RU2434152C1 - Engine control system - Google Patents

Abstract

FIELD: engines and pumps.

SUBSTANCE: system includes variable compression degree mechanism and gas distribution phase control mechanism. When the required engine output power is boundary output power or less, control is performed in order to maintain minimum fuel flow level. When the required engine output power increases above boundary output power, mechanical compression degree decreases to minimum mechanical compression degree, and then the engine output power increases.

EFFECT: reducing fuel flow at increase in the required engine power.

9 cl, 37 dwg

Description

Technical field

The present invention relates to an engine control system.

State of the art

A hybrid vehicle is known in the art for using one or both of an engine and an electric motor to drive a vehicle, the engine consisting of an engine equipped with a variable compression ratio mechanism, the compression ratio is set, whereby the overall efficiency including the efficiency of the engine, the efficiency of the electric motor, the efficiency of the power transmission system and all other efficiency becomes maximum, and the compression ratio of the engine is controlled to a compression ratio that gives this r maximum overall efficiency (see., e.g., the publication of Japanese Patent Application № 2004-44433 A).

Technical problem

However, even if only the compression ratio is controlled so that the overall efficiency becomes maximum, there is a limitation on improving fuel consumption. At present, a vehicle improvement giving a better level of fuel consumption is desired.

An object of the present invention is to provide an engine control system capable of obtaining a better level of fuel consumption, while at the same time providing the required engine output when the required engine output increases, by controlling the mechanical compression ratio and the timing of closing the intake valve, and capable of eliminating the occurrence of vibration and noise.

Solution

According to the present invention, an engine control system is provided with an output power control system that allows you to specify the desired combination of engine torque and engine speed giving the same output power, while there is a variable compression ratio mechanism capable of changing the mechanical compression ratio and a valve timing control mechanism able to control when the intake valve closes and when the required engine output is lower than twice defined boundary output power, control is performed to maintain a minimum fuel consumption level satisfying the required output power by changing the engine speed in a state that maintains a mechanical compression ratio by a predetermined compression ratio or more, and when the required engine output increases above the boundary output power, control is performed to increase the output power, increasing the engine torque by after reducing the mechanical compression ratio to a predetermined compression ratio or less.

Brief Description of the Drawings

Figure 1 - General view of the engine and the output power control system;

Figure 2 is a view for explaining the operation of the output power control system;

Figure 3 is a view showing the relationship between the engine power output and the engine torque Te and the engine speed Ne, etc .;

4 is a flowchart of operational control of the vehicle;

5 is a view illustrating the management of charge and discharge of the battery;

Fig.6 is a General view of the engine shown in Fig.1;

Fig.7 is a perspective view with exploded parts of the mechanism of variable compression ratio;

Fig. 8 is a side cross-sectional view of an engine shown schematically;

Fig.9 is a view showing a valve timing control mechanism;

Figure 10 is a view showing the amount of raising of the intake valve and exhaust valve;

11 is a view for explaining the degree of mechanical compression of the engine and the degree of actual compression and the degree of expansion;

12 is a view showing the relationship between theoretical thermal efficiency and the degree of expansion;

13 is a view showing a conventional cycle and an ultrahigh expansion ratio cycle;

14 is a view showing changes in the degree of mechanical compression in accordance with engine torque, etc .;

Fig. 15 is a view showing equal lines of a fuel consumption level and a line of operation;

Fig is a view showing changes in the level of fuel consumption and the degree of mechanical compression;

17 is a view showing equivalent lines of a fuel consumption level and a line of operation;

Fig. 18 is a view showing equivalent output power levels and marginal engine output powers;

Fig. 19 is an illustration of a state of engine torque Te and engine speed Ne when the required output power increases or decreases;

Fig. 20 is a timing chart showing changes in the degree of mechanical compression, the closing timing of the intake valve, etc .;

21 is a view showing a state of changes in engine torque Te and engine speed Ne when the required engine output increases or decreases;

FIG. 22 is a view showing a state of changes in engine torque Te and engine speed Ne when the required engine output increases or decreases;

23 is a timing chart showing changes in the degree of mechanical compression, the closing timing of the intake valve, etc .;

24 is a flowchart of operational control;

25 is a view showing a state of changes in engine torque Te and engine speed Ne when the required engine output increases or decreases;

Fig. 26 is a view showing a state of changes in engine torque Te and engine speed Ne when the required engine output increases or decreases;

Fig. 27 is a timing chart showing changes in the degree of mechanical compression, the closing timing of the intake valve, etc .; and

Fig. 28 is a flowchart of operational control.

BEST MODE FOR CARRYING OUT THE INVENTION

1 is a perspective view of a spark ignition engine 1 and an output power control system 2 installed in a hybrid vehicle.

First, it will be described with reference to FIG. 1, an output power control system 2. In the embodiment shown in FIG. 1, the output power control system 2 consists of a pair of electric motors-generators MG1 and MG2, operating as electric motors and generators, and a planetary gear mechanism 3. The planetary gear mechanism 3 is provided with a sun gear 4, a ring gear 5 planetary gears 6, located between the sun gear 4 and the ring gear 5, and the planet carrier planet 7, carrying planetary gears 6. The sun gear 4 is connected to the shaft 8 of the motor-generator MG1, while the planet carrier 7 is connected to the output shaft 9 of the engine. Additionally, the ring gear 5 is connected on one side to the shaft 10 of the motor-generator MG2, and on the other hand connected to the output shaft 12 connected to the drive wheels through the belt 11. Therefore, it will be understood that if the ring gear 5 rotates, the output shaft 12 rotates with it.

Electric motors-generators MG1 and MG2, respectively, consist of AC-synchronized electric motors equipped with rotors 13 and 15 attached to the respective shafts 8 and 10 and having many permanent electromagnets attached to external circles, and stators 14 and 16 equipped with excitation coils, forming rotating magnetic fields. The excitation coils of the stators 14 and 16 of the motor generators MG1 and MG2 are connected to the respective motor excitation control circuits 17 and 18, while these electric motor excitation control circuits 17 and 18 are connected to the battery 19, which generates a high DC voltage. In the embodiment shown in FIG. 1, the motor generator MG2 mainly operates as an electric motor, while the electric motor generator MG1 mainly operates as a generator.

The electronic control unit 20 consists of a digital computer and is equipped with a ROM 22, RAM 23, a CPU (microprocessor) 24, an input port 25 and an output port 26, which are interconnected via a bi-directional bus 21. The accelerator pedal 27 is connected to a load sensor 28 forming output voltage proportional to the pressure L of the accelerator pedal 27. The output voltage of the load sensor 28 is inputted through the corresponding ADC (analog-to-digital converter) 25a into the input port 25. In addition, the input port 25 is connected to the crankshaft angle sensor 29, which generates an output pulse each time the crankshaft rotates, for example, by 15 °. In addition, the input port 25 receives, as input, a signal expressing the charge and discharge current of the battery 19, and other various signals through the corresponding ADC 25a. On the other hand, the output port 26 is connected to the drive excitation control circuits 17 and 18 and is connected via the corresponding drive circuit 26a to components for controlling the engine 1, for example, a fuel injector, etc.

Upon excitation of the motor generator MG2, the high DC voltage of the battery 19 is converted in the circuit 18 for controlling the excitation of the electric motor into a three-phase alternating current with a frequency fm and a current value Im. This three-phase alternating current is supplied to the excitation coil of the stator 16. This frequency fm is the frequency required to cause the magnetic field generated by the field coil to rotate in synchronization with the rotation of the rotor 15. This frequency fm is calculated by the CPU 24 based on the speed of the output shaft 10. V an electric motor excitation control circuit 18, this frequency fm is equal to the frequency of a three-phase alternating current. On the other hand, the output torque of the motor generator MG2 becomes substantially proportional to the value Im of the three-phase alternating current. This current value Im is calculated based on the desired output torque of the motor generator MG2. In the motor excitation control circuit 18, this current value Im is made by a three-phase alternating current value.

Further, if the state setting uses external force to drive the motor generator MG2, the motor generator MG2 acts as a generator. The energy generated at this time is regenerated in the battery 19. The required drive torque, when an external force is used to drive the motor generator MG2, is calculated in the CPU 24. The drive drive control circuit 18 operates so that this required drive torque acts on shaft 10.

This type of drive control in the motor generator MG2 is similarly performed in the motor generator MG1. That is, upon excitation of the motor generator MG1, the high DC voltage of the battery 19 is converted in the excitation control circuit 17 of the electric motor into a three-phase alternating current with a frequency fm and a current value Im. This three-phase alternating current is supplied to the excitation coil of the stator 14. Additionally, if the state setting uses external force to drive the motor generator MG1, the motor generator MG1 acts as a generator. The energy generated at this time is regenerated in the battery 19. At this time, the motor drive control circuit 17 operates so that the calculated required drive torque acts on the shaft 8.

Next will be explained with reference to Figure 2 (A), illustrating the planetary gear mechanism 3, the ratio of the torques acting on the different shafts 8, 9 and 10, and the ratio of the speeds of the shafts 8, 9 and 10.

