BRPI0714220A2 - spark ignition type internal combustion engine - Google Patents

Abstract

INNER FUEL TYPE IGNITION ENGINE. The present invention relates to a spark ignition type internal combustion engine comprising a variable compression ratio mechanism capable of changing a mechanical compression ratio, a real compression action start sync shift mechanism capable of change a starting timing of an actual compression action and an exhaust valve. At low engine load operation, the mechanical compression ratio is maximized for maximum expansion ratio and the actual compression ratio is determined so that no blow occurs. The maximum expansion ratio is 20 or more. Large exhaust valve closing timing, exhaust purification catalyst temperature can be maintained at a relatively high temperature.

Description

Report of the Invention Patent for "INTERNAL COMBUSTION MOTOR TYPE MOTOR". TECHNICAL FIELD

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

A spark ignition type internal combustion engine is known in the art provided with a variable compression ratio mechanism capable of changing a mechanical compression ratio, and a variable valve timing mechanism capable of controlling synchronism. closing the inlet valve by performing a supercharging action by a turbocharger at the time of engine average load operation and high engine load operation and, in the state of retaining the actual compression ratio set at the time of operation. medium and high engine load, increasing the mechanical compression ratio and delaying the inlet valve closing timing when the engine load becomes smaller (for example, see Japanese Patent Publication (A) No. 2004- 218522).

However, in general, in an internal combustion engine, the higher the expansion ratio, the longer the period on an expansion stroke where the downward force acts on the piston, so the higher the expansion ratio, the more thermal efficiency is optimized. Therefore, to increase the thermal efficiency at engine running time, it is preferable to make the mechanical compression ratio as high as possible and to make the expansion ratio large.

However, if the expansion ratio is increased in this way, most of the thermal energy produced in the combustion chamber is converted to kinetic energy, so that the exhaust gas temperature drops. In addition, with this, the exhaust gas pressure in the combustion chamber at the end of the expansion stroke also becomes lower and consequently the exhaust gas exhaust from the combustion chamber becomes more difficult. This trend seems particularly noticeable when the expansion ratio is made 20 or more.

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

However, as explained above, if the expansion ratio is increased, the exhaust gas drops in temperature, so that the temperature by which the exhaust purification catalyst is raised by flow rate unit becomes lower. . Additionally, if the expansion ratio increases, the exhaust gas becomes more difficult to exhaust from the combustion chamber, so that the exhaust gas flow rate flowing within the exhaust purification catalyst becomes smaller. For this reason, operating the internal combustion engine in a state of large expansion ratio, keeping the exhaust purification catalyst at a high temperature becomes difficult. Therefore, an object of the present invention is to provide a motor

spark ignition type internal combustion engine capable of maintaining 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 an internal combustion engine of the

spark ignition type described in the claims of the claim section as a means to achieve the above objective.

In one aspect of the present invention there is provided a spark ignition type internal combustion engine comprising a variable compression ratio mechanism capable of changing a mechanical compression ratio, a variable starting synchronization mechanism. Actual compression action capable of changing a starting timing of an actual compression action and an exhaust valve, where at low engine load operation, the mechanical compression ratio is maximized to obtain a mechanical compression ratio. maximized to obtain a maximum expansion ratio and the actual compression ratio is determined so that no blow occurs, where the maximum expansion ratio is 20 or more, and where the exhaust valve close timing at Low engine load operation is done at the substantially upper inlet neutral.

In another aspect of the present invention, there is provided a spark ignition type internal combustion engine comprising a variable compression ratio mechanism capable of changing a mechanical compression ratio, a real compression action start sync shift mechanism. capable of changing a starting timing of a real compression action, and a variable exhaust valve timing mechanism capable of changing the closing timing of the exhaust valve, where at low engine operating time, the mechanical compression ratio is maximized to obtain a maximum expansion ratio and the actual compression ratio is determined so that no blow occurs, where the maximum expansion ratio is 20 or more, and where a determinable region of closing of the exhaust valve at low engine operating time is limited further to one side of the mor higher input than at high engine load operation.

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

In another aspect of the present invention, the engine further comprises an inlet variable valve timing mechanism capable of changing the inlet valve opening timing, and the exhaust valve closing timing and valve opening timing. The inlet ports are controlled so that at low engine load operation, a period where the inlet valve opening and the exhaust valve opening overlap is minimal.

In another aspect of the present invention, the engine further comprises an inlet variable valve timing mechanism capable of changing the inlet valve opening timing, and the exhaust valve closing timing and valve opening timing. The inlet ports are controlled so that at low engine operating time, the period when the inlet valve opening and the exhaust valve opening overlap becomes zero.

In another aspect of the present invention, the engine further comprises an inlet valve opening timing change mechanism capable of changing the inlet valve opening timing and, at low engine operating time, the timing of the engine. opening of the inlet valve becomes the substantially upper inlet neutral.

In another aspect of the present invention, the compression ratio

The actual compression ratio at low engine load operation is substantially the same as the actual compression ratio at medium and high engine load operation.

In another aspect of the present invention, at low engine speed, regardless of engine load, the actual compression ratio falls within a fixed 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 sync mechanism is comprised of an inlet variable valve sync mechanism capable of changing the inlet valve close sync.

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

In another aspect of the present invention, the inlet valve closure timing is shifted as the engine load becomes lower in a direction away from the lower inlet neutral to a limit close timing that allows controlling the amount of incoming air fed into the combustion chamber.

In another aspect of the present invention, in a region of a load greater than the engine load when the inlet valve shutoff timing reaches the limit shutoff timing, the amount of incoming air fed into the combustion chamber is controlled without consider a throttle valve arranged in an engine inlet port by changing the closing timing of the inlet valve.

In another aspect of the present invention, in a region of a load greater than the motor load when the inlet valve close sync reaches the limit close sync, the throttle valve is kept in the fully open state. In another aspect of the present invention, in a region of a

load less than the engine load when the inlet valve close sync reaches the close limit sync, a throttle valve arranged in a motor inlet port is used to control the amount of inlet air fed into the chamber of combustion.

In another aspect of the present invention, in a region of a load less than the engine load when the inlet valve close sync reaches the limit close sync, the smaller the load, the greater the air-fuel ratio becomes. In another aspect of the present invention, in a region of a

load less than the motor load when the inlet valve close sync reaches the limit close sync, the inlet valve close sync is maintained at the close limit sync.

