Patent Abstract:
This internal combustion engine comprising a variable compression ratio mechanism either measures an exhaust temperature or exhaust pressure which varies according to an actual expansion ratio, or measures a physical quantity which varies according to the exhaust temperature and/or the exhaust pressure, and estimates the current mechanical compression ratio on the basis of the measured value.

Full Description:
CROSS-REFERENCE TO RELATED APPLICATION 
     This is a national phase application based on the PCT International Patent Application No. PCT/JP2012/064201 filed May 31, 2012, the entire contents of which are incorporated herein by reference. 
     TECHNICAL FIELD 
     The present invention relates to an internal combustion engine which is provided with a variable compression ratio mechanism. 
     BACKGROUND ART 
     Known in the art is an internal combustion engine which is provided with a variable compression ratio mechanism which can make a cylinder block move along a cylinder axis with respect to a crankcase so as to change the mechanical compression ratio. In general, the lower the engine load, the lower the thermal efficiency, so in such an internal combustion engine which is provided with a variable compression ratio mechanism, the mechanical compression ratio is made higher the lower the engine load so as to raise the expansion ratio and raise the thermal efficiency. 
     In this way, in an internal combustion engine which is provided with a variable compression ratio mechanism, target mechanical compression ratios are respectively set for the current engine operating states and the variable compression ratio mechanism is controlled so that the current target mechanical compression ratios are realized. However, in actuality, sometimes the current target mechanical compression ratio is not realized. If the target mechanical compression ratio is not realized, the current desired expansion ratio is also not realized. 
     In this way, it is desirable to estimate the current actual mechanical compression ratio. For example, the cylinder pressure at the time of combustion is affected by the amount of fed fuel, so it has been proposed to use the cylinder pressure at top dead center during a fuel cut operation as the basis to estimate the current actual compression ratio (see PLT 1). 
     CITATIONS LIST 
     Patent Literature 
     PLT 1: Japanese Patent Publication No. 2010-174757A 
     PLT 2: Japanese Patent Publication No. 2006-046193A 
     PLT 3: International Publication WO2010/073411 
     PLT 4: International Publication WO2010/125694 
     PLT 5: Japanese Patent Publication No. 2010-024977A 
     SUMMARY OF INVENTION 
     Technical Problem 
     If the current actual compression ratio is estimated in the above way, the intake valve closing timing can be used as the basis to estimate the current actual mechanical compression ratio. However, unless a fuel cut operation is performed during operation at the current mechanical compression ratio, it is not possible to estimate the current actual compression ratio and due to this it is not possible to estimate the current actual mechanical compression ratio. 
     Therefore, an object of the present invention is to provide an internal combustion engine which is provided with a variable compression ratio mechanism which can estimate the current actual mechanical compression ratio when a fuel cut operation is not performed. 
     Solution to Problem 
     An internal combustion engine which is provided with a variable compression ratio mechanism according to claim  1  of the present invention measures an exhaust temperature or an exhaust pressure which changes in accordance with an actual expansion ratio or measures a physical quantity which changes in accordance with at least one of the exhaust temperature and the exhaust pressure, uses the measured measurement value as the basis to estimate a current mechanical compression ratio, and 
     the measurement value is a supercharging pressure at a downstream side of a compressor of a turbocharger. 
     An internal combustion engine which is provided with a variable compression ratio mechanism according to claim  3  of the present invention provides the internal combustion engine which is provided with a variable compression ratio mechanism according to claim  1  wherein the engine is provided with a detection device which directly or indirectly detects an actuating amount of an actuator of the variable compression ratio mechanism, the actuator is controlled so that the actuating amount which is detected by the detection device becomes the actuating amount which corresponds to a target mechanical compression ratio, and the actuating amount which is detected by the detecting device is corrected by a difference between the actuating amount which corresponds to a mechanical compression ratio which is estimated based on the measurement value in a specific engine operating state and the actuating amount which corresponds to a target mechanical compression ratio of the specific engine operating state. 
     An internal combustion engine which is provided with a variable compression ratio mechanism according to claim  4  of the present invention provides the internal combustion engine which is provided with a variable compression ratio mechanism according to claim  3  wherein the specific engine operating state is an engine operating state in which the target mechanical compression ratio becomes a set mechanical compression ratio or less. 
     An internal combustion engine which is provided with a variable compression ratio mechanism according to claim  5  of the present invention provides the internal combustion engine which is provided with a variable compression ratio mechanism according to claim  1  wherein a wastegate passage which bypasses a turbine of a turbocharger is provided, a wastegate valve which controls an amount of exhaust which passes through the wastegate passage is arranged in the wastegate passage, and a difference between a supercharging pressure at a downstream side of a compressor of the turbocharger which is measured when making the wastegate valve a first opening degree and the supercharging pressure of a compressor of the turbocharger which is measured when making the wastegate valve a second opening degree is made the measurement value. 
     An internal combustion engine which is provided with a variable compression ratio mechanism according to claim  6  of the present invention provides the internal combustion engine which is provided with a variable compression ratio mechanism according to claim  1  or  5  wherein when a compressor of a supercharger is arranged at an upstream side of the compressor of the turbocharger, the supercharging pressure is made a differential pressure before and after the compressor of the turbocharger. 
     An internal combustion engine which is provided with a variable compression ratio mechanism according to claim  7  of the present invention provides the internal combustion engine which is provided with a variable compression ratio mechanism according to any one of claims  1  and  3  to  6  wherein the engine estimates the current actual compression ratio and uses the estimated current actual compression ratio and the estimated current mechanical compression ratio as the basis to estimate the current closing timing of the intake valve. 
     Advantageous Effects of Invention 
     According to the internal combustion engine which is provided with a variable compression ratio mechanism according to claim  1  of the present invention, the engine measures an exhaust temperature or an exhaust pressure which changes in accordance with an actual expansion ratio or measures a physical quantity which changes in accordance with at least one of the exhaust temperature and the exhaust pressure and uses the measured measurement value as the basis to estimate a current mechanical compression ratio. Due to this, it is possible to estimate the current actual mechanical compression ratio when a fuel cut operation is not being performed. 
     According to the internal combustion engine which is provided with a variable compression ratio mechanism according to claim  1  of the present invention, in the internal combustion engine which is provided with a variable compression ratio mechanism according to claim  1 , the measurement value is a supercharging pressure at a downstream side of a compressor of a turbocharger. A generally provided supercharging pressure sensor can be used to measure the measurement value for estimating the mechanical compression ratio. 
     According to the internal combustion engine which is provided with a variable compression ratio mechanism according to claim  3  of the present invention, in the internal combustion engine which is provided with a variable compression ratio mechanism according to claim  1 , the engine is provided with a detection device which directly or indirectly detects an actuating amount of an actuator of the variable compression ratio mechanism, the actuator is controlled so that the actuating amount which is detected by the detection device becomes the actuating amount which corresponds to a target mechanical compression ratio, and the actuating amount which is detected by the detecting device is corrected by a difference between the actuating amount which corresponds to a mechanical compression ratio which is estimated based on the measurement value in a specific engine operating state and the actuating amount which corresponds to a target mechanical compression ratio of the specific engine operating state. Due to this, by such control of the actuator based on the corrected actuating amount, it is possible to realize the target mechanical compression ratio even in an engine operation other than a specific engine operating state. 