In FIG. 2 (A), r ₁ shows the radius of the initial circumference of the sun gear 4, while r ₂ shows the radius of the initial circumference of the ring gear 5. Now suppose that in the state shown in FIG. 2 (A), the torque Te is applied to the output shaft 9 of the engine 1, and a force F acting in the direction of rotation of the output shaft 9 is generated in the center of rotation of each planet gear 6. At this time, in the parts engaging with the planet gear 6, the sun gear 4 and the ring gear 5 act force f / 2 in the same direction as s la F. As a result, the torque Tes = ((F / 2) r ₁ ) acts on the shaft 8 of the sun gear 4, while the torque Ter = ((F / 2) r ₂ ) acts on the shaft 10 of the ring gear 5 . On the other hand, the torque Te acting on the output shaft 9 of the engine 1 is expressed as F (r ₁ + r ₂ ) / 2, thus, if the torque Tes acting on the shaft 8 of the sun gear 4 is expressed by r ₁ , r ₂ and Te, the result is Tes = (r ₁ / (r ₁ + r ₂ )) · Te, while if we express the torque Ter acting on the shaft 10 of the ring gear 5 by r ₁ , r ₂ and Te , the result is Ter = (r ₂ / (r ₁ + r ₂ )) · Te.

That is, the torque Te occurring on the output shaft 9 of the engine 1 is divided by the torque Tes acting on the shaft 8 of the sun gear 4 and the torque Ter acting on the shaft 10 of the ring gear 5 with a ratio of r ₁ : r ₂ . In this case, r ₂ > r ₁ , so that the torque Ter acting on the shaft 10 of the ring gear 5 always becomes larger than the torque Tes acting on the shaft 8 of the sun gear 4. It should be noted that if the radius r is determined _{1 of the} initial circumference of the sun gear / radius r _{2 of the} initial circumference of the ring gear 5, i.e., the number of teeth of the sun gear 4 / number of teeth of the ring gear 5 as ρ, Tes is expressed as Tes = (ρ / (1 + ρ)) · Te, and Ter expressed as Ter = (1 / (1 + ρ)) · Te.

On the other hand, if the direction of rotation of the output shaft 9 of the engine 1, that is, the direction of action of the torque Te, shown by the arrow sign in Figure 2 (A), is indicated as the direct direction when the rotation of the planet carrier 7 is stopped, and in this state the sun gear 4 rotates in the forward direction, ring gear 5 rotates in the opposite direction. At this time, the ratio of the speeds of the sun gear 4 and the ring gear 5 becomes r ₂ : r ₁ . The dashed line Z ₁ in FIG. 2 (B) illustrates the ratio of speeds at this time. It should be noted that in FIG. 2 (B), the ordinate shows a forward direction above 0 and a reverse direction below it. Additionally, in FIG. 2 (B), S shows the sun gear 4, C shows the planet carrier 7 and R shows the ring gear 5. As shown in FIG. 2 (B), if the distance between the planet carrier C of the planet gear and the ring gear R is set as r ₁ , the distance between planet carrier C of the planetary gear and sun gear S is set as r ₂ , and the speeds of the sun gear S, planet carrier C of the planet gear and ring gear R are shown with black dots, points showing speeds are located on the line shown by the broken line Z ₁ .

On the other hand, if the relative rotation of the sun gear 4, the ring gear 5 and the planetary gear 6 is stopped to cause the planet carrier 7 to rotate in the forward direction, the sun gear 4, the ring gear 5 and the planet carrier 7 will rotate in the forward direction with the same rotation speed. The ratio of speeds at this time is shown by the dashed line Z ₂ . Therefore, the ratio of actual speeds is expressed by the solid line Z obtained by superimposing the dashed line Z ₁ on the dashed line Z ₂ , therefore, the points showing the speeds of the sun gear S, planet carrier C and the ring gear R are located on the line shown by the solid line Z. Therefore, when any two speeds of the sun gear S, the planet carrier C and the ring gear R, are set, the remaining one speed is set automatically. It should be noted that if you use the above ratio r ₁ / r ₂ = ρ, as shown in Figure 2 (B), the distance between the sun gear C and the planet carrier C of the planet gear and the distance between the planet carrier C of the planet gear and the ring gear R becomes 1: ρ.

FIG. 2 (C) illustrates the speeds of the sun gear S, planet carrier C and the ring gear R and the torques acting on the sun gear S, planet carrier C and the ring gear R. The ordinate and abscissa in FIG. 2 (C) are the same as in Figure 2 (B). Additionally, the solid line shown in FIG. 2 (C) corresponds to the solid line shown in FIG. 2 (B). On the other hand, FIG. 2 (C) shows the torques acting on the respective shafts at black points indicating speeds. It should be noted that when the direction of action of the torque and the direction of rotation are the same for each torque, this shows the case where the drive torque is transmitted to the corresponding shaft, while the direction of the torque and the direction of rotation are opposite, this shows the case. where torque is transmitted to the corresponding shaft.

Now, in the example shown in FIG. 2 (C), an engine torque Te is applied to the planet carrier C of the planetary gear. This engine torque Te is divided by the torque Ter applied to the ring gear R and the torque Tes applied to the sun gear S. The divided ring torque Ter acts on the shaft 10 of the ring gear R, the torque Tm _{2 of the} motor generator MG2, and vehicle drive torque Tr for driving the vehicle. These torques Ter, Tm ₂ and Tr are balanced. In the case shown in FIG. 2 (C), the torque Tm ₂ is the moment where the direction of action of the torque and the direction of rotation are the same, so this torque Tm ₂ transfers the drive torque to the shaft 10 of the ring gear R. Therefore, at this time, the motor generator MG2 operates as a drive motor. In the case shown in FIG. 2 (C), the sum of the engine torque Ter divided at this time and the drive torque Tm ₂ by the motor generator MG2 becomes equal to the vehicle drive torque Tr. Therefore, at this time, the vehicle is driven by the engine 1 and the motor generator MG2.

On the other hand, the divided engine torque Tes and the torque Tm _{1 of the} motor generator MG1 are acting on the shaft 8 of the sun gear 5. These Tes and Tm ₁ torques are balanced. In the case shown in FIG. 2 (C), the torque Tm ₁ is such that the direction of action of the torque and the direction of rotation are opposite, so this torque Tm ₁ becomes the drive torque transmitted from the shaft 10 of the ring gear R Therefore, at this time, the motor generator MG1 operates as a generator. That is, the divided engine torque Tes becomes equal to the torque for driving the motor generator MG1. Therefore, at this time, the motor generator MG1 is driven by the engine 1.

In Fig. 2 (C), Nr, Ne and Ns respectively show the speeds of the shaft 10 of the ring gear R, the planet carrier shaft C of the planetary gear, that is, the drive shaft 9, and the sun gear shaft 8. Therefore, the ratio of the speeds of the shafts 8, 9 and 10 and the ratio of the torques acting on the shafts 8, 9 and 10 will be apparent at a glance in FIG. 2 (C). Figure 2 (C) is called a "nomogram." The solid line shown in FIG. 2 (C) is called the “operating line”.

Now, as shown in FIG. 2 (C), if the vehicle drive torque is Tr and the ring gear speed 5 is Nr, the output power Pr of the vehicle drive for driving the vehicle is expressed as Pr = Tr · Nr. Additionally, the output power Pe of the engine 1 at this time is expressed as the product Te · Ne of the engine torque Te and the engine speed Ne. On the other hand, at this time, the generated electric power of the motor generator MG1 is likewise expressed as a product of torque and rotational speed. Therefore, the generated electric power of the motor generator MG1 becomes equal to Tm ₁ · Ns. Additionally, the excitation energy of the motor generator MG2 is also expressed as the product of torque and speed. Therefore, the excitation energy of the motor generator MG2 becomes Tm ₂ · Nr. Here, assuming that the generated electric power Tm ₁ · Ns of the motor generator MG1 is equal to the driving energy Tm ₂ · Nr of the motor generator MG2, and the energy generated by the motor generator MG1 is used to drive the motor generator MG2, the total output power Pe of the motor 1 is used by the power output Pr of the vehicle drive. At this time, Pr = Pe, therefore, Tr · Nr = Te · Ne. That is, the engine torque Te is converted to the vehicle drive torque Tr. Therefore, the output power control system 2 performs a torque conversion action. It should be noted that, in reality, there is a loss in power generation and a gear loss, so the output power Pe of the engine 1 cannot be used for the output power Pr of the vehicle drive, but the output power control system 2 still performs the torque conversion action .

Figure 3 (A) shows the equivalent output power lines from Pe ₁ to Pe _{9 of} engine 1. Between the output powers, there is a relation Pe ₁ <Pe ₂ <Pe ₃ <Pe ₄ <Pe ₅ <Pe ₆ <Pe ₇ <Pe ₈ < Pe ₉ . It should be noted that the ordinate in FIG. 3 (A) shows the engine torque Te, while the abscissa in FIG. 3 (A) shows the engine rotation speed Ne. As will be understood from FIG. 3 (A), there are innumerable combinations of engine torque Te and engine speed Ne satisfying the required output power Pe of engine 1 requested to drive the vehicle. In this case, no matter what combination of the engine torque Te and the engine speed Ne is selected, it is possible to convert the engine torque Te to the vehicle drive torque Tr in the output power control system 2. Therefore, if you use this system 2 control the output power, it becomes possible to set the desired combination of engine torque Te and engine speed Ne, giving the same output power Pe of the engine. In the present invention, as will be explained later, a combination of engine torque Te and engine speed Ne is set, capable of providing the required output power Pe of engine 1 and obtain better fuel consumption. The ratio shown in FIG. 3 (A) is stored in advance in the ROM 22.