In another aspect of the present invention, the compression ratio

Mechanical stress is increased when the motor load becomes lower for the mechanical limit compression ratio. In another aspect of the present invention, in a region of a load less 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, provided that such a gas of

The exhaust gas is discharged from the combustion chamber to the exhaust purification catalyst, even if the internal combustion engine is operated at a state of large expansion ratio, the exhaust purification catalyst can be kept at a relatively high temperature. BRIEF DESCRIPTION OF DRAWINGS

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

Figure 1 is an overview of a spark ignition type internal combustion engine. Figure 2 is a disassembled perspective view of a mechanical

variable compression ratio mechanism.

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

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

Figures 5A and 5B are views showing the inlet valve and exhaust valve lift amounts.

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

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

Figures 8A and 8B are views explaining a normal cycle and a super high expansion ratio cycle. Figure 9 is a view showing the change in relation of

mechanical compression, etc., according to the engine load.

Figures 10A, 10B and 10C are views showing changes in inlet valve and exhaust valve elevation.

Figure 11 is a view showing a region in which an exhaust valve closure synchrony according to the mechanical compression ratio can be determined.

Figures 12A and 12B are shown showing the changes in

lift inlet valve and exhaust valve.

Figure 13 is a flow chart for operational control.

Figures 14A, 14B and 14C are views showing the actual target compression ratio, etc. Figures 15A and 15B are views showing a map of the synchro-

exhaust valve closing mechanism, etc. BEST MODE FOR CARRYING OUT THE INVENTION

Figure 1 shows a side cross-sectional view of a spark ignition type internal combustion engine. Referring to Figure 1, 1 indicates a crankshaft housing,

2 a cylinder block, 3 a cylinder head, 4 a piston, 5 a combustion chamber, 6 a spark plug disposed in the upper center of the combustion chamber 5, 7 an inlet valve, 8 an inlet port, 9 an exhaust valve, and 10 an exhaust orifice. Inlet port 8 is connected through an inlet tube 11 in a clearing chamber 12, while inlet port 11 is provided with a fuel injector 13 to inject fuel into a corresponding inlet port. 8. Note that each fuel injector 13 may be arranged in each combustion chamber 5 instead of being attached to each inlet pipe 11.

Compensation chamber 12 is connected via an inlet duct 14 to a turbocharger compressor 15a outlet 15, while a compressor inlet 15a is connected via an inlet air quantity detector 16 using, for example, a heating wire in an air cleaner 17. The inlet duct 14 is provided inside it with a throttle valve 19 actuated by an actuator 18. On the other hand, the exhaust port 10 is connected through the exhaust pipe. 20 at the exhaust turbine inlet 15b of the exhaust turbocharger 15, while an exhaust turbine outlet 15b is connected via an exhaust pipe 21 into a catalytic converter 22 housing an exhaust purification catalyst. The exhaust pipe 21 has an air-fuel ratio sensor 23 disposed therein.

Additionally, in the embodiment shown in Figure 1, the connecting portion of the crankshaft housing 1 and the cylinder block 2 is provided with a variable compression ratio mechanism A capable of changing the relative positions of the crankshaft housing 1 and the cylinder block 2 in the axial direction of the cylinder to change the volume of the combustion chamber 5 when the piston 4 is positioned at the upper dead center of compression. In addition, it is provided with a variable valve timing mechanism B capable of individually controlling the closing timing of the inlet valve 7 to change the starting timing of the actual compression action, and capable of individually controlling the timing synchronization. opening of the inlet valve 7. In addition, it is provided with an exhaust variable valve timing mechanism C capable of individually controlling 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 to each other via a bidirectional bus 31 such as ROM (read memory) 32, RAM (random access memory) 33, CPU (microprocessor). 34), inlet port 35, and outlet port 36. The output signal from the inlet air quantity detector 16 and the output signal from the fuel-to-air ratio sensor 23 are input via corresponding AD converters 37 to input port 35. Additionally, the 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 41 is introduced via a corresponding AD converter 37 into the inlet port 35. Additionally, the inlet port 35 is connected to a crank angle sensor 42 generating an output pulse each time. z that the crankshaft rotates for, for example, 30 °. On the other hand, outlet port 36 is connected via drive circuit 38 to a spark plug 6, fuel injector 13, throttle valve actuator 18, variable compression ratio mechanism A, and mechanism Valve Timing Switch B.

Figure 2 is a disassembled perspective view of the variable compression ratio mechanism A shown in figure 1, while figures 3A and 3B are side cross-sectional views of the illustrated internal combustion engine. Referring to Figure 2, at the bottom of the two side walls of the cylinder block 2, a plurality of projecting parts 50 are formed separated from each other by a certain distance. Each projecting part 50 is formed with a circular cross-section meat insert hole 51. On the other hand, the top surface of the crankshaft housing 1 is formed with a plurality of projecting parts 52 separated from one another by a certain amount. distance and engaging between corresponding projecting parts 50. These projecting parts 52 are also formed with circular cross-section meat insertion holes 53. As shown in Figure 2, a pair of camshafts 54, 55 are

provided. Each of the meat shafts 54, 55 has circular meats 56 attached thereto capable of being rotatably inserted into the meat insertion holes 51 in each other position. These circular meats 56 are coaxial with the axes of rotation of the meat shafts 54, 55. On the other hand, between the circular meats 56, as shown by the dashes in figures 3A and 3B, extend eccentric axes 57 arranged eccentrically with respect to the axes. cam cams 54, 55. Each eccentric shaft 57 has other circular cams 58 rotatably attached thereto. As shown in Figure 2, these circular cams 58 are disposed between the circular cams 56. These circular cams 58 are rotatably inserted into the corresponding cam insertion holes 53.

When the circular shafts 56 attached to the cam shafts 54, 55 are rotated in opposite directions as shown by the solid line arrows in figure 3A of the state shown in figure 3A, the eccentric shafts 57 move to the lower center, thus the circular shafts 58 rotate in opposite directions of the circular cams 56 in the cam insert holes 53 as shown by the dashed line arrows in figure 3A. As shown in Figure 3B, when the eccentric axes 57 move to the lower center, the circular meat centers 58 move below the eccentric axes 57.