     According to the internal combustion engine which is provided with a variable compression ratio mechanism according to claim  4  of the present invention, in the internal combustion engine which is provided with a variable compression ratio mechanism according to claim  3 , the specific engine operating state is an engine operating state in which the target mechanical compression ratio becomes a set mechanical compression ratio or less. The exhaust temperature or exhaust pressure changes relatively largely in response to a slight deviation in the mechanical compression ratio when the target mechanical compression ratio is not realized, so a slight deviation of the mechanical compression ratio can be accurately detected and the actuating amount of which detected by the detection device can be accurately corrected. 
     According to the internal combustion engine which is provided with a variable compression ratio mechanism according to claim  5  of the present invention, in the internal combustion engine which is provided with a variable compression ratio mechanism according to claim  1 , a wastegate passage which bypasses a turbine of a turbocharger is provided, a wastegate valve which controls an amount of exhaust which passes through the wastegate passage is arranged in the wastegate passage, and a difference between a supercharging pressure at a downstream side of a compressor of the turbocharger which is measured when making the wastegate valve a first opening degree and the supercharging pressure at a downstream side of a compressor of the turbocharger which is measured when making the wastegate valve a second opening degree is made the measurement value. Due to this, it is possible to eliminate the deviation, from the measurement value, in the supercharging pressure which occurs due to individual differences in turbochargers and possible to estimate a more accurate mechanical compression ratio. 
     According to the internal combustion engine which is provided with a variable compression ratio mechanism according to claim  6  of the present invention, in the internal combustion engine which is provided with a variable compression ratio mechanism according to claim  1  or  5 , when a compressor of a supercharger is arranged at an upstream side of the compressor of the turbocharger, the supercharging pressure is made a differential pressure before and after the compressor of the turbocharger. Due to this, it is possible to eliminate the effect of supercharging by the compressor of the supercharger and possible to estimate a more accurate mechanical compression ratio. 
     According to the internal combustion engine which is provided with a variable compression ratio mechanism according to claim  7  of the present invention, in the internal combustion engine which is provided with a variable compression ratio mechanism according to any of claims  1  and  3  to  6 , the engine estimates the current actual compression ratio and uses the estimated current actual compression ratio and the estimated current mechanical compression ratio as the basis to estimate the current closing timing of the intake valve. It becomes possible to estimate an accurate closing timing of the intake valve. 
    
    
     
       BRIEF DESCRIPTION OF DRAWINGS 
         FIG. 1  is an overview of an internal combustion engine. 
         FIG. 2  is a disassembled perspective view of a variable compression ratio mechanism. 
         FIGS. 3(A) to 3(C)  are side cross-sectional views of an internal combustion engine which is shown schematically. 
         FIG. 4  is a view which shows a variable valve timing mechanism. 
         FIG. 5  is a view which shows lift amounts of an intake valve and an exhaust valve. 
         FIGS. 6(A) to 6(C)  are views for explaining an mechanical compression ratio, actual compression ratio, and expansion ratio. 
         FIG. 7  is a view which shows a relationship between a theoretical thermal efficiency and an expansion ratio. 
         FIGS. 8(A) and 8(B)  are views for explaining a normal cycle and superhigh expansion ratio cycle. 
         FIG. 9  is a view which shows changes in a mechanical compression ratio etc. in accordance with an engine load. 
         FIG. 10  is a first flow chart for estimating an actual mechanical compression ratio. 
         FIG. 11  is a map for setting a correction amount of a mechanical compression ratio which is used in a first flow chart. 
         FIG. 12  is a second flow chart for estimating an actual mechanical compression ratio. 
         FIG. 13  is a map for setting a correction amount of a mechanical compression ratio which is used in a second flow chart. 
         FIG. 14  is a schematic overview of an internal combustion engine in the case where a turbocharger is arranged. 
         FIG. 15  is a third flow chart for estimating an actual mechanical compression ratio. 
         FIG. 16  is a map for setting a correction amount of a mechanical compression ratio which is used in a third flow chart. 
         FIG. 17  is a fourth flow chart for setting a supercharging pressure true value. 
         FIG. 18  is a graph which shows a relationship between an opening degree of a wastegate valve and a supercharging pressure. 
     
    
    
     DESCRIPTION OF EMBODIMENTS 
       FIG. 1  shows a side cross-sectional view of an internal combustion engine which is provided with a variable compression ratio mechanism according to the present invention. Referring to  FIG. 1, 1  indicates a crankcase,  2  a cylinder block,  3  a cylinder head,  4  a piston,  5  a combustion chamber,  6  a spark plug which is arranged at the top center of a combustion chamber  5 ,  7  an intake valve,  8  an intake port,  9  an exhaust valve, and  10  an exhaust port. Each intake port  8  is connected through an intake runner  11  to a surge tank  12 . At each intake runner  11 , a fuel injector  13  is arranged for injecting fuel toward the inside of the respectively corresponding intake port  8 . Note that, each fuel injector  13  may be arranged in a combustion chamber  5  instead of attached to an intake runner  11 . 
     The surge tank  12  is connected with an air cleaner  15  through an intake duct  14 . Inside the intake duct  14 , a throttle valve  17  which is driven by an actuator  16  and an intake air detector  18  which uses for example hot wires are arranged. The exhaust port  10  is connected through an exhaust manifold  19  to a catalyst device  20  which has for example a three-way catalyst built into it. Inside the exhaust manifold  19 , an air-fuel ratio sensor  21  is arranged. Further, inside the exhaust manifold  19 , a temperature sensor  28  for measuring the exhaust temperature and a pressure sensor  29  for measuring the exhaust pressure are arranged. 
     On the other hand, in the embodiment which is shown in  FIG. 1 , at the connecting part of the crankcase  1  and the cylinder block  2 , a variable compression ratio mechanism A is provided which can change the relative positions of the crankcase  1  and the cylinder block  2  in the cylinder axial direction and thereby change the volume of the combustion chamber  5  when the piston  4  is positioned at compression top dead center. Furthermore, an actual compression action start timing changing mechanism B which is able to change the start timing of the actual compression action is provided. Note that, in the embodiment which is shown in  FIG. 1 , this actual compression action start timing changing mechanism B is comprised of a variable valve timing mechanism which can control the closing timing of the intake valve  7 . 
     As shown in  FIG. 1 , the crankcase  1  and the cylinder block  2  have a relative position sensor  22  attached to them for detecting the relative positional relationship between the crankcase  1  and the cylinder block  2 . From this relative position sensor  22 , an output signal which shows the change in distance between the crankcase  1  and the cylinder block  2  is output. Further, the variable valve timing mechanism B has a valve timing sensor  23  which generates an output signal which shows a closing timing of the intake valve  7  attached to it. The actuator  16  for driving the throttle valve has a throttle opening degree sensor  24  which generates an output signal which shows the throttle valve opening degree attached to it. 