Figure 3 (B) shows the equivalent lines of the degree of opening of the accelerator of the accelerator pedal 27, that is, the equivalent depression lines L. Values L depressions are shown as percentages relative to the equivalent lines L depressions. It should be noted that the ordinate in FIG. 3 (B) shows the required vehicle drive torque TrX requested to drive the vehicle, while the abscissa in FIG. 3 (B) shows the rotation speed Nr of the ring gear 5. From Figure 3 (B) it will be understood that the required torque TrX of the vehicle drive is determined from the amount of pressing L of the accelerator pedal 27 and the rotation speed Nr of the ring gear 5 at this time. The ratio shown in FIG. 3 (B) is stored in advance in the ROM 22.

Next, referring to FIG. 4, a basic control procedure for driving a vehicle will be explained. It should be noted that this procedure is performed by interruption at predetermined time intervals.

Turning to FIG. 4, first, at step 100, the rotation speed Nr of the ring gear 5 is determined. Next, at step 101, the amount of pressing L of the accelerator pedal 27 is read. Next, in step 102, the required vehicle drive torque TrX is calculated from the ratio shown in FIG. 3 (B). Next, in step 103, the rotation speed Nr of the ring gear 5 is multiplied by the required vehicle drive torque TrX to calculate the required vehicle drive output power Pr = (TrX Nr). Next, in step 104, the required output power Pr of the vehicle’s drive is added to the output power Pd of the engine to increase or decrease to charge or discharge the battery 19, and the output power Ph of the engine required for auxiliary driving devices to calculate the output power Pn required from engine 1. It should be noted that the output power Pd of the engine for charging and discharging the battery 19 is calculated by the procedure shown in FIG. 5 (B) explained below.

Next, in step 105, the output power Pr required by the engine 1 is divided by the torque conversion efficiency ηt of the torque control system 2 of the output power so as to calculate the final required output power Pe of the engine 1 = (Pn / ηt). Next, in step 106, from the relation shown in FIG. 3 (A), the required engine torque TeX and the required engine rotation speed NeX, etc., are established that satisfy the required engine output Pe and give minimum fuel consumption. The required engine torque TeX and the required engine speed NeX, etc. set by the procedure shown in Figs. 24 and 28 explained below. It should be noted that in the present invention, "minimum fuel consumption" means minimum fuel consumption when not only the efficiency of the engine 1 is considered, but also the efficiency of the transmission mechanism of the output power control system 2 and etc.

Next, in step 107, the required torque Tm ₂ X of the motor generator MG2 = (TrX-Ter = TrX-TeX / (1 + ρ)) is calculated from the required vehicle drive torque TrX and the required engine torque TeX. Next, in step 108, the required rotation speed NsX of the sun gear 4 is calculated from the rotation speed Nr of the ring gear 5 and the required rotation speed NeX of the engine. It should be noted that from the relation shown in FIG. 2 (C), (NeX-Ns) :( Nr-NeX) = 1: ρ, thus, the required rotation speed NsX of the sun gear 4 is expressed as Nr- (Nr-NeX ) · (1 + ρ) / p, as shown by step 108 in FIG. 4.

Next, in step 109, the motor generator MG1 is controlled so that the rotation speed of the motor generator MG1 becomes the desired rotation speed NsX. If the rotation speed of the motor generator MG1 becomes the required rotation speed NsX, the engine rotation speed Ne becomes the required engine rotation speed NeX and, therefore, the engine rotation speed Ne is controlled by the motor generator MG1 to the desired engine rotation speed NeX. Next, in step 110, the motor generator MG2 is controlled so that the torque of the motor generator MG2 becomes the required torque Tm ₂ X. Next, in step 111, the amount of fuel injection required to obtain the required engine torque TeX and the target degree throttle openings are calculated. At 112, the engine 1 is controlled based on them.

In this regard, in a hybrid vehicle, it is necessary to maintain the accumulated charge of the battery 19 at a constant value or more all the time. Therefore, in the embodiment according to the present invention, as shown in FIG. 5 (A), the accumulated charge SOC is stored between the lower limit value SC ₁ and the upper limit value SC ₂ . That is, in the embodiment according to the present invention, if the accumulated charge SOC falls below the lower limit value SC ₁ , the output of the engine is forcibly increased in order to increase the amount of power generation. If the accumulated charge SOC exceeds the upper limit value SC ₂ , the engine output is forcibly reduced so as to increase the amount of energy consumed by the motor-generator. That is, the accumulated charge SOC, for example, is calculated by the combined addition of the charge current I and the discharge of the battery 19.

5 (B) shows a control procedure for charging and discharging the battery 19. This procedure is performed by interruption at predetermined time intervals.

Referring to FIG. 5 (B), first, at step 120, the accumulated charge SOC is added to the current I of the charge and discharge of the battery 19. This current I value is set to positive during charge and made negative during discharge. Next, in step 121, it is judged if the battery is in the middle of a charge. When not in the midst of the charge, the procedure proceeds to step 122, where it is evaluated whether the accumulated charge SOC is lower than the lower limit value SC ₁ . If SOC <SC ₁ , the procedure proceeds to step 124, where the output power Pd of the engine in step 104 of FIG. 4 becomes equal to the predetermined value Pd ₁ . At this time, the engine output is forcibly increased, and the battery 19 is forcibly charged. If the battery 19 is forcibly charged, the procedure proceeds from step 121 to step 123, where it is assessed whether the forced charge action has been completed. The procedure proceeds to step 124 until the action of the forced charge is completed.

On the other hand, when it is estimated that SOC≥SC _{1 is} in step 122, the procedure proceeds to step 125, where it is judged whether the battery 19 is in the middle of a forced discharge. When it is not in the middle of the forced discharge, the procedure proceeds to step 126, where it is evaluated whether the accumulated charge SOC has exceeded the upper limit value SC ₂ . If SOC> SC ₂ , the procedure proceeds to step 128, where the output power Pd of the engine in step 104 of FIG. 4 becomes equal to a predetermined value of Pd ₂ . At this time, the output of the engine is forcibly reduced, and the battery 19 is forcibly discharged. If the battery 19 is forcedly discharged, the procedure proceeds from step 125 to step 127, where it is evaluated whether the forced discharge action has been completed or not. The procedure proceeds to step 128 until the end of the forced discharge.

Next, the spark ignition type internal combustion engine shown in FIG. 1 will be explained with reference to FIG. 6.

6, reference numeral 30 denotes a crankcase, 31 - a cylinder block, 32 - a cylinder head, 33 - a piston, 34 - a combustion chamber, 35 - a spark plug placed at the top dead center of a combustion chamber 34, 36 - an intake valve, 37 - inlet, 38 - exhaust valve, 39 - outlet. The inlet 37 is connected through the inlet 40 to the surge tank 41, while each inlet 40 is provided with a fuel nozzle 42 for injecting fuel into the corresponding inlet 37. It should be noted that each fuel nozzle 42 can be placed in each combustion chamber 34 instead of attaching to each inlet pipe 40.

The surge tank 41 is connected through the inlet channel 43 to the air filter 44, while the inlet channel 43 is provided inside with a throttle valve 46 driven by the actuator 45 and an intake air volume sensor 47 using, for example, a high resistance wire. On the other hand, the exhaust port 39 is connected through the exhaust manifold 48 to a catalytic converter 4 containing, for example, a three-way catalyst, while the exhaust manifold 48 is provided internally with an air-fuel mixture quality sensor 49a.

On the other hand, in the embodiment shown in FIG. 6, the connecting part of the engine crankcase 30 and the cylinder block 31 is provided with a variable compression ratio mechanism A capable of changing the relative positions of the engine crankcase 30 and the cylinder block 2 in the axial direction of the cylinders so as to change the volume of the combustion chamber 34, when the piston 33 is located at the top dead center of compression, and is additionally equipped with a variable valve timing mechanism to control the closing timing of the intake valve 7 so as to control the amount of intake air actually supplied to the combustion chamber 34.

FIG. 7 is an exploded perspective view of a variable compression ratio mechanism A shown in FIG. 6, while FIG. 8 is a side cross-sectional view of an illustrated internal combustion engine. Referring to FIG. 7, at the bottom of two side walls of the cylinder block 31, a plurality of protruding parts 50 are formed, separated from each other by a certain distance. Each protruding portion 50 is formed with a circular cross-section hole 51 for inserting a cam. On the other hand, the upper surface of the crankcase 30 of the engine is formed with a plurality of protruding parts 52, separated from each other by a certain distance and installed between the respective protruding parts 50. These protruding parts 52 are also formed with round in cross section openings 53 for inserting the cam.

As shown in FIG. 7, a pair of cam shafts 54, 55 are provided. Each of the cam shafts 54, 55 has disk cams 56 fixed thereto, rotatably inserted into the cam insertion holes 51 in every second position. These disk cams 56 are coaxial with the axes of rotation of the cam shafts 54, 55. On the other hand, between the disk cams 56, as shown by shading in FIG. 8, the extending eccentric shafts 57 are eccentric with respect to the axes of rotation of the cam shafts 54, 55. Each the eccentric shaft 57 has other disk cams 58 rotatably attached to them eccentrically. As shown in FIG. 7, these disk cams 58 are interposed between the disk cams 56. These disk cams 58 are rotationally inserted into the respective cam insertion holes 53.