As will be understood from a comparison of FIG. 3A and FIG. 3B, the relative positions of the crank shaft housing 1 and the cylinder block 2 are determined by the distance between the circular meat centers 56 and the circular meat centers 58. The greater the distance between the circular meat centers 56 and the circular meat centers 58, the further away the cylinder block 2 is from the crankshaft housing 1. If the cylinder block 2 moves away from the gear housing crankshaft 1, the volume of the combustion chamber 5 when the piston 4 is positioned when the upper compression dead center increases, thus making the cam shafts 54, 55 rotate, the volume of the combustion chamber 5 when the piston 4 is positioned as top compression dead center can be changed.

As shown in Figure 2, to make cam shafts 54, 55 rotate in opposite directions, the shaft of a drive motor 59 is provided with a pair of helical gears 61, 62 with opposite thread directions. The gears 63, 64 engaging with these helical gears 61, 62 are secured at the 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 top dead center compression over a wide range. Note that the variable compression ratio mechanism A shown in Figures 1 to 3 shows an example. Any type of variable compression ratio mechanism can be used.

Additionally, Figure 4 shows a variable valve timing mechanism B provided on a cam shaft 70 to drive the inlet valve 7 in Figure 1. Referring to Figure 4, the variable valve timing mechanism B is comprised of a meat phase switch B1 fixed at one end of the meat axis 70 and changing the meat phase of the meat axis 70, and a meat actuation angle switch B2 disposed between the meat axis 70 and the hanger 24 of the inlet valve 7 and changing the working angle of the camshaft meat 70 to different working angles for the transmission to the inlet valve 7. Note that figure 4 is a sectional view side and plane view of cam actuation angle switch B2.

First, by explaining cam phase switch B1 of the variable valve timing mechanism B, this cam phase switch B1 is provided with a timing pulley 71 made to rotate by a motor crankshaft through a belt. in the direction of the arrow, a cylindrical housing 72 rotating together with the timing pulley 71, an axis 73 capable of rotating along a cam axis 70 and rotating with respect to the cylindrical housing 72, a plurality of partitions 74 are spaced apart. extending from an inner circumference of the cylindrical housing 72 to an outer circumference of the shaft 73, and vanes 75 extending between the divisions 74 of the outer circumference of the shaft 73 to the inner circumference of the cylindrical housing 72, the two sides of the vanes 75 formed with advanced use of hydraulic chambers 76 and delayed use of hydraulic chambers 77.

The working oil supply of the hydraulic chambers 76, 77 is controlled by a working oil supply control valve 78. This working oil supply control valve 78 is supplied 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 reel valve 85 to control the connection and disconnection of the ports. - 79, 80, 82, 83 and 84.

To advance the camshaft cam phase 70, reel valve 85 is made to move downward in FIG. 4, working oil fed from feed port 82 is fed through hydraulic port 79 to advanced use hydraulic chambers 76, and working oil for delayed use hydraulic chambers 77 are drained from drain hole 84. At this time, shaft 73 is rotated with respect to cylindrical housing 72 in the direction of the arrow. X.

As opposed to this, to slow the camshaft cam phase 70, in Fig. 4, the reel valve 85 moves upwards, the working oil fed from the feed port 82 and fed through the port. 80 for the hydraulic chambers 77 to retard, and the working oil in the hydraulic chambers 76 to advance is drained from the drainage hole 83. At this time, the shaft 73 is rotated with respect to the cylindrical housing 72 in the opposite direction to the arrows. X.

When the shaft 73 is rotated with respect to the cylindrical housing 72, if the reel valve 85 is returned to the neutral position shown in figure 4, the operation for relative rotation of the shaft 73 is terminated, and the shaft 73 is held in relative rotational position at this time. Therefore, cam phase switch B1 can be used to advance or delay the cam phase of cam shaft 70 by exactly the desired amount. That is, cam phase switch B1 can freely advance or slow the opening timing of inlet valve 7.

Next, explaining the cam actuation angle switch B2 of the variable valve timing mechanism B, this cam actuation angle switch B2 is provided with a control rod 90 arranged in parallel with the cam shaft 70 and made moving by an actuator 91 in the axial direction, an intermediate cam 94 engaging with a cam 92 of the cam shaft 70 and slidingly engaging with a groove 95 extending in a spiral formed on the control rod 90 and extending into the axial steering, and a pivoting cam 96 engaging with a valve hanger 24 to drive the inlet valve 7 and slidably engage with a groove 95 extending in a spiral formed on the control rod 90. The pivoting cam 96 is formed with a cam 97. When the meat shaft 70 rotates, the meat 92 causes the intermediate meat 94 to pivot at exactly a constant angle each time. This time, pivoting meat 96 is also oscillated at exactly a constant angle. On the other hand, the intermediate cam 94 and the pivoting cam 96 are non-movably supported in the axial direction of the control rod 90, so when the control rod 90 is moved by actuator 91 in the axial direction, the pivoting cam 96 it is rotated with respect to intermediate meat 94.

When the cam 92 of the camshaft 70 begins to engage with the intermediate cam 94 due to the relative rotational position relationship between the intermediate cam 94 and the pivoting cam 96, if the cam 97 of the pivoting cam 96 begins to engage with the cam. valve hanger 24, as shown by a in figure 5B, the opening time and elevation of the inlet valve 7 become maximum. When opposed to this, when actuator 91 is used to pivot cam 96 to rotate relative to intermediate cam 94 in the direction of arrow Y of FIG. 4, cam 92 of cam shaft 70 engages with intermediate cam 94, then thereafter. the cam 97 of the cam cam 96 engages with the valve hanger 24. In this case, as shown by b in figure 5 (B), the opening time and lift amount of the inlet valve 7 is makes it smaller than a.

When the pivoting cam 96 is rotated relative to the intermediate cam 94 in the direction of arrow Y of FIG. 4, as shown by c in FIG. 5B, the opening time period and elevation amount of the inlet valve 7 become further. smaller ones. That is, by using actuator 91 to change the relative rotational position of the intermediate cam 94 and pivoting cam 96, the opening time of the inlet valve 7 can be changed slightly. However, in this case, the lift amount of the inlet valve 7 becomes smaller the shorter the inlet valve 7 opens. Cam Phase Switch B1 can be used to change

freely opening timing of inlet valve 7 and cam actuation angle switch B2 can be used to freely change the timing of opening of inlet valve 7 in this way, so that the meat phase switch B1 and the meat angle switch B2, i.e. the variable input valve timing mechanism B, can be used to freely change the opening timing and opening time period of the input valve. 7, that is, the opening timing and closing timing of the inlet valve 7.