     An electronic control unit  30  is comprised of a digital computer. This is provided with components connected to 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 . Output signals of the intake air detector  18 , the air-fuel ratio sensor  21 , relative position sensor  22 , valve timing sensor  23 , throttle opening degree sensor  24 , later explained cam rotational angle sensor  25 , temperature sensor  28 , and pressure sensor  29  are input through respectively corresponding AD converters  37  to an input port  35 . Further, the accelerator pedal  40  is connected to a load sensor  41  which generates an output voltage which is proportional to the amount of depression L 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 . Furthermore, the input port  35  is connected to a crank angle sensor  42  which generates an output pulse each time a crankshaft rotates by for example 30°. On the other hand, the output port  36  is connected through a corresponding drive circuit  38  to each spark plug  6 , fuel injector  13 , throttle valve drive-use actuator  16 , variable compression ratio mechanism A, and variable valve timing mechanism B. 
       FIG. 2  is a disassembled perspective view of the variable compression ratio mechanism A which is shown in  FIG. 1 , while  FIGS. 3(A) to 3(C)  are side cross-sectional views of the illustrated internal combustion engine. Referring to  FIG. 2 , at the bottom of the two sides of the cylinder block  2 , a plurality of projecting parts  50  are formed spaced apart from each other. That is, cylinder block side supports are formed. Inside these projecting parts  50 , cam insertion holes  51  of round cross-sectional shapes are formed. On the other hand, at the top surface of the crankcase  1 , a plurality of projecting parts  52  are formed which are spaced apart from each other and fit between the corresponding projecting parts  50 . That is, crankcase side supports are formed. Inside these projecting parts  52  as well, round cross-section cam insertion holes  53  are formed. 
     As shown in  FIG. 2 , a pair of cam shafts  54  and  55  are provided. On these cam shafts  54  and  55 , at every other position, a concentric part  58  which is inserted rotably into a cam insertion hole  53  is positioned. These concentric parts  58  are coaxial with the axes of rotation of the cam shafts  54 ,  55 . On the other hand, at the two sides of each concentric part  58 , as shown in  FIGS. 3(A) to 3(C) , eccentric parts  57  which are arranged eccentrically with respect to the axes of rotation of the cam shafts  54 ,  55  are positioned. Other circular cams  56  are attached eccentrically on these eccentric parts  57  in a rotatable manner. That is, the eccentric parts  57  engage with the eccentric holes which are formed in the circular cams  56  and the circular cams  56  pivot about the eccentric parts  57  centered about the eccentric holes. As shown in  FIG. 2 , these circular cams  56  are arranged at the two sides of each concentric part  58 . These circular cams  56  are rotatably inserted into the corresponding cam insertion holes  51 . Further, as shown in  FIG. 2 , the cam shaft  55  has a cam rotational angle sensor  25  which generates an output signal which expresses the rotational angle of the cam shaft  55  attached to it. 
     If making the concentric parts  58  of the cam shafts  54  and  55  rotate from the state such as shown in  FIG. 3(A)  in the mutually opposite directions as shown by the solid arrows in  FIG. 3(A) , the eccentric parts  57  move in directions away from each other, so the circular cams  56  rotate in opposite directions from the concentric parts  58  in the cam insertion holes  51  and, as shown in  FIG. 3(B) , the positions of the eccentric parts  57  change from high positions to intermediate height positions. If next making the concentric parts  58  further rotate in the direction shown by the arrow, as shown by  FIG. 3(C) , the eccentric parts  57  become the lowest position. 
     Note that,  FIG. 3(A) ,  FIG. 3(B) , and  FIG. 3(C)  show the positional relations between the center “a” of the concentric part  58  and the center “b” of the eccentric part  57  at their respective states. 
     As will be understood from a comparison of  FIG. 3(A)  to  FIG. 3(C) , the relative positions of the crankcase  1  and the cylinder block  2  are determined by the distance between the center “a” of the concentric part  58  and the center “c” of the circular cam  56 . The larger the distance between the center “a” of the concentric part  58  and the center “c” of the circular cam  56  is made, the further the cylinder block  2  is away from the crankcase  1 . That is, the variable compression ratio mechanism A uses a crank mechanism using a rotating cam so as to make the relative positions between the crankcase  1  and cylinder block  2  change. If the cylinder block  2  moves away from the crankcase  1 , the volume of the combustion chamber  5  when the piston  4  is positioned at compression top dead center increases. Therefore, by rotating the cam shafts  54  and  55 , 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  and  55  rotate in opposite directions, the shaft of a drive motor  59  is formed with a pair of worms  61  and  62  with opposite spiral directions. The worm gears  63  and  64  which engage with these worms  61  and  62  are fastened to the ends of the cam shafts  54  and  55 . In this embodiment, by operating the drive motor  59 , the volume of the combustion chamber  5  when the piston  4  is positioned at compression top dead center can be changed over a wide range. 
     On the other hand,  FIG. 4  shows a variable valve timing mechanism B which is attached to the end part of the cam shaft  70  for driving the intake valve  7  at  FIG. 1 . If referring to  FIG. 4 , this variable valve timing mechanism B is provided with a timing pulley  71  which is made to rotate in the arrow direction by the engine crankshaft through the timing belt, a cylindrical housing  72  which rotates together with the timing pulley  71 , a rotary shaft  73  which can rotate together with the intake valve drive-use cam shaft  70  and which can rotate with respect to the cylindrical housing  72 , a plurality of partition walls  74  which extend from the inside circumferential wall of the cylindrical housing  72  to the outside wall circumference of the rotary shaft  73 , and vanes  75  which extend from the outside circumferential surface of the rotary shaft  73  to the inside circumferential surface of the cylindrical housing  72  between the partition walls  74 . At the two sides of each vane  75 , an advance-use hydraulic chamber  76  and a delay-use hydraulic chamber  77  are formed. 
     The feed of working oil to the hydraulic chambers  76  and  77  is controlled by a working oil feed control valve  78 . This working oil feed control valve  78  is provided with hydraulic ports  79  and  80  which are respectively connected to the hydraulic chambers  76  and  77 , a feed port  82  of working oil which is discharged from the hydraulic pump  81 , a pair of drain ports  83  and  84 , and a spool valve  85  which controls the connections and disconnections of the ports  79 ,  80 ,  82 ,  83 , and  84 . 
     When advancing the phase of the cam of the intake valve drive-use cam shaft  70 , the spool valve  85  is made to move to the right in  FIG. 4 , working oil which is supplied from the feed port  82  is supplied to the advance-use hydraulic chamber  76  through a hydraulic port  79 , and working oil in the delay-use hydraulic chamber  76  is exhausted from the drain port  84 . At this time, the rotary shaft  73  is made to rotate relative to the cylindrical housing  72  in the arrow direction. 
     As opposed to this, when delaying the phase of the cam of the intake valve drive-use cam shaft  70 , the spool valve  85  is made to move to the left in  FIG. 4 , working oil which is supplied from the feed port  82  is supplied to the delay-use hydraulic chamber  77  through the hydraulic port  80 , and working oil in the advance-use hydraulic chamber  76  is exhausted from the drain port  83 . At this time, the rotary shaft  73  is made to rotate relative to the cylindrical housing  72  in the opposite direction to the arrow mark. 
