8, the exhaust gas exhausted from the combustion chamber 5 to the exhaust manifold 20 will become smaller by that amount.

Therefore, when the overlap period is large as shown in FIG. 12B, the exhaust gas will often flow into the intake port 8 in a large amount. Therefore, the exhaust gas flowing out from the combustion chamber 5 to the exhaust manifold 20 will often become smaller. For this reason, in this case, the flow rate of the exhaust gas flowing into the exhaust purification catalyst will become smaller.

Therefore, in this embodiment, when the superhigh expansion ratio cycle shown in FIG. 8B is executed, that is, when the mechanical compression ratio is high, as shown in FIG. 12A, the closing timing of the exhaust valve 7 and the opening timing of the intake valve 9 are controlled to become minimum in the range in which the overlap period can be set. Therefore, for example, in an internal combustion engine where the settable overlap period becomes 10° to 60°, when the mechanical compression ratio is high, the overlap period, is made 10°, while in an internal combustion engine where the settable overlap period becomes 0° to 50°, when the mechanical compression ratio is high, the overlap period is made 0°.

In this way, when the mechanical compression ratio is high, by minimizing the overlap period, the exhaust gas flowing into the intake port 8 becomes smaller, so the exhaust gas exhausted from the combustion chamber 5 to the exhaust manifold 20 becomes great and accordingly the flow rate of the exhaust gas flowing into the exhaust purification catalyst becomes greater.

Note that the overlap period when the mechanical compression ratio is high need not necessarily be made the minimum so long as 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 need only be 10° or less of the settable range even at the minimum.

Further, as explained above, if raising the mechanical compression ratio, the combustion chamber volume at intake top dead center becomes smaller. Accordingly, depending on the opening timing of the intake valve 7, the intake valve 7 will end up interfering with the piston 4.

FIGS. 12A and 12B show the piston interference line showing the limit where the exhaust valve 9 or intake valve 7 interferes with the piston 4. If the lift curve of the intake valve 7 intersects with the piston interference line, the intake valve 7 will interfere with the piston 4. Here, in FIG. 12B, the lift curve of the intake valve 7 intersects with the piston interference line. This means that if increasing the overlap period, the intake valve 7 and piston 4 will end up interfering with each other. That is, in the present embodiment, as explained above, the closing timing of the exhaust valve 9 is made substantially intake top dead center. The overlap period being large means that the opening timing of the intake valve 7 is made to greatly advance. If the opening timing of the intake valve 7 is made to greatly advance, the intake valve 7 and the piston 4 will end up interfering with each other.

As opposed to this, according to the present embodiment, when the mechanical compression ratio is high, the overlap period is made the minimum, so the opening timing of the intake valve 7 is made substantially intake top dead center or less. For this reason, as shown in FIG. 12A, even if the mechanical compression ratio becomes high, the intake valve 7 can be prevented from interfering with the piston.

FIG. 13 shows a control routine of operational control of a spark ignition type internal combustion engine of the present embodiment. Referring to FIG. 13, first, at step 101, the engine load L and engine speed Ne are fetched. Next, at 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 becomes higher the higher the engine speed Ne. Next, at step 103, the map shown in FIG. 14B is used to calculate the mechanical compression ratio CR. That is, the mechanical compression ratio CR required for making the actual compression ratio the target actual compression ratio is stored as a function of the engine load L and engine speed Ne in the form of a map as shown in FIG. 14B in advance in the ROM 32. This map is used to calculate the mechanical compression ratio CR.

Further, the closing timing IC of the intake valve 7 required for feeding the required amount of intake air into the combustion chamber 5 is stored as a function of the engine load L and engine speed Ne in the form of a map as shown in FIG. 14C in advance in the ROM 32. At step 104, this map is used for calculating the closing timing IC of the intake valve 7.

Next, at step 105, it is judged if the engine load L is smaller than a predetermined value L₃. Here, this predetermined value L₃ is, for example, made a value equal to the engine load at which when the engine load becomes smaller, the drop in the temperature of the exhaust gas may be accompanied with a drop in the temperature of the exhaust purification catalyst to below the activation temperature. When it is judged at step 105 that the engine load L is smaller than the predetermined value L₃, the routine proceeds to step 106. At step 106, the closing timing EC of the exhaust valve 9 is made substantially intake top dead center. Next, at step 107, the overlap period ΔOL is made the minimum and the routine proceeds to step 110.

On the other hand, when it is judged at step 105 that the engine load is the predetermined value L₃ or more, the routine proceeds to step 108. At step 108, the map shown in FIG. 15A is used to calculate the closing timing EC of the exhaust valve 9, next, at step 109, the map shown in FIG. 15B is used to calculate the overlap period ΔOL. That is, the closing timing EC of the exhaust valve 9 and overlap period ΔOL are stored as functions of the engine load L and engine speed Ne in the form of the maps shown in FIGS. 15A and 15B in advance in the ROM 32. These maps are used to calculate the closing timing EC of the exhaust valve 9 and overlap period ΔOL. After this, the routine proceeds to step 110.