Note that the variable valve timing mechanism B shown in FIG. 1 and FIG. 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 FIG. 4.

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

In the following, the meaning of the terms used in the present application will be explained with reference to figures 6A to 6C. It is noted that Figures 6A, 6B and 6C show, for explanatory purposes, an engine with a combustion chamber volume of 50 ml and a piston stroke volume of 500 ml. In these figures 6A, 6B and 6C, the combustion chamber volume shows, the combustion chamber volume when the piston is in top dead center of compression. Figure 6A explains the mechanical compression ratio. The rela-

Mechanical compression ratio is a mechanically determined value from the piston stroke volume 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 Figure 6A, this mechanical compression ratio (50 ml + 500 ml) / 50 ml = 11.

Figure 6B explains the actual compression ratio. This actual compression ratio is a value determined from the actual stroke volume of the piston when the compression action is actually initiated 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 figure 6B, even if the piston begins to rise in the compression stroke, no compression action is performed while the inlet valve is open. The actual compression action is initiated after the inlet valve closes. Therefore, the actual compression ratio is expressed as follows using the actual stroke volume. In the example shown in figure 6B, the actual compression ratio becomes (50 ml + 450 ml) / 50 ml = 10.

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

In the following, the most basic aspects of the present invention will be explained with reference to figures 7, 8A and 8B. Note that Figure 7 shows the relationship between theoretical thermal efficiency and expansion ratio, while Figures 8A and 8B show a comparison between the normal cycle and the super high expansion ratio cycle, used selectively. according to the charge in the present invention. Figure 8A shows the normal cycle when the inlet valve

closes near the bottom dead center, and the piston compression action is initiated substantially near the bottom dead center. In the example shown in Fig. 8A also, just as the examples shown in Fig. 6A, 6B and 6C, the combustion chamber volume is 50 ml, and the piston stroke volume is 500 ml. As will be understood from Figure 8A, in a normal cycle the mechanical compression ratio is (50ml + 500ml) / 50ml = 11, the actual compression ratio is also about 11, and the expansion ratio is also becomes (50 ml + 500 ml) / 50 ml = 11. That is, in a common internal combustion engine, the mechanical compression ratio and the actual compression ratio and the expansion ratio become substantially equal.

The solid line in figure 7 shows the change in thermal efficiency.

theoretical in the case where 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 higher the expansion ratio, that is, the higher the actual compression ratio, the higher the theoretical thermal efficiency. Therefore, in a normal cycle, to increase the theoretical thermal efficiency, the actual compression ratio must become larger. However, due to restrictions on the occurrence of crash at the time of high engine load operation, the actual compression ratio can only be increased even to a maximum of about 12, consequently in a normal cycle. The theoretical thermal efficiency cannot be high enough.

On the other hand, under this situation, the inventors strictly differentiated between the mechanical compression ratio and the actual compression ratio and studied the theoretical thermal efficiency and as a result found that in the theoretical thermal efficiency, the expansion ratio is dominated. and the theoretical thermal efficiency is not greatly affected by the actual compression ratio. That is, if you increase the actual compression ratio, the explosive force goes up, but compression requires a lot of energy, so even if you increase the actual compression ratio, the theoretical thermal efficiency will not rise much. As opposed to this, if you increase the expansion ratio, how much

longer is the period during which a force acts by depressing the piston down at the moment the piston gives a rotational force to the crankshaft. Therefore, the larger the expansion ratio becomes, the greater the theoretical thermal efficiency becomes. The dashed line in figure 7 shows the theoretical thermal efficiency in the case of setting the actual compression ratio to 10 and increasing the expansion ratio in this state. Thus, it is found that the amount of theoretical thermal efficiency improvement when increasing the expansion ratio in the state where the actual compression ratio is kept low and the amount of theoretical thermal efficiency increase in the case where the Actual compression ratio is increased with the expansion ratio as shown in the solid line of figure 7 will not differ so much.

If the actual compression ratio is kept at a low value

In this way, no tapping will occur, so if you raise the expansion ratio in the state where the actual compression ratio is kept to a low value, tapping can be prevented and the theoretical thermal efficiency can be greatly increased. Figure 8B shows an example of the case when using variable compression ratio mechanism A and variable valve timing mechanism B to keep the actual compression ratio to a low value and raise the expansion ratio.

Referring to Figure 8B, in this example, the variable compression ratio mechanism A is used to decrease the combustion chamber volume from 50 ml to 20 ml. On the other hand, variable valve timing mechanism B is used to slow the inlet valve closing timing 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 figure 8 (A), 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 figure 8B, it is found that only the expansion ratio is high to 26. This is why the so-called "super high expansion ratio cycle" is called below.

As explained above, generally speaking, in an internal combustion engine, the lower the engine load, the worse the thermal efficiency, so to optimize the thermal efficiency at the time of vehicle operation, ie to optimize the fuel consumption. fuel efficiency, it is necessary to improve the thermal efficiency at the time of low engine load operation. On the other hand, in the super high expansion ratio cycle shown in figure 8B, the actual stroke volume of the piston at the moment of the compression stroke becomes smaller, thus the amount of inlet air that can be sucked into the chamber. of combustion 5 becomes smaller, so this super high expansion ratio cycle can only be employed when the engine load is relatively low. Therefore, in the present invention, at the moment of low engine load operation, the super high expansion ratio cycle shown in figure 8B is determined, while at the moment of high engine load operation, the ordinary cycle shown in figure 8A is determined. This is the basic aspect of the present invention.

Figure 9 shows overall operational control at constant operating time when engine speed is low. Below, the operational control as a whole will be explained with reference to figure 9.