     When the rotary shaft  73  is made to rotate relative to the cylindrical housing  72 , if the spool valve  85  is returned to the neutral position which is shown in  FIG. 4 , the relative rotation operation of the rotary shaft  73  is made to stop and the rotary shaft  73  is held at the relative rotation position at that time. Therefore, the variable valve timing mechanism B can be used to make the phase of the cam of the intake valve drive-use cam shaft  70  advance or can be used to make it delayed by exactly a desired amount. 
     In  FIG. 5 , the solid line shows when the variable valve timing mechanism B is used to make the phase of the cam of the intake valve drive-use cam shaft  70  the most advanced, while the broken line shows when it is used to make the phase of the cam of the intake valve drive-use cam shaft  70  the most delayed. Therefore, the opening time period of the intake valve  7  can be freely set between the range which is shown by the solid line and the range which is shown by the broken line in  FIG. 5 . Therefore, the closing timing of the intake valve  7  can also be set to any crank angle in the range which is shown by the arrow C in  FIG. 5 . 
     The variable valve timing mechanism B which is shown in  FIG. 1  and  FIG. 4  shows one example. For example, it is possible to use a variable valve timing mechanism which enables only the closing timing of the intake valve to be changed while maintaining the opening timing of the intake valve constant or various other types of variable valve timing mechanisms. 
     Next, referring to  FIGS. 6(A) to 6(C) , the meaning of the terms which are used in the present application will be explained. Note that,  FIGS. 6(A), 6(B) , and  6 (C) show engines with combustion chamber volumes of 50 ml and piston stroke volumes of 500 ml for explanatory purposes. In these  FIGS. 6(A), 6(B) , and  6 (C), the “combustion chamber volume” expresses the volume of a combustion chamber when the piston is positioned at top dead center of the compression stroke. 
       FIG. 6(A)  explains the mechanical compression ratio. The mechanical compression ratio is a value which is mechanically determined from only the piston stroke volume at the time of the compression stroke and the combustion chamber volume. This mechanical compression ratio is expressed by (combustion chamber volume+stroke volume)/combustion chamber volume. In the example which is shown in  FIG. 6(A) , this mechanical compression ratio becomes (50 ml+500 ml)/50 ml=11. 
       FIG. 6(B)  explains the actual compression ratio. This actual compression ratio is a value which is determined from the actual piston stroke volume from when the compression action actually is 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. 6(B) , even if the piston starts to rise in the compression stroke, no compression action is performed while the intake valve is open. The actual compression action is started after the intake valve closes. Therefore, the actual compression ratio is expressed as shown above using the actual stroke volume. In the example which is shown in  FIG. 6(B) , the actual compression ratio becomes (50 ml+450 ml)/50 ml=10. 
       FIG. 6(C)  explains the expansion ratio. The expansion ratio is a value which is determined from the stroke volume of the piston and the combustion chamber volume at the time of the expansion stroke. This expansion ratio is expressed by the (combustion chamber volume +stroke volume)/combustion chamber volume. In the example which is shown in  FIG. 6(C) , this expansion ratio becomes (50 ml+500 ml)/50 ml=11. 
     Next, while referring to  FIG. 7  and  FIGS. 8(A) and 8(B) , the super expansion ratio cycle which is used in the present invention will be explained. Note that,  FIG. 7  shows the relationship between the theoretical thermal efficiency and the expansion ratio, while  FIGS. 8(A) and 8(B)  show a comparison between a normal cycle and a superhigh expansion ratio cycle which are selectively used in accordance with the load in the present invention. 
       FIG. 8(A)  shows the normal cycle in the case where the intake valve closes near bottom dead center and the compression action by the piston is started from near substantially bottom dead center of the intake stroke. In the example which is shown in  FIG. 8(A) , in the same way as the examples which are shown in  FIGS. 6(A) , (B), (C), the combustion chamber volume is made 50 ml and the piston stroke volume is made 500 ml. As will be understood from  FIG. 8(A) , in the 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 a normal internal combustion engine, the mechanical compression ratio, the 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 the expansion ratio are substantially equal, that is, at the time of a normal cycle. In this case, it is learned that the larger the expansion ratio becomes, that is, the higher the actual compression ratio becomes, the higher the theoretical thermal efficiency becomes. Therefore, in the normal cycle, to raise the theoretical thermal efficiency, it is sufficient to raise the actual compression ratio. However, due to restrictions on the occurrence of knocking at the time of engine high load operation, the actual compression ratio can only be raised up to about 12 even at the maximum. Therefore, in a normal cycle, the theoretical thermal efficiency cannot be made sufficiently high. 
     On the other hand, in view of this situation, studies have been conducted to strictly separate the mechanical compression ratio and the actual compression ratio while raising the theoretical thermal efficiency. As a result, it was discovered that in the theoretical thermal efficiency, the expansion ratio is dominant and that the actual compression ratio does not have almost any effect on the theoretical thermal efficiency. That is, if raising the actual compression ratio, the explosive force rises, but a large energy is required for compression. Therefore, even if raising the actual compression ratio, the theoretical thermal efficiency does not become much larger at all. 
     As opposed to this, if increasing the expansion ratio, the time period during which a pushdown force acts on the piston at the time of the expansion stroke becomes longer and therefore the time period during which the piston gives a rotational force to the crankshaft becomes longer. Therefore, the greater the expansion ratio is made, the more the theoretical thermal efficiency rises. The broken line ε=10 of  FIG. 7  shows the theoretical thermal efficiency when raising the expansion ratio in the state setting the actual compression ratio at 10. In this way, it is learned that there is no large difference between the amount of rise of the theoretical thermal efficiency when raising the expansion ratio in the state maintaining the actual compression ratio ε at a low value and the amount of rise of the theoretical thermal efficiency when making the actual compression ratio increase together with the expansion ratio as shown by the solid line of  FIG. 7 . 
     If the actual compression ratio is maintained at a low value in this way, knocking will never occur. Therefore, if raising the expansion ratio in a state maintaining the actual compression ratio at a low value, it is possible to prevent the occurrence of knocking while greatly raising the theoretical thermal efficiency.  FIG. 8(B)  shows one example of using the variable compression ratio mechanism A and the variable valve timing mechanism B to maintain the actual compression ratio at a low value while raising the expansion ratio. 
     If referring to  FIG. 8(B) , 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, the variable valve timing mechanism B is used to delay the closing timing of the intake valve until the actual piston stroke volume is reduced from 500 ml to 200 ml. As a result, in this example, the actual compression ratio becomes (20 ml+200 ml)/20 ml=11, while the expansion ratio becomes (20 ml+500 ml)/20 ml=26. In the normal cycle which is shown in  FIG. 8(A) , as explained above, the actual compression ratio becomes about 11 and the expansion ratio becomes 11. If compared with this case, in the case which is shown in  FIG. 8(B) , it is learned that only the expansion ratio is raised to 26. This is called the “superhigh expansion ratio cycle”. 