At step 110, the mechanical compression ratio is made the mechanical compression ratio CR by controlling the variable compression ratio mechanism A, while the closing timing of the intake valve 7 is made the closing timing IC and the overlap period is made the overlap period ΔOL by controlling the intake variable valve timing mechanism B. Further, the closing timing of the exhaust valve 9 is made the closing timing EC by controlling the exhaust variable valve timing mechanism C.

While the invention has been described by reference to specific embodiments chosen for purposes of illustration, it should be apparent that numerous modifications could be made thereto by those skilled in the art without departing from the basic concept and scope of the invention.

Claims

1. A spark ignition type internal combustion engine comprising a variable compression ratio mechanism able to change a mechanical compression ratio, an intake variable valve timing mechanism able to change the closing timing of the intake valve, and an exhaust valve,

wherein at the time of engine low load operation, the variable compression ratio mechanism controls the mechanical compression ratio so as to be maximized to obtain a maximum expansion ratio and the intake variable valve timing mechanism controls the closing timing of the intake valve so that the actual compression ratio is set so that no knocking occurs,

wherein said maximum expansion ratio is 20 or more, and

wherein the closing timing of the exhaust valve at the time of engine low load operation is made substantially intake top dead center.

2. A spark ignition type internal combustion engine comprising a variable compression ratio mechanism able to change a mechanical compression ratio, an intake variable valve timing mechanism able to change the closing timing of the intake valve, and an exhaust variable valve timing mechanism able to change the closing timing of the exhaust valve,

wherein at the time of engine low load operation, the variable compression ratio mechanism controls the mechanical compression ratio so as to be maximized to obtain a maximum expansion ratio and the intake variable valve timing mechanism controls the closing timing of the intake valve so that the actual compression ratio is set so that no knocking occurs,

wherein said maximum expansion ratio is 20 or more, and

wherein a settable region of the closing timing of the exhaust valve at the time of engine low load operation is limited more to an intake top dead center side than that at the time of engine high load operation.

3. A spark ignition type internal combustion engine as set forth in claim 2, wherein at the time of engine low load operation, the closing timing of the exhaust valve is made substantially intake top dead center.

4. A spark ignition type internal combustion engine as set forth in claim 2, wherein the closing timing of the exhaust valve and the opening timing of the intake valve are controlled so that at the time of engine low load operation a period where the opening of the intake valve and the opening of the exhaust valve overlap is made minimum.

5. A spark ignition type internal combustion engine as set forth in claim 2, wherein the closing timing of the exhaust valve and the opening timing of the intake valve are controlled so that at the time of engine low load operation the period where the opening of the intake valve and the opening of the exhaust valve overlap becomes zero.

6. A spark ignition type internal combustion engine as set forth in claim 1, wherein at the time of engine low load operation, the opening timing of the intake valve is made substantially intake top dead center.

7. A spark ignition type internal combustion engine as set forth in claim 1, wherein the actual compression ratio at the time of engine low load operation is made substantially the same actual compression ratio as at the time of engine medium and high load operation.

8. A spark ignition type internal combustion engine as set forth in claim 7, wherein, at the time of engine low speed, regardless of the engine load, said actual compression ratio falls within a range of 9 to 11.

9. A spark ignition type internal combustion engine as set forth in claim 8, wherein the higher the engine speed, the higher the actual compression ratio.

10. (canceled)

11. A spark ignition type internal combustion engine as set forth in claim 1, wherein the amount of intake air fed into the combustion chamber is controlled by changing the closing timing of the intake valve.

12. A spark ignition type internal combustion engine as set forth in claim 11, wherein the closing timing of the intake valve is shifted as the engine load becomes lower to a direction away from intake bottom dead center until a limit closing timing enabling control of the amount of intake air fed into the combustion chamber.

13. A spark ignition type internal combustion engine as set forth in claim 12, wherein in a region of a load higher than the engine load when the closing timing of the intake valve reaches said limit closing timing, the amount of intake air fed into the combustion chamber is controlled without regard to a throttle valve arranged in an engine intake passage by changing the closing timing of the intake valve.

14. A spark ignition type internal combustion engine as set forth in claim 13, wherein in a region of a load higher than the engine load when the closing timing of the intake valve reaches said limit closing timing, the throttle valve is held at the fully opened state.

15. A spark ignition type internal combustion engine as set forth in claim 12, wherein in a region of a load lower than the engine load when the closing timing of the intake valve reaches said limit closing timing, a throttle valve arranged in an engine intake passage is used to control the amount of intake air fed into the combustion chamber.

16. A spark ignition type internal combustion engine as set forth in claim 12, wherein in a region of a load lower than the engine load when the closing timing of the intake valve reaches said limit closing timing, the lower the load, the larger the air-fuel ratio is made.

17. A spark ignition type internal combustion engine as set forth in claim 12, wherein in a region of a load lower than the engine load when the closing timing of the intake valve reaches said limit closing timing, the closing timing of the intake valve is held at said limit closing timing.

18. A spark ignition type internal combustion engine as set forth in claim 1, wherein said mechanical compression ratio is increased as the engine load becomes lower to the limit mechanical compression ratio.

19. A spark ignition type internal combustion engine as set forth in claim 18, wherein in a region of a load lower than the engine load when said mechanical compression ratio reaches said limit mechanical compression ratio, the mechanical compression ratio is held at said limit mechanical compression ratio.