Figure 9 shows the changes in the mechanical compression ratio, expansion ratio, valve close and inlet timing 7, actual compression ratio, the amount of air inlet, choke valve opening degree 17 , and loss of pumping, with the motor load. It is noted that in the present embodiment, to enable the three-way catalyst in catalytic converter 22 to simultaneously reduce unburned HC, CO and NOx in the exhaust gas, normally the average air-fuel ratio in the combustion chamber 5 is controlled by feedback on 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 high engine load operation, the normal cycle shown in figure 8A is executed. Therefore, as shown in figure 9, at this time, the mechanical compression ratio is low, so that the expansion ratio becomes low, and as shown by the solid line in figure 9, the inlet valve closing timing 7 is advanced. Additionally, at this time, the amount of inlet air is large. At this time, the opening of throttling valve 17 is kept fully open or substantially completely open, thus pumping loss becomes zero. On the other hand, as shown in figure 9, with the reduction in engine load, the mechanical compression ratio is increased, so the expansion ratio is also increased. Additionally, at this time, the actual compression ratio is kept substantially constant, as shown by the solid line in figure 9, slowing the closing timing of the inlet valve 7 when the motor load becomes lower. Note that at this time also, the throttle valve 17 is kept in the fully open or substantially fully open state, so the amount of inlet air fed into the combustion chamber 5 is controlled not by the throttle valve 17, but changing the closing timing of the inlet valve 7. At this time too, the pumping loss becomes zero.

Thus, when the engine load becomes smaller from the high load operating state of the engine, the mechanical compression ratio is increased with the drop in intake air quantity under a substantially constant actual compression ratio. That is, the volume of the combustion chamber 5 when the piston 4 reaches the upper compression dead center is reduced proportionally to the reduction in the amount of inlet air. Therefore, the volume of the combustion chamber 5 when the piston 4 reaches the upper compression dead center proportionally changes the amount of inlet air. It is noted that at this time, the air-fuel ratio in the combustion chamber 5 becomes the stoichiometric air-fuel ratio, thus the volume of the combustion chamber 5 when the piston 4 reaches the top compression dead center. proportionally changes the amount of fuel.

If the engine load becomes even lower, the mechanical compression ratio is further increased. When the mechanical compression ratio reaches the limit compression ratio forming the structural limit of the combustion chamber 5, in the region of a load less than the engine load Li, when the mechanical compression ratio reaches the limit mechanical compression ratio, the ratio Mechanical compression ratio is maintained at the limit motor compression ratio. Therefore, at low engine load operation, the mechanical compression ratio becomes maximum, and the expansion ratio also becomes maximum. Putting this another way, in the present invention, in order to obtain the maximum expansion ratio at the moment of low engine load operation, the mechanical compression ratio is made maximum. Additionally, at this time, the actual compression ratio is maintained at an actual compression ratio substantially the same as that at the time of medium and high engine load operation.

On the other hand, as shown by the solid line in figure 9, the inlet valve close sync 7 is delayed so that the limit close sync allows the control of the amount of incoming air fed into the combustion chamber 5 when the load is controlled. the engine becomes smaller. In the region of a load less than the L2 motor load when the inlet valve close sync 7 reaches the limit close sync, the inlet valve close sync 7 is maintained in the close sync. limit. If the inlet valve close sync 7 is maintained at the limit close sync, the amount of inlet air will no longer be able to be controlled by changing the inlet valve close sync 7. The amount of inlet air has to be controlled by some other method.

In the embodiment shown in Figure 9, at this time, that is, in the region of a load smaller than the L2 motor load when the inlet valve close sync 7 reaches the limit close sync, the throttle valve 17 is used to controlling the amount of incoming air fed to the combustion chamber 5. However, if you use the throttle valve 17 to control the amount of incoming air as shown in figure 9, the pumping loss increases.

Note that to prevent this loss of pumping, in the region of a load less than motor load L2 when the inlet valve close sync 7 reaches the limit close sync, the throttle valve 17 is maintained completely. open or substantial and completely open, and the air-fuel ratio may be higher, the lower the engine load. At this time, the fuel injector 13 is preferably disposed in the combustion chamber 5 to perform stratified combustion.

As shown in figure 9, at the moment of low speed of the

Regardless of the engine load, the actual compression ratio is kept substantially constant. The actual compression ratio at this time is made when the range of the actual compression ratio over the engine's high and medium load operating time ± 10%, preferably ± 5%. It is noted that in the embodiment according to the present invention, the actual compression ratio at the moment when the low engine speed is about 10 ± 1, ie from 9 to 11. However, if the speed is As the engine becomes larger, the air-fuel mixture in the combustion chamber 5 is distributed, so the beating becomes difficult, so in the embodiment according to the present invention, the higher the engine speed, the higher the fuel ratio. Real compression.

On the other hand, as explained above, in the super high expansion ratio cycle shown in figure 8B, the expansion ratio is 26. The higher this expansion ratio, the better, but if 20 or more, a considerably higher theoretical thermal efficiency can be obtained. Therefore, in the present invention, the variable compression ratio mechanism A is formed such that the expansion ratio becomes 20 or more.

Additionally, in the example shown in figure 9, the mechanical compression ratio is continuously changed according to the engine load. However, the mechanical compression ratio can also be changed in stages according to engine load.

On the other hand, as shown by the dashed line in figure 9, when the engine load becomes smaller by advancing the closing timing of the inlet valve 7 as well, it is possible to control the amount of inlet air without relying on the inlet valve. choke. Therefore, in Figure 9, if both cases shown by the solid line and the dashed line are comprehensively expressed, in the embodiment according to the present invention, the closing timing of the inlet valve 7 is displaced when the The engine load becomes smaller in a direction away from the lower BDC inlet dead center until the L2 limit close timing allows control of the amount of incoming air fed into the combustion chamber.

In the following, the closing timing of exhaust valve 9 will be explained by focusing on the low load operation where the superhigh expansion ratio cycle shown in figure 8B is performed.

In general, at the time of low load operation where a super high expansion ratio cycle is performed, the amount of heat generated due to combustion of the air-fuel mixture in the combustion chamber 5 is small, so that the temperature of the exhaust gas from the combustion chamber 5 becomes easily low. In addition to this, in an internal combustion engine, the higher the expansion ratio, the longer the period during which a downward pushing force on the piston acts at an expansion stroke, thus most of the energy The thermal energy produced by combustion of the air-fuel mixture in the combustion chamber is converted to the kinetic energy of the piston. As a result, the temperature of the flue gas in the combustion chamber at the end of the expansion course becomes lower. For this reason, when the super high expansion ratio cycle shown in figure 8B is executed at the time of an exhaust stroke, the exhaust gas temperature from the combustion chamber 5 to the exhaust pipe 20 becomes extremely low. This trend seems particularly noticeable when the expansion ratio is made 20 or more. Between running a super high expansion ratio cycle where the expansion ratio is 20 or more and a normal cycle where the expansion ratio is 12 or so, the combustion chamber exhaust gas temperature 5 differs by about 100 ° C.