     Generally speaking, in an internal combustion engine, the lower the engine load, the worse the thermal efficiency. Therefore, to improve the thermal efficiency at the time of engine operation, that is, to improve the fuel economy, it is necessary to raise the thermal efficiency when the engine load is low. On the other hand, in the superhigh expansion ratio cycle, which is shown in  FIG. 8(B) , to enable the actual piston stroke volume at the time of the compression stroke to be made smaller, the amount of intake air which is sucked into a combustion chamber  5  becomes smaller. Therefore, this superhigh expansion ratio cycle can be employed only when the engine load is relatively low. Therefore, in the present invention, when the engine load is relatively low, the superhigh expansion ratio cycle, which is shown in  FIG. 8(B)  is used, while at the time of the engine high load operation, the normal cycle which is shown in  FIG. 8(A)  is used. 
     Next, referring to  FIG. 9 , the operational control as a whole will be schematically explained.  FIG. 9  shows the changes in the intake air amount, intake valve closing timing, mechanical compression ratio, expansion ratio, actual compression ratio, and opening degree of the throttle valve  17  in accordance with the engine load at a certain engine speed. Note that,  FIG. 9  shows the case where the mean air-fuel ratio in a combustion chamber  5  is feedback controlled to the stoichiometric air-fuel ratio based on the output signal of the air-fuel ratio sensor  21  so that the three-way catalyst in the catalyst device  20  can be used to simultaneously reduce the unburned HCs, CO, and NO x  in the exhaust gas. 
     As explained above, at the time of engine high load operation, the normal cycle which is shown in  FIG. 8(A)  is performed. Therefore, as shown in  FIG. 9 , at this time, the mechanical compression ratio is made low, so the expansion ratio is low. As shown by the solid line in  FIG. 9 , the closing timing of the intake valve  7  is advanced as shown by the solid line in  FIG. 5 . Further, at this time, the amount of intake air is large. At this time, the opening degree of the throttle valve  17  is held wide open, so the pumping loss becomes zero. 
     On the other hand, as shown in  FIG. 9  by the solid line, if the engine load becomes lower, along with this, the closing timing of the intake valve  7  is delayed so as to reduce the amount of intake air. Further, at this time, the mechanical compression ratio is increased as the engine load becomes lower as shown in  FIG. 9  so that the actual compression ratio is held substantially constant. Therefore, the expansion ratio is also increased as the engine load becomes lower. Note that, at this time as well, the throttle valve  17  is held in the wide open state. Therefore the amount of intake air which is supplied to the inside of the combustion chamber  5  is controlled without relying on the throttle valve  17  and by changing the closing timing of the intake valve  7 . 
     In this way, when the engine load becomes lower from the engine high load operation state, the mechanical compression ratio is made to increase under a substantially constant actual compression ratio as the amount of intake air decreases. That is, the volume of the combustion chamber  5  when the piston  4  reaches compression top dead center is made to decrease in proportion to the decrease in the amount of intake air. Therefore, the volume of the combustion chamber  5  when the piston  4  reaches compression top dead center changes in proportion to the amount of intake air. Note that, at this time, in the example which is shown in  FIG. 9 , 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 in proportion to the amount of fuel. 
     If the engine load becomes further lower, the mechanical compression ratio is made to further increase. If the engine load falls down to the intermediate load L 1  somewhat near the low load, the mechanical compression ratio reaches the limit mechanical compression ratio (upper limit mechanical compression ratio) forming the structural limit of the combustion chamber  5 . When the mechanical compression ratio reaches the limit mechanical compression ratio, in the region with a load lower than the engine load L 1  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 low load side engine medium load operation and at the time of engine low load operation, that is, at the engine low load operation side, the mechanical compression ratio becomes maximum and the expansion ratio also becomes maximum. In other words, at the engine low load operation side, the mechanical compression ratio is made maximum so that the maximum expansion ratio is obtained. 
     On the other hand, in the embodiment which is shown in  FIG. 9 , if the engine load falls to L 1 , the closing timing of the intake valve  7  becomes the limit closing timing by which the amount of intake air which is supplied to the inside of a combustion chamber  5  can be controlled. If the closing timing of the intake valve  7  reaches the limit closing timing, in the region of a lower load than the engine load L1 when the closing timing of the intake valve  7  reached 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, it is no longer possible to control the intake air amount by changing the closing timing of the intake valve  7 . In the embodiment which is shown in  FIG. 9 , at this time, that is, in the region of a load lower than the engine load L 1  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 which is supplied to the inside of the combustion chamber  5 . The lower the engine load is, the smaller the opening degree of the throttle valve  17  is made. 
     On the other hand, as shown by the broken line in  FIG. 9 , it is possible to control the amount of intake air without relying on the throttle valve  17  by advancing the closing timing of the intake valve  7  as the engine load becomes lower as shown by the broken line in  FIG. 9 . Therefore, if expressing the invention to be able to include both the case which is shown by the solid line in  FIG. 9  and the case which is shown by the broken line, in the embodiment according to the present invention, as the engine load becomes lower, the closing timing of the intake valve  7  is made to move in a direction away from suction bottom dead center BDC until the limit closing timing L 1  where the amount of intake air which is supplied to the inside of the combustion chamber can be controlled. In this way, the intake air amount can be controlled by changing the closing timing of the intake valve  7  as shown by the solid line in  FIG. 9  and can be controlled by changing it as shown by the broken line. 
     As explained above, in the superhigh expansion ratio cycle which is shown in  FIG. 8(B) , the expansion ratio is made  26 . This expansion ratio is preferably as high as possible, but as will be understood from  FIG. 7 , a considerably high theoretical thermal efficiency can be obtained if  20  or more compared with the practically usable lower limit actual compression ratio ε=5. Therefore in the present embodiment, the variable compression ratio mechanism A is formed so that the expansion ratio becomes 20 or more. 
     In the internal combustion engine of the present embodiment, as shown in  FIG. 9 , a target mechanical compression ratio is set for the current engine load. For the actuator of the variable compression ratio mechanism A, that is, the drive motor  59 , to realize the current target mechanical compression ratio, the actuating amount is controlled to become an actuating amount corresponding to the current target mechanical compression ratio. The actuating amount of the drive motor  59  (number of rotations having parts below decimal place) may be detected directly by a specific sensor (not shown), but may also be indirectly detected based on the relative positions between the crankcase  1  and cylinder block  2  which are detected by the above-mentioned relative position sensor  22  or the rotational angle of the cam shaft  55  which is detected by the above-mentioned cam rotational angle sensor  25 . 
     However, even if the actuator of the variable compression ratio mechanism A is controlled in this way, in actuality, the current target mechanical compression ratio is sometimes not realized. If the target mechanical compression ratio is not realized, the current desired expansion ratio is also not realized and the thermal efficiency also cannot be sufficiently raised. 
     The internal combustion engine which is provided with a variable compression ratio mechanism of the present embodiment is designed to estimate the current mechanical compression ratio in the specific engine operating state based on the first flow chart which is shown in  FIG. 10 . First, at step  101 , the current engine load which is detected by the load sensor  41  and the current engine speed which is detected by the crank angle sensor  42  are used as the basis to judge if the current steady engine operating state where the engine load and the engine speed are not changing is the specific engine operating state. When this judgment is no, the routine ends as is, but when the specific engine operating state, the judgment of step  101  is yes. At step  102 , the temperature sensor  28  is used to detect the current exhaust gas temperature T, which changes in accordance with the actual expansion ratio. 