When the disk cams 56 attached to the cam shafts 54, 55 rotate in opposite directions, as shown by the solid line arrows in FIG. 8 (A), from the state shown in FIG. 8 (A), the eccentric shafts 57 move to a lower point thus, the disk cams 58 rotate in opposite directions from the disk cams 56 in the cam insertion holes 53, as shown by the broken line arrows in FIG. 8 (A). As shown in FIG. 8 (B), when the eccentric shafts 57 move to a lower point, the centers of the disk cams 58 move below the eccentric shafts 57.

As will be understood from the comparison of Figs. 8 (A) and 8 (B), the relative positions of the engine crankcase 30 and the cylinder block 31 are determined by the distance between the centers of the disk cams 56 and the centers of the disk cams 58. The greater the distance between the centers of the disk cams 56 and the centers of the disk Cams 58, the farther the cylinder block 31 from the crankcase 31 of the engine. If the cylinder block 31 moves away from the engine crankcase 30, the volume of the combustion chamber 34 when the piston 33 is located at the top dead center of compression increases, therefore, creating the rotation of the cam shafts 54, 55, the volume of the combustion chamber 34 when the piston 33 is located at the top dead center compression may vary.

As shown in FIG. 7, in order to rotate the cam shafts 54, 55 in opposite directions, the shaft of the drive motor 59 is provided with a pair of worm gears 61, 62 with opposite threads. Gears 63, 64 engaged with these worm gears 61, 62 are attached to the ends of the cam shafts 54, 55. In this embodiment, the drive motor 59 may be driven to vary the volume of the combustion chamber 34 when the piston 33 is at top dead center. compression over a wide range. It should be noted that the variable compression ratio mechanism A shown in FIGS. 6-8 is exemplary. Any type of variable compression ratio mechanism may be used.

On the other hand, FIG. 9 shows a valve timing control mechanism B attached to the end of a cam shaft 70 to drive the intake valve 36 in FIG. 6. Referring to Fig. 9, this valve timing control mechanism B is provided with a toothed pulley 71, rotated by the output shaft 9 of the engine 1 through the toothed belt of the drive in the direction of the arrow, a cylindrical housing 72 that rotates with the toothed pulley 71, the shaft 73 that can rotate with the cam the intake valve drive shaft 70 and rotate relative to the cylindrical body 72, a plurality of parts 74 extending from the inner circumference of the cylindrical body 72 to the outer circumference of the shaft 73, and vanes 75 extending between the parts 74 from the outer circumference of the shaft 73 to the inner circumference of the cylindrical body 72, the two sides of the blades 75 are formed with hydraulic chambers 76 for advancing and use the hydraulic chambers 77 for delay.

The oil-based fluid supply to the hydraulic chambers 76, 77 is controlled by an oil-based fluid control valve 78. This oil-based fluid control valve 78 is provided with fluid openings 79, 80 connected to hydraulic chambers 76, 77, an oil-based fluid supply port 82 discharged from the hydraulic pump 81, a pair of drain openings 83, 84 and a spool valve 85 to control the connection and disconnection of holes 79, 80, 82, 83, and 84.

In order to advance the cam phase of the camshaft of the intake valve actuator 70 in FIG. and the oil-based working fluid in the hydraulic chambers 77 for delay flows from the drain hole 84. At this time, the shaft 73 rotates relative to the cylindrical body 72 in the direction of the arrow.

In contrast, in order to delay the phase of the cams of the camshaft of the intake valve actuator 70, in FIG. 9, the spool valve 85 is made to move to the left, the oil-based working fluid supplied from the supply hole 82 is supplied through the hydraulic hole 80 to the hydraulic chambers 77 for delay and the oil-based working fluid in the hydraulic chambers 76 for advancing flows from the drain hole 83. At this time, the shaft 73 rotates relative to the cylindrical housing 72 in the opposite direction to the Cams.

When the shaft 73 rotates relative to the cylindrical body 72, if the spool valve 85 is returned to the neutral position shown in FIG. 9, the relative rotation operation of the shaft 73 ends and the shaft 73 is held in a relative rotating position at this time. Therefore, it is possible to use the camshaft adjuster B so that the cam phase of the camshaft 70 of the intake valve actuator is advanced or delayed by the exact desired amount.

In Fig. 10, a solid line shows when the camshaft adjuster B is used to advance the cam phase of the cam shaft 70 of the intake valve actuator in advance, while the broken line indicates when it is used to cam phase of the cams cam shaft 70 of the intake valve actuator. Therefore, the opening moment of the intake valve 36 can be freely set between the range shown by the solid line in FIG. 10 and the range shown by the dashed line, therefore, the closing time of the intake valve 36 can be set at any angle of rotation of the crankshaft in the range shown by the arrow C in FIG. 10.

The valve timing control mechanism B shown in FIGS. 6 and 9 is one example. For example, a valve timing control mechanism or other various types of valve timing control mechanisms that can only change the timing of closing the intake valve while maintaining a constant timing of opening the intake valve can be used.

Next, the meaning of the terms used in this application will be explained with reference to Fig.11. It should be noted that FIGS. 11 (A), 11 (B) and 11 (C) show, for explanatory purposes, an engine with a volume of combustion chambers of 50 ml and a cylinder displacement of 500 ml above the piston. In these Figures 11 (A), 11 (B) and 11 (C), the volume of the combustion chamber indicates the volume of the combustion chamber when the piston is at compression top dead center.

11 (A) illustrates the degree of mechanical compression. The degree of mechanical compression is a value determined mechanically from the working volume of the cylinder and the volume of the combustion chamber at the time of the compression stroke. This degree of mechanical compression is expressed by the value (volume of the combustion chamber + working volume) / volume of the combustion chamber. In the example shown in FIG. 11 (A), this mechanical compression ratio becomes (50 ml + 500 ml) / 50 ml = 11.

11 (B) illustrates the degree of actual compression. This degree of actual compression is a value determined from the actual working volume of the cylinder from the moment when the compression action actually began to the moment when the piston reaches top dead center and the volume of the combustion chamber. This actual compression ratio is expressed by the value (volume of the combustion chamber + actual working volume) / volume of the combustion chamber. That is, as shown in FIG. 11 (B), even if the piston starts to rise during compression, the compression action is not performed while the intake valve is open. The actual compression action begins after the intake valve closes. Therefore, the actual compression ratio is expressed as follows using the actual working volume. In the example shown in FIG. 11 (B), the actual compression ratio becomes (50 ml + 450 ml) / 50 ml = 10.

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

Next, with reference to Figs. 12 and 13, an ultrahigh expansion ratio cycle in the present invention will be explained. It should be noted that FIG. 12 shows the relationship between theoretical thermal efficiency and expansion ratio, while FIG. 13 shows a comparison between a conventional cycle and an ultra-high expansion cycle used selectively in accordance with the load in the present invention.

13 (A) shows a normal cycle when the intake valve closes near the bottom dead center and the compression action by the piston starts close essentially from the bottom dead center of the compression. In the example also shown in this FIG. 13 (A), in the same method as in the examples shown in FIGS. 11 (A), 11 (B) and 11 (C), the volume of the combustion chamber is 50 ml, and the working volume of the cylinder is 500 ml. As will be understood from FIG. 13 (A), in a typical cycle, the mechanical compression ratio is (50 ml + 500 ml) / 50 ml = 11, the actual compression ratio is also approximately 11, and the expansion ratio is also (50 ml + 500 ml ) / 50 ml = 11. That is, in a conventional internal combustion engine, the mechanical compression ratio and the actual compression ratio and the expansion ratio become substantially the same.

The solid line in FIG. 12 shows the change in theoretical thermal efficiency in the case where the actual compression ratio and the expansion ratio are substantially equal, that is, in a normal cycle. In this case, it was studied that the greater the degree of expansion, i.e. the higher the actual compression ratio, the higher the theoretical thermal efficiency. Therefore, in a normal cycle, in order to increase the theoretical thermal efficiency, the actual compression ratio should be higher. However, due to restrictions on the occurrence of detonation during engine operation under high load, the actual compression ratio can only be increased uniformly to a maximum of approximately 12, respectively, in a normal cycle, the theoretical thermal efficiency cannot be made sufficiently high.

On the other hand, in such a situation, a method is being studied to increase the theoretical thermal efficiency with a clear distinction between the degree of mechanical compression and the actual compression ratio, and as a result, it was found that in the theoretical thermal efficiency the degree of expansion is dominant, and the theoretical thermal efficiency is almost completely unaffected by the degree of actual compression. That is, if the actual compression ratio increases, the explosive force increases, but the compression requires a lot of energy, respectively, even if the actual compression ratio increases, the theoretical thermal efficiency will almost not increase.

In contrast, if the degree of expansion increases, the longer the period during which the force acts as the force pressing the piston at the time of the expansion stroke, the longer the time during which the piston transfers the rotational force to the crankshaft. Therefore, the greater the degree of expansion, the higher the theoretical thermal efficiency becomes. The dashed line in FIG. 12 shows the theoretical thermal efficiency in the case of fixing the actual compression ratio at 5, 6, 7, 8, 9, 10, respectively, and increasing the expansion ratio in this state. Note that in Fig. 12, black dots show the peak positions of the theoretical thermal efficiency when the actual compression ratios are 5, 6, 7, 8, 9, 10. It is clear from Fig. 12 that the growth rate of the theoretical thermal efficiency with increasing degree of expansion in the state where the actual compression ratio ε is kept at a low value, and the theoretical thermal efficiency growth rate when the actual compression ratio increases together with the expansion ratio, as shown by the solid line in FIG. 12, will hardly differ.