On the other hand, in most internal combustion engines,

Harmful ingredients contained in the exhaust gas (eg HC, CO, NOx, etc.) are removed by providing within the engine exhaust port a three-way catalyst, NOx storage and reduction catalyst, or other exhaust purification. Such an exhaust purification catalyst cannot effectively remove harmful ingredients in the exhaust gas unless its temperature becomes the activation temperature or higher. Here, on most internal combustion engines, the exhaust gas temperature is considerably higher than the activation temperature, so the exhaust gas is flowed into the exhaust purification catalyst to maintain the exhaust purification catalyst temperature at activation temperature or higher.

However, if the super high expansion ratio cycle shown in figure 8B is executed, the exhaust gas exhaust temperature from the combustion chamber 5 to the exhaust pipe 20 will become slightly higher than the exhaust temperature. Activation, even if the exhaust gas flows within the exhaust purification catalyst, it becomes difficult to maintain the exhaust purification catalyst temperature at or above the activation temperature. Therefore, when the super high expansion ratio cycle is performed, to maintain the exhaust purification catalyst temperature at or above the activation temperature, it is necessary to make as much exhaust gas as possible flow into the exhaust purification catalyst.

Referring here to Figures 10A-10C, we will consider the relationship between the exhaust valve close timing 9 and the exhaust gas fluid rate from the combustion chamber 5 to the exhaust pipe 20. A Figure 10A shows the elevation changes of exhaust valve 9 and inlet valve 7 in the case where exhaust valve 9 is closed at the substantially inlet upper dead center, Figure 10B shows the same in the case where exhaust valve 9 is closed. Exhaust 9 is closed before the upper inlet neutral, while Figure 10C shows the same in the case where the exhaust valve 9 is closed after the upper inlet neutral.

As shown in Figure 10B, when exhaust valve 9 closes before the upper inlet neutral, the volume of the combustion chamber 5 when closing the exhaust valve 9 is greater than the volume of the combustion chamber when the piston is positioned at the upper inlet neutral (combustion chamber volume). After the exhaust valve 9 closes, the exhaust gas that corresponds 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 for the exhaust pipe 20 and the exhaust gas flow rate in. exhaust purification catalyst becomes small.

On the other hand, as shown in Fig. 10C, when exhaust valve 9 closes after the upper inlet neutral, exhaust valve 9 is opened even at the upper inlet neutral, so that when piston 4 reaches the upper inlet dead center, almost all of the exhaust gas in the combustion chamber 5 flows out of the exhaust port 10. However, if the exhaust valve 9 is opened even after the upper inlet neutral, part of the exhaust gas Exhaust that flows out once into the exhaust port 10 will again flow into the combustion chamber 5 as piston 4 descends.

In particular, when the super high expansion ratio cycle is executed, at the moment of the expansion stroke, the flue gas in the combustion chamber 5 expands considerably, so that the flue gas pressure at the end of the expansion stroke will be relatively low. For this reason, the resistance of the exhaust gas flowing from the combustion chamber 5 to the exhaust port 10 in the exhaust stroke will be poor. Therefore, if the piston 4 descends after reaching the upper inlet dead center, part of the exhaust gas flowing out of the exhaust port 10 will easily flow back into the combustion chamber 5.

In this way, when the exhaust valve 9 closes after the upper inlet neutral, the exhaust gas flowing out once into the exhaust port 10 will return back to the interior of the combustion chamber 5, thus Exhaust gas in the combustion chamber 5 will not be capable of being sufficiently exhausted in the exhaust pipe 20 and the exhaust gas flow rate flowing within the exhaust purification catalyst will be small.

Therefore, in the present embodiment, when the super high expansion ratio cycle shown in figure 8B is executed, that is, when the mechanical compression ratio is high, to prevent the exhaust valve 9 from being too tightly synchronized. Before or long after the upper inlet dead center, the region where the closing timing of the exhaust valve 9 can be determined is limited to the upper inlet dead side.

Figure 11 is a view showing a region in which the exhaust valve closure timing 9 according to the mechanical compression ratio can be established.

As shown in Figure 11, the region in which the exhaust valve 9 can be placed becomes the region between the maximum determinable advance amount and the maximum delay amount. As will be understood from the figure, the amount of advance by which the exhaust valve closing timing 9 can be set is smaller (after) the higher the mechanical compression ratio, while inversely the maximum amount. The delay time by which the closing timing of exhaust valve 9 can be established is shorter (earlier) the higher the mechanical compression ratio. For this reason, the region in which the exhaust valve closure timing 9 can be established becomes smaller the higher the mechanical compression ratio, that is, the narrower the greater the mechanical compression ratio. For example, as shown in Figure 11, when the mechanical compression ratio is low, the region at which the exhaust valve close timing 9 can be set is ATOCI, while when the mechanical compression ratio is high, the region in that the closing timing of exhaust valve 9 can be set is ΔΤ002 (ATOC2 <ATOC1).

Alternatively, when the super high expansion ratio cycle shown in figure 8B is executed, that is, when the mechanical compression ratio is high, to safely prevent exhaust valve 9 closing timing from being too high. advanced or very delayed from the upper inlet dead center, as shown in figure 10A, the exhaust valve 9 close timing can be done at the upper inlet dead center. Thus, when the mechanical compression ratio is

high, limiting the region where the exhaust valve 9 close timing can be set at the upper inlet side of the exhaust or the exhaust valve 9 close the substantially inlet upper neutral, It is possible to sufficiently exhaust the exhaust gas in the combustion chamber 5 for the exhaust pipe 20 and make the exhaust gas flow rate flowing within the purification large.

That is, exhaust valve 9 is made to close near the upper inlet neutral, as shown in Figure 10B, compared to closing the exhaust valve 9 before the upper inlet neutral, the volume of the exhaust chamber. 5 when the exhaust valve 9 is closed 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. Additionally, the exhaust valve 9 is made to close near the upper inlet neutral, as shown in figure 10C, compared to when closing the exhaust valve 9 after the upper inlet neutral, the amount of exhaust gas flowing into the combustion chamber 5 at the Exhaust gas flowing into the exhaust port 10 may be reduced. For this reason, as shown in Fig. 10A, when exhaust valve 9 is closing close to the upper inlet neutral, as shown in Figs. 10B and 10C, compared when exhaust valve 9 closes away from the upper inlet neutral. , the exhaust gas in the combustion chamber 5 may be sufficiently exhausted within the exhaust pipe and the exhaust gas flow rate flowing within the exhaust purification catalyst may be increased. As a result, even at low load operation where the super high expansion ratio cycle is performed, it is possible to keep the exhaust purification catalyst at or above the activation temperature.