     Next, at step  103 , the temperature difference ΔT between the current exhaust gas temperature T and the ideal exhaust gas temperature T′ when the target mechanical compression ratio is realized in the specific engine operating state is calculated. Next, at step  104 , the map which is shown in  FIG. 11  is used as the basis to set the mechanical compression ratio correction amount AE for the temperature difference ΔT. In the map of  FIG. 11 , if the temperature difference ΔT is 0, that is, if the current exhaust gas temperature T is equal to the ideal exhaust gas temperature T′, the target mechanical compression ratio is realized, the desired expansion ratio is also realized, and the mechanical compression ratio correction amount ΔE becomes 0. However, when the temperature difference ΔT is larger than 0, the current exhaust gas temperature T is higher than the ideal exhaust gas temperature T′, the current mechanical compression ratio is lower than the target mechanical compression ratio, the expansion ratio becomes lower than the desired value, and the thermal efficiency deteriorates. Further, when the temperature difference ΔT is smaller than 0, the current exhaust gas temperature T is lower than the ideal exhaust gas temperature T′, so the current mechanical compression ratio becomes higher than the target mechanical compression ratio, the expansion ratio also becomes higher than the desired value, and the thermal efficiency is improved more than necessary. Further, at this time, the actual compression ratio becomes higher than the constant value which is shown in  FIG. 9  and knocking easily occurs. 
     In the map which is shown in  FIG. 11 , overall, the larger the temperature difference ΔT, the smaller the mechanical compression ratio correction amount ΔE is set to become. When the temperature difference ΔT is larger than 0, the mechanical compression ratio correction amount ΔE is made a minus value, while when the temperature difference ΔT is smaller than 0, the mechanical compression ratio correction amount AE is made a plus value. 
     Next, at step  105 , the current actual mechanical compression ratio Er is calculated by adding the mechanical compression ratio correction amount ΔE which was set at step  104  to the current target mechanical compression ratio Et. In this way, the actual mechanical compression ratio Er at the time of the specific engine operating state can be estimated. 
     Of course, if the ideal exhaust gas temperature T′ where the target mechanical compression ratio is realized is set at each engine operating state in advance in a map etc., the temperature difference ΔT from the current exhaust gas temperature T can be calculated. As a result, in each engine operating state, if a map of the mechanical compression ratio correction amount ΔE for temperature difference ΔT such as shown in  FIG. 11  is set, the current actual mechanical compression ratio can be estimated at each engine operating state. 
     At step  106 , the actuating amount Ar of the actuator of the variable compression ratio mechanism A corresponding to the estimated current actual mechanical compression ratio Er is calculated. Next, at step  107 , the difference between the actuating amount Ar which was calculated at step  106  and the target actuating amount At of the actuator corresponding to the current (specific engine operating state) target mechanical compression ratio Et is calculated as the actuating amount correction amount ΔA. 
     The thus calculated actuating amount correction amount ΔA is an amount of deviation between the actual actuating amount of the actuator and the actuating amount of the actuator which is calculated based on the output of a detection device such as the relative position sensor  22  or the cam rotational angle sensor  25 . By correcting by addition the actuating amount which is calculated based on the output of the detection device, it is possible to calculate the current actual actuating amount. Due to this, if controlling the actuator of the variable compression ratio mechanism A so that the thus corrected actuating amount becomes an actuating amount corresponding to the target mechanical compression ratio of each engine operating state, it is possible to realize the target mechanical compression ratio at each engine operating state. 
     Further, the internal combustion engine which is provided with a variable compression ratio mechanism of the present embodiment enables estimation of the current mechanical compression ratio at the specific engine operating state by the second flow chart which is shown in  FIG. 12 . First, at step  201 , the current engine load which is detected by the load sensor  41  and the current engine speed which is detected by the crank angle sensor  42  are used as the basis to judge if the current steady engine operating state is a specific engine operating state. When this judgment is no, the routine is ended as it is, but when a specific engine operating state, the judgment of step  201  is yes. At step  202 , the pressure sensor  29  is used to detect the current exhaust gas pressure PE which changes according to the actual expansion ratio. 
     Next, at step  203 , the pressure difference ΔPE between the current exhaust gas pressure PE and the ideal exhaust gas pressure PE′ when the target mechanical compression ratio is realized at the specific engine operating state is calculated. Next, at step  204 , the map which is shown in  FIG. 13  is used as the basis to set the mechanical compression ratio correction amount ΔE for the pressure difference ΔPE. In the map of  FIG. 13 , if the pressure difference ΔPE is 0, that is, if the current exhaust gas pressure PE is the ideal exhaust gas pressure PE′, the target mechanical compression ratio is realized, the desired expansion ratio is also realized, and the mechanical compression ratio correction amount ΔE becomes 0. However, when the pressure difference ΔPE is larger than 0, the current exhaust gas pressure PE is higher than the ideal exhaust gas pressure PE′, so the current mechanical compression ratio becomes lower than the target mechanical compression ratio and the expansion ratio also becomes lower than the desired value and the thermal efficiency deteriorates. Further, when the pressure difference ΔPE is smaller than 0, the current exhaust gas pressure PE is lower than the ideal exhaust gas pressure PE′, so the current mechanical compression ratio becomes higher than the target mechanical compression ratio and the expansion ratio also becomes higher than the desired value, so the thermal efficiency is improved more than necessary. Further, at this time, the actual compression ratio becomes higher than the constant value shown in  FIG. 9  and knocking easily occurs. 
     In the map which is shown in  FIG. 13 , overall, the larger the pressure difference ΔPE, the smaller the mechanical compression ratio correction amount ΔE is set. When the pressure difference ΔPE is larger than 0, the mechanical compression ratio correction amount ΔE is made a minus value, while when the pressure difference ΔPE is smaller than 0, the mechanical compression ratio correction amount ΔE is made a plus value. 
     Next, at step  205 , the current actual mechanical compression ratio Er is calculated by adding the mechanical compression ratio correction amount ΔE which was set at step  204  to the current target mechanical compression ratio Et. In this way, it is possible to estimate the actual mechanical compression ratio Er at the time of a specific engine operating state. 
     However, if the ideal exhaust gas pressure PE′ when the target mechanical compression ratio is realized is set at each engine operating state in advance in a map etc., it is possible to calculate the pressure difference ΔPE with the current exhaust gas pressure PE. As a result, if a map of the mechanical compression ratio correction amount ΔE with respect to the pressure difference ΔPE such as shown in  FIG. 13  is set at each engine operating state, it is possible to estimate the current actual mechanical compression ratio at each engine operating state. 
     At step  206 , the actuating amount Ar of the actuator of the variable compression ratio mechanism A which corresponds to the estimated current actual mechanical compression ratio Er is calculated. Next, at step  207 , the difference between the actuating amount Ar which was calculated at step  206  and the target actuating amount At of the actuator which corresponds to the current (specific engine operating state) target mechanical compression ratio Et is calculated as the actuating amount correction amount ΔA. 