If the actual compression ratio ε is kept at a low value in this method, knocking will not occur, therefore, if the expansion ratio is increased in a state where the actual compression ratio is kept at a low value, the occurrence of detonation can be prevented, and the theoretical thermal efficiency can be significantly increased. 13 (B) shows an example of a case where a variable compression ratio mechanism A and a variable valve timing mechanism B are used to maintain the actual compression ratio at a low value and increase the expansion ratio.

Referring to FIG. 13 (B), in this example, a variable compression ratio mechanism A is used to reduce the volume of the combustion chamber from 50 ml to 20 ml. On the other hand, valve timing control mechanism B is used to delay the closing timing of the intake valve until the actual piston stroke has changed from 500 ml to 200 ml. As a result, in this example, the actual compression ratio is (20 ml + 200 ml) / 20 ml = 11, and the expansion ratio is (20 ml + 500 ml) / 20 ml = 26. In the conventional cycle shown in FIG. 13 (A), as explained above, the actual compression ratio is approximately 11 and the expansion ratio is 11. Compared to this case in the case shown in FIG. 13 (B), it has been studied that only the degree of expansion rises to 26. For this reason, the cycle is called the "cycle of superhigh degree of expansion".

As explained above, if the degree of expansion increases, the theoretical thermal efficiency improves and fuel consumption improves. Therefore, the degree of expansion preferably grows in as wide a work area as possible. However, as shown in FIG. 13 (B), in the super-high expansion ratio cycle, since the actual piston stroke at the time of the compression stroke is made smaller, the amount of intake air supplied to the combustion chamber 34 becomes smaller. Therefore, this super-high expansion ratio cycle can only be applied when the amount of intake air supplied to the combustion chamber 34 is small, that is, when the required engine torque Te is low. Therefore, in the embodiment according to the present invention, when the required engine torque Te is low, the super-high expansion ratio cycle shown in FIG. 13 (B) is applied, while the required engine torque Te is high, the normal cycle shown in FIG. Fig. 13 (A).

Next, referring to FIG. 14, it is explained how the engine 1 is controlled in accordance with the required torque Te.

14 shows changes in the degree of mechanical compression, the degree of expansion, the closing timing of the intake valve 36, the actual compression ratio, the amount of intake air, the opening degree of the throttle valve 46 and the fuel consumption level in accordance with the required engine torque Te. The fuel consumption level shows the fuel consumption when the vehicle travels a predetermined travel distance in a predetermined driving mode. Therefore, the value indicating the level of fuel consumption becomes lower, the better the level of fuel consumption. It should be noted that in the embodiment according to the present invention, usually the average air-fuel ratio in the combustion chamber 34 is feedback controlled based on the output of the air-fuel ratio sensor 49a to the stoichiometric air-fuel ratio, so that the three-way catalyst in the catalyst 49 The exhaust gas can simultaneously reduce unburned CH, CO and NO _X in the exhaust gas. 12 shows theoretical thermal efficiency when the average air-fuel ratio in the combustion chamber 34 is equal to the stoichiometric air-fuel ratio in this way.

On the other hand, in this method, in the embodiment according to the invention, the average air-fuel ratio in the combustion chamber 34 is controlled to a stoichiometric air-fuel ratio, so that the engine torque Te becomes proportional to the amount of intake air supplied to the combustion chamber 34. Therefore, as shown in FIG. 14, the more the required engine torque Te falls, the more the intake air volume decreases. Therefore, in order to further reduce the intake air volume when the required engine torque Te falls, as shown by the solid line in FIG. 14, the closing timing of the intake valve 36 is delayed. The throttle valve 46 is held fully open while the intake air volume is controlled by delaying the closing moment of the throttle valve 36 in this manner. On the other hand, if the required engine torque Te becomes lower than the determined value Te ₁ , it is no longer possible to control the intake air volume to the required intake air volume by controlling the closing timing of the intake valve 36. Therefore, when the required engine torque Te is lower than this the value of Te ₁ , the limit value of Te ₁ , the closing moment of the intake valve 36 is held at the limit moment of closure during the limit value of Te ₁ . At this time, the intake air volume is controlled by the throttle valve 46.

On the other hand, as explained above, when the required engine torque Te is low, an ultra-high expansion ratio cycle is applied, therefore, as shown in FIG. 14, when the required engine torque Te is low, the mechanical compression ratio increases, whereby the expansion ratio becomes higher . In this regard, as shown in FIG. 12, when, for example, the actual compression ratio ε is 10, the theoretical thermal efficiency reaches a peak when the expansion ratio is 35 or so. Therefore, when the required engine torque Te is low, it is preferable to increase the mechanical compression ratio until the expansion ratio is 35 or so. However, it is difficult to increase the mechanical compression ratio until the expansion ratio is 35 or so, due to structural limitations. Therefore, in the embodiment according to the present invention, when the required engine torque Te is low, the mechanical compression ratio becomes the maximum structurally possible mechanical compression ratio, so that a high expansion ratio is obtained as much as possible.

On the other hand, if the closing moment of the intake valve 36 is advanced in such a way that the volume of intake air increases in a state that maintains the degree of mechanical compression at the maximum degree of mechanical compression, the degree of actual compression becomes higher. However, the actual compression ratio should be kept at 12 or less even at maximum. Therefore, when the required engine torque Te becomes high and the intake air volume increases, the mechanical compression ratio is lowered so that the actual compression ratio is maintained at the optimum actual compression ratio. In the embodiment according to the present invention, as shown in FIG. 14, when the required engine torque Te exceeds the limit value Te ₂ , the mechanical compression ratio decreases when the required engine torque Te increases so that the actual compression ratio is maintained at the optimum actual degree compression.

If the required engine torque Te becomes higher, the mechanical compression ratio is reduced to a minimum mechanical compression ratio. At this time, the cycle becomes the normal cycle shown in FIG. 13 (A).

In this regard, in the embodiment according to the present invention, when the engine rotation speed Ne is low, the actual compression ratio ε is 9-11. However, if the engine rotation speed Ne becomes higher, the air-fuel mixture in the combustion chamber 34 is disturbed, thus, knocking is not so easy. Therefore, in the embodiment according to the present invention, the higher the engine rotation speed Ne, the higher the actual compression ratio ε.

On the other hand, in the embodiment according to the present invention, the expansion ratio when the superhigh expansion ratio cycle is used is 26-30. On the other hand, in FIG. 12, the actual compression ratio ε = 5 shows the lower limit of the practicable actual compression ratio. In this case, the theoretical thermal efficiency reaches a peak when the degree of expansion is approximately 20. The degree of expansion, where the theoretical air-fuel ratio reaches a peak, becomes greater than 20 when the degree ε of actual expansion becomes greater than 5. Therefore, if we consider the practicable degree ε of actual compression, it can be said that the expansion ratio is preferably equal to 20 or more. Therefore, in the embodiment according to the present invention, the variable compression ratio mechanism A is formed so that the expansion ratio becomes 20 or more.

In addition, in the example shown in FIG. 14, the mechanical compression ratio continuously changes in accordance with the required engine torque Te. However, the mechanical compression ratio can be changed in stages in accordance with the required engine torque Te.

On the other hand, as shown by the dashed line in FIG. 14, when the required engine torque Te becomes lower, it is possible to control the intake air volume by shifting in advance the closing timing of the intake valve 36. Therefore, if the invention is described so as to include it as the case shown by the solid line and the case shown by the dashed line in FIG. 14, in the embodiment according to the present invention, the closing timing of the intake valve 36 moves away from the bottom dead center and an intake BDC to a closing limit point capable of controlling the amount of intake air supplied to the combustion chamber 34 when the required engine torque Te becomes lower.

In this regard, if the expansion ratio becomes higher, the theoretical thermal efficiency becomes higher and the fuel consumption becomes better, that is, the fuel consumption level becomes lower. Therefore, in FIG. 14, when the required engine torque Te is the limit value Te ₂ or less, the fuel consumption level becomes the smallest. However, between the limit value of Te ₁ and Te _{2, the} actual compression ratio drops when the required engine torque Te becomes lower, thus, the fuel consumption level deteriorates quite a bit, that is, the fuel consumption level becomes higher. Further, in an area where the required engine torque Te is lower than the limit value Te ₁ , the throttle valve 46 is closed, so that the fuel consumption level becomes further higher. On the other hand, if the required engine torque Te becomes higher than the limit value Te ₂ , the expansion ratio drops, thus, the fuel consumption level increases when the required torque Te becomes higher. Therefore, when the required torque Te is the limit value Te ₂ , that is, at the boundary of the region where the mechanical compression ratio Te is decreased by increasing the required engine torque Te and the region where the mechanical compression ratio is maintained at the maximum mechanical compression ratio, the fuel consumption level becomes the smallest.

The limit value Te _{2 of} the engine torque Te, where the fuel consumption becomes the smallest, partly changes in accordance with the engine speed Ne, but in any case, if it is possible to maintain the engine torque Te at the limit value Te ₂ , the minimum fuel consumption is achieved. In the present invention, the output power control system 2 is used to maintain the engine torque Te at the limit value Te ₂ , even if the required output power Pe of the engine 1 is changed.

Next, referring to Fig. 15, a method for controlling the engine 1 will be explained.