Note that the "substantially upper input dead center" indicates within 10 ° before and after the upper input dead center, preferably within 5 ° before and after the upper input dead center.

In addition, if the mechanical compression ratio is increased, the volume of the combustion chamber in the upper inlet neutral becomes smaller and therefore depending on the closing timing of the exhaust valve 9, the exhaust valve 9 will eventually interfere with the piston. 4

Figures 10A to 10C show the piston interference line showing the limit where exhaust valve 9 or inlet valve 7 interferes with piston 4. When the lift curve of exhaust valve 9 interferes with the piston interference, exhaust valve 9 interferes with piston 4. Here, in Figure 10C, the elevation curve of exhaust valve 9 intersects with the piston interference line. This means that when exhaust valve 9 closes before the upper inlet dead center, while also depending on the delay length, exhaust valve 9 and piston 4 will eventually interfere.

When opposed to this, according to the present invention, when the mechanical compression ratio is high, the region in which the exhaust valve closing synchrony 9 can be set is limited to the upper dead side of in particular, the amount of maximum delay by which the closing timing of exhaust valve 9 may be set lower. For this reason, as shown in Figure 10A, even if the mechanical compression ratio becomes larger, the exhaust valve 9 may be prevented from interfering with the piston 4.

However, when there is a valve overlap where the inlet valve opening time period 7 and the exhaust valve opening time period 9 overlap, the amount of exhaust gas exhausted from the interior of the combustion chamber 5 for exhaust pipe 20 changes even during this period. Referring to Figures 12A and 12B below, consider the relationship between the overlap period where the inlet valve opening time period 7 and the exhaust valve opening time period 9 overlap and the amount of gas exhaust pipe from combustion chamber 5 to exhaust pipe 20. Figure 12A shows the case where the overlap period is zero, while Figure 12B shows the elevation changes of exhaust valve 9 and inlet valve 7 when the overlap period is large.

In general, when inlet 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 flowing from the combustion chamber 5 to the exhaust port 10 will sometimes flow into the inlet port 8. Thus, when part As the exhaust gas flows into the inlet port 8, the exhaust gas exhausted from the combustion chamber 5 to the exhaust pipe 20 will become smaller by this amount.

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

Therefore, in this embodiment, when the super high expansion ratio cycle shown in figure 8B is executed, that is, when the mechanical compression ratio is high, as shown in figure 12A, the closing sync of the exhaust valve 9 and the opening timing of the inlet valve 7 and are controlled to be minimal in the range in which the overlap period can be set. Therefore, for example, in an internal combustion engine where the determinable overlap period becomes 10 ° to 60 °, when the mechanical compression ratio is high, the overlap period is 10 °, while in an engine internal combustion where the determinable 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, minimizing the overlap period, the exhaust gas flowing into the inlet port 8 becomes smaller so that the exhaust chamber exhaust gas is exhausted. combustion 5 for the exhaust pipe becomes large, and consequently the exhaust gas flow rate flowing into the exhaust purification catalyst becomes larger.

It is noted that the overlap period when the ratio of

Mechanical compression is high, it need not necessarily be the least 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 determinable range even at a minimum.

Additionally, as explained above, if the mechanical compression ratio is increased, the combustion chamber volume in the upper inlet dead center becomes smaller. Consequently, depending on the opening timing of the inlet valve 7, the inlet valve 7 will eventually interfere with the piston 4.

Figures 12A and 12B show the piston interference line showing the limit where exhaust valve 9 or inlet valve 7 interferes with piston 4. If the lift curve of inlet valve 7 intersects the interference line Inlet piston 7 will interfere with piston 4. Here, in Figure 12B, the lift curve of inlet valve 7 intersects the piston interference line. This means that if the overlap period is increased, the inlet valve 7 and the piston 4 will eventually interfere with each other. That is, in the present embodiment, as explained above, the closing timing of exhaust valve 9 is at the substantially inlet upper dead center. The overlap period being large means that the opening timing of the input valve 7 is greatly advanced. If the opening timing of inlet valve 7 is advanced enormously, the inlet valve 7 and piston 4 will eventually interfere with each other.

When opposed to this, according to the present invention, when the mechanical compression ratio is high, the overlap period is minimal, so the opening timing of the inlet valve 7 is made substantially the upper inlet dead center or any less. For this reason, as shown in Figure 12A, even if the mechanical compression ratio becomes high, the inlet valve 7 may be prevented from interfering with the piston.

Figure · 13 shows an operating control control routine.

internal combustion engine of the spark ignition type of the present embodiment. Referring to Figure 13, first, in step 101, motor load L and motor speed Ne are sought. Next, in step 102, the map shown in Fig. 14A is used to calculate the actual target compression ratio. As shown in Fig. 14A, this actual target compression ratio becomes greater as motor speed Ne. Next, in step 103, the map shown in figure 14B is used to calculate the mechanical compression ratio CR. That is, the mechanical compression ratio CR required to make the actual compression ratio, the target actual compression ratio is stored as a function of engine load L and engine speed Ne in the form of a map as shown in figure 14B above. in ROM 32. This map is used to calculate the mechanical compression ratio CR.

Additionally, the inlet valve closure timing IC 7 required to feed the required amount of inlet air into the combustion chamber 5 is stored as a function of engine load L and engine speed Ne in the form of a map as shown in Figure 14C in ROM 32. In step 104, this map is used to calculate the closing synchronism IC of inlet valve 7.

Next, in step 105, it is judged whether the motor load L is less than a predetermined value L3. Here, this predetermined value L3 is, for example, made of a value equal to the engine load where when the engine load becomes lower, the drop in exhaust gas temperature may be accompanied by a drop in the catalyst temperature. exhaust purification below activation temperature. When it is judged at step 105 that the engine load L is less than the predetermined value L3, the routine proceeds to step 106. In step 106, the exhaust valve EC shut-off timing 9 is made substantially higher neutral. input. Next, at step 107, the overlap period AOL is minimal and the routine proceeds to step 110.