     The thus calculated actuating amount correction amount ΔA is the amount of deviation between the actual actuating amount of the actuator and the actuating amount of the actuator which is calculated based on the output of a detection device such as the relative position sensor  22  or cam rotational angle sensor  25 . By using the output of the detection device as the basis to correct by addition the calculated actuating amount, it is possible to calculate the current actual actuating amount. Due to this, if controlling the actuator of the variable compression ratio mechanism A so that the corrected actuating amount becomes an actuating amount corresponding to the target mechanical compression ratio at each engine operating state, it is possible to realize the target mechanical compression ratio at each engine operating state. 
       FIG. 14  is a schematic overall view of the case where an internal combustion engine which is provided with a variable compression ratio mechanism has a turbocharger. In the figure, members which were explained in  FIG. 1  are assigned the same reference numerals and explanations will be omitted. In the present embodiment, the intake duct  14 ′ between the surge tank  12  and the air cleaner  15  has a compressor  90  of the turbocharger arranged at it. At the upstream side of the compressor  90  of the turbocharger, the compressor  91  of the supercharger is arranged. 
     The compressor  90  of the turbocharger cannot sufficiently raise the supercharging pressure when the exhaust pressure is low such as the time of engine low speed. To assist the supercharging of the turbocharger at the time of engine low speed, the compressor  91  of the supercharger is provided. 
     The compressor  91  of the supercharger is an engine driven type. It is connected to an engine drive shaft through a solenoid clutch (not shown) and can be stopped by disengaging it from the engine drive shaft by the solenoid clutch. 
     If the compressor  91  of the supercharger is designed to be efficiently driven by the engine drive shaft at the time of engine low speed operation, it can be damaged by excessive rotation at the time of engine high speed operation, so the compressor  91  of the supercharger is disengaged from the engine drive shaft by the solenoid clutch if the engine speed becomes a set speed or more. 
       92  is a supercharging pressure sensor for measuring the intake pressure at the downstream side of the compressor  90  of the turbocharger of the intake duct  14 ′ as the supercharging pressure, while  93  is an intake pressure sensor for measuring the intake pressure of the intake duct  14 ′ between the compressor  90  of the turbocharger and the compressor  91  of the supercharger.  94  is an intercooler for cooling the intake which is supercharged by the compressor  90  of the turbocharger. 
     On the other hand, at the exhaust duct  95  at the downstream side of the exhaust manifold  19 , a turbine  96  of the turbocharger is arranged at the upstream side of the catalyst device  20 .  97  is a wastegate passage which bypasses the turbine  96 , while a wastegate valve  98  which controls the amount of exhaust which passes through the wastegate passage  97  is arranged at the wastegate passage  97 . 
     The internal combustion engine which is provided with a variable compression ratio mechanism of the present embodiment is designed to use the third flow chart which is shown in  FIG. 15  to estimate the current mechanical compression ratio at a specific engine operating state. First, at step  301 , the current engine load which is detected by the load sensor  41  and the current engine speed which is detected by the crank angle sensor  42  is used as the basis to judge if the current steady engine operating state is in the specific engine operating state. When this judgment is no, the routine is ended as it is, but when in a specific engine operating state, the judgment of step  301  is yes. At step  302 , the supercharging pressure sensor  92  is used to detect the current supercharging pressure PI of the turbocharger which changes in accordance with the exhaust temperature and exhaust pressure. 
     When the compressor  91  of the supercharger is not provided or the compressor  91  of the supercharger does not operate in the specific engine operating state, the current supercharging pressure PI which is detected by the supercharging pressure sensor  92  becomes a physical quantity which changes in accordance with the actual expansion ratio, but when the compressor  91  of the supercharger is used for supercharging, it is necessary to eliminate that effect. Specifically, as the supercharging pressure PI, the differential pressure before and after the compressor  90  of the turbocharger, that is, the differential pressure between the pressure which is detected by the supercharging pressure sensor  92  and the pressure which is detected by the intake pressure sensor  93 , is detected. 
     Next, at step  303 , the supercharging pressure difference ΔPI between the current supercharging pressure PI and the ideal supercharging pressure PI′ when the target mechanical compression ratio is realized in a specific engine operating state is calculated. Next, at step  304 , the map which is shown in  FIG. 16  is used as the basis to set the mechanical compression ratio correction amount ΔE for the supercharging pressure difference ΔPI. In the map of  FIG. 16 , if the supercharging pressure difference ΔPI is 0, that is, if the current supercharging pressure PI is equal to the ideal supercharging pressure PI′, the target mechanical compression ratio is realized, the desired expansion ratio is also realized, and the mechanical compression ratio correction amount ΔE becomes zero. However, when the supercharging pressure difference ΔPI is larger than 0, the current supercharging pressure PI is higher than the ideal supercharging pressure PI′, so the current mechanical compression ratio becomes lower than the target mechanical compression ratio, the expansion ratio also becomes lower than the desired value, and the thermal efficiency deteriorates. Further, when the supercharging pressure difference ΔPI is smaller than 0, the current supercharging pressure PI is lower than the ideal supercharging pressure PI′, the current mechanical compression ratio becomes higher than the target mechanical compression ratio, the expansion ratio also becomes higher than the desired value, and the thermal efficiency is improved more than necessary. Further, at this time, the actual compression ratio becomes higher than the constant value which is shown in  FIG. 9  and knocking easily occurs. 
     In the map which is shown in  FIG. 16 , overall, the larger the supercharging pressure difference ΔPI, the smaller the mechanical compression ratio correction amount ΔE is set to be. When the supercharging pressure difference ΔPI is larger than 0, the mechanical compression ratio correction amount ΔE is made a minus value, while when the supercharging pressure difference ΔPI is smaller than 0, the mechanical compression ratio correction amount ΔE is made a plus value. 
     Next, at step  305 , the current actual mechanical compression ratio Er is calculated by adding the mechanical compression ratio correction amount ΔE which is set at step  304  to the current target mechanical compression ratio Et. In this way, it is possible to estimate the actual mechanical compression ratio Er at the time of the specific engine operating state. 
     However, if the ideal supercharging pressure PI′ when the respective target mechanical compression ratios are realized is set at each engine operating state in a map etc. in advance, the supercharging pressure difference ΔPI relating to the current supercharging pressure PI can be calculated. As a result, if a map of the mechanical compression ratio correction amount ΔE for the supercharging pressure difference ΔPI such as shown in  FIG. 16  is set at each engine operating state, it is possible to estimate the current actual mechanical compression ratio at each engine operating state. 
     At step  306 , the actuating amount Ar of the actuator of the variable compression ratio mechanism A corresponding to the estimated current actual mechanical compression ratio Er is calculated. Next, at step  307 , the difference between the actuating amount Ar which was calculated at step  306  and the target actuating amount At of the actuator corresponding to the current (specific engine operating state) target mechanical compression ratio Et is calculated as the actuating amount correction amount ΔA. 