Fig. 15 shows equivalent lines a ₁ , a ₂ , a ₃ , a ₄ , a ₅ , a ₆ , a ₇ and a _{8 of} the fuel consumption level expressed in two-dimensional coordinates with an ordinate equal to the engine torque Te and an abscissa equal to engine speed Ne. The equivalent fuel consumption level lines a ₁ -a ₈ are equivalent fuel consumption level lines obtained by controlling the engine 1 shown in FIG. 6, as shown in FIG. The greater a ₁ -a ₈ , the higher the fuel consumption level. That is, inside a ₁ there is a region of the lowest level of fuel consumption. Point O ₁ shown in the inner region a ₁ is an operating state giving the lowest fuel consumption level. In the engine 1 shown in FIG. 6, the point O ₁ where the fuel consumption level becomes minimal is the case when the engine torque Te is low and the engine rotation speed Ne is approximately 2000 rpm.

In Fig. 15, the solid line K1 shows the ratio of the engine torque Te and the engine speed Ne, where the engine torque Te becomes the limit value Te ₂ shown in Fig. 14, that is, where the fuel consumption level becomes minimal. Therefore, when setting the engine torque Te and the engine speed Ne equal to the engine torque Te and the engine speed Ne on the solid line K1, the fuel consumption level becomes minimal. Therefore, the solid line K1 is called the “minimum fuel consumption line”. This minimum fuel consumption line K1 takes the form of a curve passing through point O ₁ in the direction of increasing the engine rotation speed Ne.

As will be understood from FIG. 15, on the line with the minimum fuel consumption level K1, the engine torque Te does not change at all. Therefore, when the required output power Pe of the engine 1 increases, the required output power Pe of the engine 1 is satisfied by increasing the rotation speed Ne of the engine. On this line K1 of operation with a minimum level of fuel consumption, the degree of mechanical compression is fixed at the maximum degree of mechanical compression. The closing moment of the intake valve 36 is also fixed at the moment giving the required intake air volume.

Depending on the engine design, it is possible to set this line K1 of operation with a minimum fuel consumption level so that it runs directly in the direction of increasing the engine rotation speed Ne until the engine rotation speed Ne becomes maximum. However, when the engine rotation speed Ne becomes high, the losses due to the increase in friction become larger. Therefore, in the engine 1 shown in FIG. 6, when the required output power Pe of the engine 1 increases compared to when the degree of mechanical compression is maintained at the maximum degree of mechanical compression, and in this state only the rotation speed Ne of the engine increases when the torque is the engine moment Te increases along with an increase in the engine rotation speed Ne, a drop in the mechanical compression ratio causes the theoretical thermal efficiency to fall, but the practical thermal efficiency increases. That is, in the engine 1 shown in FIG. 6, when the engine speed Ne becomes high, the fuel consumption becomes smaller when the engine speed Ne and the engine torque Te increase than when only the engine speed Ne increases.

Therefore, in the embodiment according to the present invention, the minimum fuel consumption line K1, as shown by K1 ′ in FIG. 15, extends toward the high engine torque Te along with the increase in engine speed Ne if the engine speed Ne becomes higher . On this line K1 ′ of operation with a minimum level of fuel consumption, the further from the line K1 of operation with a minimum level of fuel consumption, the closer the closing moment of the intake valve 36 is to the bottom dead center of the inlet, and the degree of mechanical compression decreases more from the maximum degree of mechanical compression.

Now, as explained above, in the embodiment according to the present invention, the ratio of the engine torque Te and the engine speed Ne when the fuel consumption becomes minimal, if expressed in two-dimensional coordinates as a function of the engine torque Te and the engine speed Ne, is expressed as line K1 work with a minimum level of fuel consumption, forming a curve passing in the direction of increasing the speed Ne of rotation of the engine. In order to minimize the fuel consumption level, while it is possible to satisfy the required output power Pe of the engine 1, it is preferable to change the engine torque Te and the engine speed Ne along this line K1 with a minimum fuel consumption level.

Therefore, in the embodiment according to the present invention, while the required output power Pe of the engine 1 can be satisfied, the engine torque Te and the engine rotation speed Ne change along the operation line K1 with a minimum fuel consumption level in accordance with the change in the required output power Pe of the engine. It should be noted that, as one would expect, this line K1 of operation with a minimum level of fuel consumption by itself is not stored in advance in the ROM 22. The ratios of the engine torque Te and the engine speed Ne, showing the lines of operation K1 and K1 'with a minimum level the fuel consumption are stored in advance in the ROM 22. Additionally, in the embodiment according to the present invention, the engine torque Te and the engine speed Ne vary in the range of the operation line K1 with a minimum fuel consumption level along the line and K1 of operation with a minimum level of fuel consumption, but the range of variation of the engine torque Te and the engine speed Ne can also be expanded to the line K1 'of operation with a minimum level of fuel consumption.

Next, operation lines other than the operation lines K1 and K1 ′ with a minimum fuel consumption level will be explained.

Turning to FIG. 15, when expressed in two-dimensional coordinates as a function of engine torque Te and engine speed Ne, the high-torque operation line shown by the dashed line K2 is set from the high-torque side Te of the engine operation lines K1 and K1 ′ minimum fuel consumption. In fact, the ratio of the engine torque Te and the engine speed Ne showing this high torque line K2 is determined in advance. This ratio is stored in advance in the ROM 22.

Next, this high torque operation line K2 will be explained with reference to FIG. 17 shows the equivalent lines b ₁ , b ₂ , b ₃ , and b _{4 of} the fuel consumption level expressed in two-dimensional coordinates with an ordinate equal to the engine torque Te and an abscissa equal to the engine rotation speed Ne. The equivalent fuel consumption level lines b ₁ -b ₄ show the fuel consumption level lines when the engine 1 shown in FIG. 6 is operating in a state of decreasing the mechanical compression ratio to the lowest value in the engine 1, i.e. in the case of the normal cycle shown in FIG. 13 (A). From b ₁ to b ₄ fuel consumption becomes higher. That is, inside b ₁ there is a region of the lowest level of fuel consumption. The point shown as O ₂ inside region b ₁ becomes the operating state of the lowest fuel consumption level. In the engine 1 shown in FIG. 17, the point O ₂ where the fuel consumption level becomes minimal is the case when the engine torque Te is high and the engine rotation speed Ne is approximately 2400 rpm.

In an embodiment according to the present invention, the high-torque line K2 is defined by a curve where the fuel consumption level becomes minimum when the engine 1 is operating in a state where the mechanical compression ratio is reduced to a minimum.

Referring again to FIG. 15, when expressed in two-dimensional coordinates as a function of engine torque Te and engine rotation speed Ne, the full-load line K3 along which full-load operation is performed is set further in the direction of the higher torque from the K2 line work with high torque. The relationship between the engine torque Te and the engine speed Ne, showing this full load operation line K3, is predetermined. This ratio is stored in advance in the ROM 22.

Fig. 16 (A) and (B) show a change in the fuel consumption level and a change in the mechanical compression ratio when viewed along the line ff in Fig. 15. As shown in FIG. 16, the fuel consumption level becomes minimal at the point O ₁ on the operation line K1 with the minimum consumption level and becomes higher towards the point O ₂ on the high torque line K2. Additionally, the degree of mechanical compression becomes maximum at point O ₁ on line K1 of operation with a minimum level of fuel consumption and gradually decreases towards point O ₂ . Additionally, the intake air volume becomes larger the higher the engine torque Te, thus the intake air volume increases from the point O ₁ on the minimum flow rate line K1 to the point O ₂ , while the closing timing of the intake valve 36 approaches to the bottom inlet dead center along with the movement from point O ₁ to point O ₂ .

Fig. 18 shows the equivalent output power lines Pe ₁ -Pe ₉ and the operation lines K1, K2 and K3 of the engine 1 shown in Fig. 3 (A), as well as two boundary output powers, consisting of a first boundary output power PY and a second boundary power output PZ. As will be understood from FIG. 18, the equivalent line of the first boundary output power PY passes through the intersection of the lines of operation K1 and K1 'with a minimum level of fuel consumption, while the equivalent line of the second boundary output power PZ passes so as to be in contact with the line K2 of operation with high torque.

Now, as explained above, in the embodiment according to the present invention, when the required output power Pe of the engine 1 increases while it is possible to satisfy the required output power Pe of the engine 1, the engine torque Te and the engine speed Ne change along the operation line K1 with a minimum flow rate fuel. That is, in the embodiment according to the present invention, when the required output power Pe of the engine 1 is lower than the predetermined first boundary output power PY, control is performed to maintain a minimum fuel consumption level satisfying the required output power of the engine by changing the engine speed in the state supporting a mechanical compression ratio at a predetermined compression ratio or more.

In contrast, when the required output power Pe of the engine 1 is not satisfied by the engine torque Te and the engine rotation speed Ne along the operation line K1 with a minimum fuel consumption level, i.e. the required engine output increases above the first boundary output power PY, control is performed to increase the output power, increasing the engine torque Te after lowering the mechanical compression ratio to a predetermined compression ratio, i.e., 20 or less.

This control to increase the output power is performed by controlling the closing timing of the intake valve 36 to increase the amount of intake air to the combustion chamber 34 and, therefore, change the engine torque Te and the engine rotation speed Ne from the points on the operation line K1 with a minimum level fuel consumption in the direction of increasing torque Te.