On the other hand, when it is judged at step 105 that the motor load is the predetermined value L3 or more, the routine proceeds to step

108. In step 108, the map shown in Fig. 15A is used to calculate the EC close timing of exhaust valve 9, then in step

109, the map shown in figure 15B is used to calculate the overlap period AOL. That is, the EC closing timing of the e-valve

9 and the AOL overlap period are stored as functions of engine load L and engine speed Ne in the form of the maps shown in figures 15A and 15B earlier in ROM 32. These maps are used to calculate the synchronism of EC closure of exhaust valve 9 and the overlap period AOL. After this, the routine proceeds to step 110.

In step 110, the mechanical compression ratio is made the mechanical compression ratio CR controlling the variable compression ratio mechanism A, while the inlet valve close timing 7 is made the IC close timing and the time period. overlap is made the overlap period AOL controlling the input variable valve timing mechanism B. In addition, the exhaust valve closing timing S is made the EC closing timing controlling the variable valve timing mechanism of exhaustion C.

While the invention has been described by reference to the specific embodiments chosen for illustration purposes, 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. Reference Listing

1 Carter 2 Cylinder Block 3 Cylinder Head 4 Piston Combustion Chamber 7 Inlet Valve 9 Discharge Valve A Variable Compression Rate Mechanism B Variable Inlet Valve Timing Mechanism C Variable Discharge Valve Synchronization Mechanism

Claims (19)

1. Spark ignition type internal combustion engine comprising a variable compression ratio mechanism capable of changing a mechanical compression ratio, a real compression action start sync shift mechanism capable of changing a compression ratio. starting from an actual compression action and an exhaust valve, where at low engine operating time, the mechanical compression ratio is maximized to obtain a maximum expansion ratio and the actual compression ratio is determined so that no blow occurs, where the maximum expansion ratio is 20 or more, and the exhaust valve closing timing at low engine load operation is made to substantially the upper inlet neutral. 2. Spark-ignition type internal combustion engine comprising a variable compression ratio mechanism capable of changing a mechanical compression ratio, a real compression action start sync shift mechanism capable of changing a compression ratio. starting from a real compression action, and a variable exhaust valve timing mechanism capable of changing the exhaust valve closing timing, where at low engine operating time, the compression ratio is maximized to obtain a maximum expansion ratio and the actual compression ratio is determined so that no blow occurs, where the maximum expansion ratio is 20 or more, and where a determinable region of exhaust valve at low engine load operation is limited more to an upper inlet neutral side than at operating high engine load will. Spark ignition type internal combustion engine according to claim 2, wherein at the time of low engine load operation, the exhaust valve closing timing is made to the substantially inlet upper dead center. Spark ignition type internal combustion engine according to claim 2, further comprising an inlet variable valve timing mechanism capable of changing the inlet timing of the inlet valve, wherein the inlet timing closing the exhaust valve and the timing of the inlet valve opening are controlled so that at the time of low engine load operation, a period where the inlet valve opening and the exhaust valve opening are overlap is minimal. Spark ignition type internal combustion engine according to claim 2, further comprising an inlet variable valve timing mechanism capable of changing the inlet valve opening timing, and in which Exhaust valve closing and inlet valve opening timing are controlled so that at low engine operating time, the period in which the inlet valve opening and the exhaust valve opening overlap, becomes zero. Spark ignition type internal combustion engine according to claim 1 or 2, further comprising an inlet valve opening timing mechanism capable of changing the inlet valve opening timing and, in the case of At the moment of low engine load operation, the opening timing of the inlet valve becomes the substantially upper inlet neutral. Spark ignition type internal combustion engine according to claim 1 or 2, wherein the actual compression ratio at the moment of low engine load operation is substantially the same as the actual compression ratio at the medium and high engine load operating moment. Spark ignition type internal combustion engine according to claim 7, wherein at the moment of low engine speed, regardless of engine load, said actual compression ratio falls within a range of 9 ° C. to 11 A spark ignition type internal combustion engine according to claim 8, wherein the higher the engine speed, the higher the actual compression ratio. Spark ignition type internal combustion engine according to claim 1 or 2, wherein said actual compression action starting timing shift mechanism is comprised of a variable valve timing mechanism capable of changing the closing timing of the inlet valve. Spark ignition type internal combustion engine according to claim 10, wherein the amount of inlet air fed into the combustion chamber is controlled by changing the closing synchrony of the inlet valve. Spark ignition type internal combustion engine according to claim 11, wherein the closing timing of the inlet valve is shifted when the engine load becomes lower in a direction away from neutral. lower inlet up to a limit closing synchrony that allows control of the amount of incoming air fed into the combustion chamber. A spark ignition type internal combustion engine according to claim 12, wherein in a region of a load greater than the engine load when said inlet valve close sync reaches the limit close sync. , the amount of inlet air fed into the combustion chamber is controlled without regard to a throttle valve disposed in an engine inlet passage by changing the close timing of the inlet valve. Spark ignition type internal combustion engine according to claim 13, wherein in a region of a load greater than the engine load when the inlet valve close sync reaches said limit close sync. , the throttle valve is kept in the fully open state. A spark ignition type internal combustion engine according to claim 12, wherein in a region of a load less than the engine load when the inlet valve close timing reaches said inlet timing. limit closure, a throttle valve arranged in an engine inlet port is used to control the amount of incoming air fed into the combustion chamber. A spark ignition type internal combustion engine according to claim 12, wherein in a region of a load less than the engine load when the inlet valve close timing reaches said inlet timing. limit closure, the lower the load, the higher the air-fuel ratio becomes. A spark ignition type internal combustion engine according to claim 12, wherein in a region of a load less than the engine load when the inlet valve closing timing reaches said supply timing. limit closing, the inlet valve closing timing is maintained at said limit closing timing. The spark ignition type internal combustion engine according to claim 2, wherein said mechanical compression ratio is increased when the engine load becomes lower to the limit mechanical compression ratio. A spark ignition type internal combustion engine according to claim 18, wherein in a region of a load less than the engine load when said mechanical compression ratio reaches said limit mechanical compression ratio, the mechanical compression ratio is maintained at said limit mechanical compression ratio.