     The thus calculated actuating amount correction amount ΔA is an amount of deviation between the actual actuating amount of the actuator and the actuating amount of the actuator which is calculated based on the output of a detection device such as the relative position sensor  22  or cam rotational angle sensor  25 . By correcting by addition the actuating amount which is calculated based on the output of the detection device, the current actual actuating amount can be calculated. Due to this, if controlling the actuator of the variable compression ratio mechanism A so that the thus corrected actuating amount becomes an actuating amount corresponding to the target mechanical compression ratio at each engine operating state, it is possible to realize the target mechanical compression ratio at each engine operating state. 
     In the present embodiment, as the physical quantity which changes in accordance with at least one of the exhaust temperature and exhaust pressure, which change in accordance with the actual expansion ratio, the supercharging pressure of the turbocharger is measured and the measured supercharging pressure is used as the basis to estimate the current mechanical compression ratio. The supercharging pressure at the downstream side of the compressor  90  of the turbocharger can be measured using a generally provided supercharging pressure sensor  92 . It is not necessary to newly provide a sensor for measuring a physical quantity. 
     Further, as the physical quantity which changes in accordance with at least one of the exhaust temperature and exhaust pressure, which change in accordance with the actual expansion ratio, for estimating the current mechanical compression ratio, in addition to the supercharging pressure of the turbocharger, the speed of the turbine of the turbocharger etc. can also be measured. 
     In the first, second, and third flow charts, the specific engine operating state is preferably made an engine operating state whereby the target mechanical compression ratio becomes the set mechanical compression ratio or less. In this way, if the target mechanical compression ratio is small, the exhaust temperature or the exhaust pressure changes relatively largely according to the slight deviation in mechanical compression ratio when the target mechanical compression ratio is not realized, so it is possible to accurately detect a slight deviation in the mechanical compression ratio and possible to accurately correct the actuating amount A of the actuator of the variable compression ratio mechanism A which is detected by the detection device. 
     In the third flow chart, the supercharging pressure difference ΔPI between the supercharging pressure PI of the turbocharger which is measured at the time of a specific engine operating state (when a supercharger is provided, differential pressure before and after the compressor  90  of the turbocharger) and the ideal supercharging pressure PI′ of the turbocharger when the desired expansion ratio of the specific engine operating state is realized is used as the basis to estimate the actual mechanical compression ratio of the specific engine operating state. 
     However, strictly speaking, due to individual differences, for each turbocharger, sometimes deviation occurs in the measurement value PI of the supercharging pressure in a specific engine operating state. Due to this, by using the supercharging pressure true value PIr where this amount of deviation is eliminated instead of the measurement value PI at step  303  of the third flow chart to calculate the supercharging pressure difference ΔPI, it is possible to accurately estimate the mechanical compression ratio. 
     As shown in  FIG. 14 , when the exhaust duct  95  of an internal combustion engine which is provided with a variable compression ratio mechanism is provided with a wastegate passage  97  which bypasses the turbine  96  of the turbocharger, the fourth flow chart which is shown in  FIG. 17  can be used to set the supercharging pressure true value PIr. 
     First, at step  401 , in the same way as step  301  of the third flow chart, it is judged if the engine is in a specific engine operating state. If this judgment is no, the routine is ended as it is, but when the judgment of step  401  is yes, at step  402 , the first supercharging pressure PI 1  of the turbocharger when making the opening degree of the wastegate valve  98  the desired opening degree at the specific engine operating state, defined as the first opening degree TA 1  (for example, the fully closed opening degree), is measured. 
     Next, at step  403 , the opening degree of the wastegate valve  98  is made a second opening degree TA 2  larger than the first opening degree TA 1  (for example, a half opening degree), then, at step  404 , the second supercharging pressure PI 2  of the turbocharger when making the opening degree of the wastegate valve  98  the second opening degree TA 2  at the specific engine operating state is measured. Next, at step  405 , the difference dPI between the first supercharging pressure PI 1  and the second supercharging pressure PI 2  is calculated. 
     At this difference dPI, even if the measurement value PI 1  of the supercharging pressure of the first opening degree TA 1  of the wastegate valve  98  includes the amount of deviation due to individual differences of the turbocharger, since the measurement value PI 2  of the supercharging pressure of the second opening degree TA 2  of the wastegate valve  98  includes the same amount of deviation, the amount of deviation due to individual differences is cancelled out. In this way, at step  406 , the supercharging pressure true value PIr is set based on the difference dPI. 
     The larger the opening degree TA of the wastegate valve  98 , the greater the amount of exhaust gas which passes through the wastegate passage  97  and does not pass through the turbine  96  of the turbocharger and the lower the supercharging pressure.  FIG. 18  shows the changes in the design supercharging pressure with respect to the opening degree of the wastegate valve  98  in a specific engine operating state. The plurality of solid lines are differences in the expansion ratio. As shown in  FIG. 18 , the higher the supercharging pressure PI when the wastegate valve  98  is the first opening degree TA 1  in a specific engine operating state, the larger the fall in the supercharging pressure when making the wastegate valve  98  the second opening degree TA 2 , that is, the above-mentioned difference dPI. 
     In this way, it is possible to unambiguously set in advance the supercharging pressure true value PIr for when the wastegate valve  98  is the first opening degree TA 1  for the difference dPI. That is, as illustrated in  FIG. 18 , the supercharging pressure true value PIr when the difference dPI is dPI 1  (supercharging pressure not including amount of deviation due to individual differences when the wastegate valve  98  is made the first opening degree TA 1  in a specific engine operating state) becomes PIr 1 , the supercharging pressure true value PIr when the difference dPI is dPI 2  becomes PIr 2 , and the supercharging pressure true value PIr when the difference dPI is dPI 3  becomes PIr 3 . In this way, it is possible to set in advance the corresponding supercharging pressure true value for other values of the difference dPI. In this way, at step  406 , it is possible to use the difference dPI as the basis to set the supercharging pressure true value PIr. 
     The second opening degree TA 2  of the wastegate valve  98  was made an opening degree larger than the first opening degree TA 1 , but unless the desired first opening degree TA 1  of the wastegate valve  98  at the specific engine operating state is the fully closed opening degree, the second opening degree TA 2  may be made smaller than the first opening degree TA 1 . 
     If, like in the first, second, and third flow charts, the accurate mechanical compression ratio Er at the time of a specific engine operating state is estimated, by estimating the actual compression ratio at the time of a specific engine operating state, it is possible to use the estimated mechanical compression ratio and the actual compression ratio as the basis to estimate the accurate closing timing of the intake valve at the time of the specific engine operating state and possible to calculate the correction amount of the closing timing of the intake valve which is detected by the valve timing sensor  23 . It is possible to use any method to estimate the actual compression ratio of a specific engine operating state and for example estimate the actual compression ratio based on the fact that the higher the fuel pressure or the easier knocking is to occur, the higher the actual compression ratio becomes. 
     REFERENCE SIGNS LIST 
     
         
           28  temperature sensor 
           29  pressure sensor 
           90  compressor of turbocharger 
           92  supercharging pressure sensor 
           96  turbine of turbocharger 
           97  wastegate passage 
           98  wastegate valve 
         A variable compression ratio mechanism 
         B variable valve timing mechanism

Technology Classification (CPC): 5