Next, referring to Figs. 19-28, two embodiments will be explained showing a method for controlling engine torque Te and engine speed Ne. It should be noted that FIGS. 18, 19, 21, 22, 25 and 26 show equivalent lines Pe ₁ -Pe ₉ of the engine output power, operation lines K1, K2 and K3 and boundary output powers PY and PZ, which are the same as and Fig. 18.

FIG. 19 shows a case where when the output power of engine 1 is Pe ₂ and it is in the operating state shown by point R on the minimum fuel consumption line K1, the required output power Pe of engine 1 becomes Pe ₄ . In this case, said control is performed to maintain a minimum level of fuel consumption. That is, in accordance with a change in the required engine power output Pe, the engine torque Te and the engine speed Ne change, as shown by the arrow, along the operation line K1 with a minimum fuel consumption level from point R to point Pe.

It should be noted that at this time, in reality, the required engine torque TeX and the required engine rotation speed NeX are detected at several points on the operation line K1 with a minimum fuel consumption level from point R to point Pe, and the required engine torque TeX and the required the engine rotation speed NeX is sequentially set from the required engine torques TeX and the required engine rotational speeds NeX so that the engine torque Te and the engine rotation speed Ne vary along l K1 lines with minimum fuel consumption from point R to point Pe.

On the other hand, FIG. 19 also shows a case where when the output power of engine 1 is Pe ₄ and it is in the operating state shown by the point Pe on the operation line K1 with a minimum fuel consumption, the required output power of the engine 1 becomes Pe ₂ . In this case, the aforementioned control is also performed to maintain a minimum level of fuel consumption. That is, the engine torque Te and the engine speed Ne change, as shown by the arrow, along the operation line K1 with a minimum fuel consumption level from the point Pe to the point R.

FIG. 20 shows changes in the degree of mechanical compression, the closing timing of the intake valve 36, the torque Te, and the engine rotation engine Ne when such control is performed to maintain a minimum fuel consumption level. It should be noted that, in FIG. 20, MAX shows the maximum degree of mechanical compression, while MIN shows the minimum degree of mechanical compression. Additionally, in FIG. 20, the time t ₁ shows the point in time when the required engine output changes from Pe ₂ to Pe, while the time t ₂ shows the time when the required engine output changes from Pe to Pe ₂ .

As will be understood from FIG. 20, when control is performed to maintain a minimum fuel consumption level, even if the required engine output is changed, the mechanical compression ratio and the closing timing of the intake valve 36 will not change. In contrast, when the required engine power output changes from Pe ₂ to Pe, the engine speed Ne increases, while when the required engine power output changes from Pe to Pe ₂ , the engine speed Ne decreases. At this time, the engine torque Te hardly changes at all.

FIG. 21 shows a case where, when the output power of engine 1 is Pe ₂ , and it is in the operating state shown by point R on line K1 of operation with a minimum fuel consumption level, an acceleration operation is performed, and the required output power Pe of engine 1 becomes Pe ₇ . In this case, the required output power Pe of the engine 1 is higher than the first boundary output power PY, thus, control is performed to increase the output power. That is, the mechanical compression ratio decreases first, then the engine torque Te and the engine rotation speed Ne increase from the point R on the operation line K1 with the minimum fuel consumption to the point Pe on the high-torque operation line K2, where the engine output becomes the required output power Pe ₇ .

In this case, the required engine torque TeX and the required engine speed NeX are located at several points on a straight line R-Pe connecting the point R on the line K1 with a minimum fuel consumption and the point Pe on the line K2 with high torque, and the required engine torque TeX and the required engine speed NeX are sequentially set from the required required engine torques TeX and the required engine speeds NeX so that the engine torque Te and the rotational speed Ne Nia engine vary along the R-Pe straight line from the point R to the point Pe.

On the other hand, in the state shown by the point on the high-torque line K2, if the required engine output becomes Pe ₂ , as shown in FIG. 21, the engine torque Te and the engine speed Ne decrease from the point Pe on the line K2 work with high torque along a straight line R-Pe to point R on line K1 work with a minimum level of fuel consumption, where the output of the engine becomes the required output power Pe ₂ .

FIG. 22 shows a case where, when the output power of engine 1 is Pe ₂ , and it is in the operating state indicated by point R on the minimum fuel consumption line K1, an acceleration operation is performed, and the required output power Pe of engine 1 exceeds the second boundary output power PZ, in particular the case where the required output power Pe of the engine 1 becomes the maximum output power. In this case, the mechanical compression ratio decreases first, then the engine torque Te and the engine rotation speed Ne increase from the point R on the operation line K1 with the minimum fuel consumption to the point Pe on the operation line K3 with full load, where the engine output becomes the required output power Pe, along a straight line R-Pe connecting the point R on line K1 with a minimum fuel consumption level to point Pe on line K3 with full load.

On the other hand, in the state shown by the point Pe on the full load line K3, if the required engine output becomes Pe ₂ , as shown in FIG. 22, the engine torque Te and the engine speed Ne decrease from the point Pe on the line K3 full load operation up to point R on line K1 with minimum fuel consumption, where the engine output becomes the required output Pe ₂ , along a straight R-Pe line.

FIG. 23 shows changes in the mechanical compression ratio, the closing timing of the intake valve 36, the engine torque Te, and the engine rotation speed Ne when the control for increasing the output power shown in FIGS. 21 and 22 is executed. It should be noted that in FIG. 23 also, moment t ₁ shows the time when the required engine output changes from Pe ₂ to Pe, while time t ₂ shows the time when the required engine output changes from Pe to Pe ₂ .

As shown in FIG. 23, when the required engine power output changes from Pe ₂ to Pe, the mechanical compression ratio is first reduced. After the action of reducing the mechanical compression ratio is completed, the closing moment of the intake valve 34 is changed so as to approach the bottom dead center of the intake, that is, the amount of intake air increases, and therefore the engine torque Te and the engine speed Ne increase. In this case, in reality, the target mechanical compression ratio for the time when the engine torque Te and the engine rotation speed Ne are on the high torque operation line K2 is predetermined. When control is performed to increase the output power, the mechanical compression ratio is reduced to the target mechanical compression ratio, then the engine torque Te and the engine rotation speed Ne are changed from the values on the operation line K1 with a minimum fuel consumption to the values on the high-torque operation line K2 .

As explained above, in the embodiment according to the present invention, the high torque line K2 is a curve where the fuel consumption level becomes minimal when the engine is running in a state where the mechanical compression ratio is reduced to a minimum level. Therefore, in the embodiment according to the present invention, the target mechanical compression ratio becomes the minimum mechanical compression ratio. Therefore, in the embodiment according to the present invention, when control is performed to increase the output power, the mechanical compression ratio is reduced from the maximum mechanical compression ratio MAX to the minimum mechanical compression ratio.

Reducing the degree of mechanical compression takes some time. Therefore, if the closing timing of the intake valve 34 approaches the bottom dead center of the intake to increase the intake air volume immediately when the required engine output increases, the effect of reducing the mechanical compression ratio cannot continue, thus, the actual compression ratio will be extremely high. The result is detonation. Therefore, in the present invention, in this method, in order to prevent the occurrence of detonation, the mechanical compression ratio is first reduced when the required engine output is increased.

If the mechanical compression ratio drops, the actual compression ratio drops. At this time, in order to prevent the actual compression ratio from dropping too much, in the embodiment according to the present invention, as shown in FIG. 2, when the mechanical compression ratio drops, the closing timing of the intake valve 34 gradually approaches the bottom dead center of the intake. When the mechanical compression ratio becomes minimal, the mechanical compression ratio MIN, the closing timing of the intake valve 34 quickly approaches the intake dead center in order to increase the engine torque Te and the engine rotation speed Ne. It should be noted that when the action to increase the requested output power required from the vehicle during the acceleration operation is not covered by the action to increase the engine power output, the lack of engine power output is compensated by the increase in the output power of the motor generator MG1.

On the other hand, when the required engine power output changes from Pe to Pe ₂ at time t ₂ , the closing timing of the intake valve 34 moves away from the bottom dead center of the intake, and the engine torque Te and the engine speed Ne decrease. Further, when the engine power output becomes Pe ₂ , the mechanical compression ratio is increased from the minimum mechanical compression ratio MIN to the maximum mechanical compression ratio MAX. The closing timing of the intake valve 34 moves away from the bottom dead center of the intake, then the mechanical compression ratio is increased in this way so as to prevent knocking.

After the control to increase the output power, such as that shown in FIG. 21, is performed, and the engine torque Te and the engine speed Ne reach the values on the high torque line K2, the engine torque Te and the engine speed Ne change along the high torque operation line K2, while the required engine output is the first boundary output power PY or more.

Further, after a control for increasing the output power, such as that shown in FIG. 22, is performed, and the engine torque Te and the engine speed Ne reach the values on the full load line K3, the engine torque Te and the engine speed Ne vary along the high torque line K3, while the required engine output is the second boundary output PZ or more.

That is, in the embodiment according to the present invention, the engine torque Te and the engine rotation speed Ne are usually stored in the values on the operation lines K1, K2 and K3. Except control time to increase power output, etc. the engine torque Te and the engine speed Ne are not stored between the operation lines K1 and K2 or the operation lines K2 and K3. For example, if the engine torque Te and the engine speed Ne are maintained at any value between the operation lines K1 and K2, the mechanical compression ratio often changes, and the durability of the variable compression ratio mechanism A becomes a problem.

As opposed to this, as in the embodiment of the present invention, if the engine torque Te and the engine speed Ne are not stored between the operation lines K1, K2 and K3, but they are stored on the operation lines K1, K2 and K3, the rate of change of the mechanical compression ratio significantly reduced, and therefore, the durability of mechanism A of variable compression ratio can be ensured.

Fig. 24 shows an operational control procedure for executing the first embodiment shown in Figs. 19-23. This procedure is performed by interruption at predefined time intervals.

Referring to FIG. 24, first, at step 200, it is judged whether, during a previous interruption, the required output power Pe of the engine was really lower than the first boundary output power PY. When during the previous interruption Pe <PY, the procedure proceeds to step 201, where it is assessed whether during the current interruption Pe <PY. When during the current interruption Pe <PY, that is, when the required output power Pe of the engine continues to be lower than the first boundary output power PY, the procedure proceeds to step 202, where the engine torque Te and the engine speed Ne are controlled along the line K1 of operation with maintaining fuel consumption.

On the other hand, when it is estimated in step 200 that Pe was not less than PY during the previous interruption, the routine proceeds to step 203 where it is evaluated whether Pe <PY during the current interruption. When Pe is not less than PY during the current interruption, that is, when the required output motor power Pe continues to be higher than the first boundary output power PY, the procedure proceeds to step 204, where it is evaluated whether the required output motor Pe is lower than the second boundary output power PZ. When Pe <PZ, the procedure proceeds to step 205, where the engine torque Te and the engine speed Ne are controlled along the high torque line K2. In contrast, when it is estimated that Pe> PZ in step 204, the procedure proceeds to step 206, where the engine torque Te and the engine speed Ne are controlled along the full load line K3.

On the other hand, when it is estimated at step 201 that Pe is not less than PY during the current interruption, that is, when the required engine output exceeds the first boundary output power PY, the procedure jumps to step 207. At step 207, it is judged if the degree reduction process is completed mechanical compression. When the process of lowering the degree of mechanical compression is not completed, the procedure proceeds to step 208, where the degree of mechanical compression is reduced. Further, when it is estimated at step 207 that the process of lowering the mechanical compression ratio is completed, the procedure proceeds to step 209, where control is performed to increase the output power.

On the other hand, when it is estimated at step 203 that Pe <PY during the current interruption, that is, when the desired engine output becomes lower than the first boundary output power PY, the procedure jumps to step 210. At step 210, it is judged whether the control is completed to lower the engine power output. When the control for lowering the engine power output is not completed, the procedure proceeds to step 211, where control is performed for lowering the engine power output. Further, when it is judged in step 210 that the control for lowering the engine output is completed, the procedure proceeds to step 212, where processing is performed to increase the degree of mechanical compression.

25-28 show a second embodiment. In this embodiment, as shown in FIG. 25 and FIG. 27, when the output power of engine 1 is Pe ₂ and it is in the operating state shown by point R on line K1 of operation with a minimum fuel consumption level, if the required output power The engine Pe exceeds the first boundary output power PY, first the engine rotation speed Ne increases along the operation line K1 with a minimum fuel consumption to the engine speed on the high-torque operation line K2 satisfying the required output power and Pe, then the mechanical compression ratio is reduced from the maximum mechanical compression MAX to MIN minimum mechanical compression, further control to increase the output power is performed in a state that preserves the engine rotation speed Ne constant.

That is, in this case, the target engine torque and the target engine speed are first found on the high-torque line K2 satisfying the required output power Pe, then the engine speed Ne becomes the target engine speed, then the mechanical compression ratio is reduced to the target degree mechanical compression, then the engine torque Te increases from the value on the line K1 with a minimum fuel consumption to the target torque on the line K2 slave you with high torque.

In contrast, when the engine is operational, where the output of engine 1 is indicated by a point Pe on the high-torque line K2, if the required output of the engine Pe becomes lower than the first boundary output PY, the engine rotation speed Ne is first maintained constant, and in this state control is performed to reduce the output power, then the degree of mechanical compression increases from the minimum degree of mechanical MIN compression to the maximum degree of mechanical MAX-ethnic compression, further rotation speed Ne of the engine is reduced.

On the other hand, as shown in FIG. 26, when the output power of engine 1 is Pe ₂ , and it is in the operating state shown by point R on line K1 of operation with a minimum fuel consumption level, if the required output power Pe of the engine exceeds the second boundary output power PZ in the same manner as in the case of FIG. 25, as shown in FIG. 27, first, the engine speed Ne increases along the operation line K1 with a minimum fuel consumption to the engine speed on the full load operation line K3, satisfy the required output power Pe, then the degree of mechanical compression decreases from the maximum degree of mechanical MAX compression to the minimum degree of mechanical MIN compression, then the rotation speed Ne of the engine is kept constant, and in this state control is performed to increase the output power.

In contrast, when the output power of the engine 1 is the operating state shown by the point Pe on the full load line K3, and when the required output power Pe of the engine becomes lower than the first boundary output power PY, the control is first performed to reduce the output power in the state keeping the engine speed Ne constant. Further, the degree of mechanical compression is increased from the minimum degree of mechanical MIN compression to the maximum degree of mechanical MAX compression, further, the engine rotation speed Ne decreases.

When the engine torque Te and the engine speed Ne are controlled before the operation line K1 with a minimum fuel consumption level, if the engine speed Ne is increased, the engine output is increased. On the other hand, as explained above, the engine rotation speed Ne is controlled by the motor generator MG1. Controlling the engine speed Ne by this motor generator MG1 has a very good response. Therefore, as shown in FIG. 25 or FIG. 26, if the engine rotation speed Ne increases when an acceleration request exists, the output of the engine immediately increases, and therefore, a good acceleration sensation can be obtained.

Fig. 28 shows an operational control procedure for executing the second embodiment shown in Figs. 25-27. This procedure is also performed by interruption at each constant time interval. It should be noted that in this procedure, steps 300-306 are the same as steps 200-206 of the procedure shown in FIG. 24, so an explanation of steps 300-306 will be omitted.

Referring to FIG. 28, in this second embodiment, when it is estimated in step 301 that Pe is not less than PY during the current interruption, that is, when the required motor output exceeds the first boundary output PY, the procedure skips to step 307. At step 307, it is judged whether the process of increasing the engine speed Ne is completed. When the process of increasing the engine speed is not completed, the procedure proceeds to step 308, where the engine speed Ne is increased. Further, when it is estimated at step 307 that the process of increasing the engine speed Ne is completed, the procedure proceeds to step 309.

At step 309, it is evaluated whether the process of lowering the degree of mechanical compression is completed. When the process of lowering the degree of mechanical compression is not completed, the procedure proceeds to step 310, where the degree of mechanical compression is reduced. Further, when it is estimated at step 309 that the process of lowering the mechanical compression ratio is completed, the procedure proceeds to step 311, where control is performed to increase the output power.

On the other hand, when it is estimated at step 303 that Pe <PY during the current interrupt, i.e. when the desired engine output becomes lower than the first marginal output PY, the procedure jumps to step 312. At step 312, it is judged if control has been completed to reduce the engine output. When the control for lowering the engine power output is not completed, the procedure proceeds to step 313, where control is performed for lowering the engine power output. Further, when it is estimated at step 312 that the control for lowering the engine output is completed, the procedure proceeds to step 314.

At 314, it is judged if the process of increasing the mechanical compression ratio has been completed. When the process of increasing the mechanical compression ratio is not completed, the process proceeds to step 315, where the mechanical compression ratio is increased. Further, when it is estimated at step 314 that the process of increasing the mechanical compression ratio is completed, the procedure proceeds to step 316, where the engine speed Ne decreases.

Claims (9)

1. The engine control system, equipped with an output power control system that allows you to specify the desired combination of engine torque and engine speed, providing the same output power, while there is a variable compression ratio mechanism that can change the mechanical compression ratio, and a valve timing control mechanism capable of control when the intake valve closes and when the required engine output is lower than the predetermined boundary output power, control is performed to maintain a minimum level of fuel consumption, satisfying the required output power, by changing the engine speed in a state that maintains a mechanical compression ratio by a predetermined compression ratio or more, and when the required engine output power increases above the boundary output power, control to increase power output, increasing engine torque after decreasing the degree of mechanical compression Ia to a predetermined degree of compression or less. 2. The system according to claim 1, in which control to increase the output power is performed by controlling the moment of closing the intake valve to increase the amount of intake air into the combustion chamber. 3. The system of claim 1, wherein said predetermined compression ratio is 20. 4. The system according to claim 1, in which the ratio between the engine torque and engine speed when fuel consumption becomes minimum is expressed as a line of operation with a minimum fuel consumption level, having the shape of a curve extending in the direction of increasing engine speed when expressed in two-dimensional coordinates as a function of engine torque and engine speed, and when control is performed to maintain a minimum level of fuel consumption, engine torque and rotational speed Engine joints are varied along said line of operation with a minimum level of fuel consumption in accordance with a change in the required engine output. 5. The system according to claim 4, in which, when expressed in two-dimensional coordinates as a function of engine torque and engine speed, the ratio between engine torque and engine speed, expressed as a line of work with high torque, is pre-set on the high side torque from the line of operation with a minimum level of fuel consumption, the target degree of mechanical compression when the engine torque and engine speed are on the line of work with high torque torque is pre-set, and when control is performed to increase the output power, the engine torque and engine speed change from values on the said line of operation with a minimum level of fuel consumption to values on the said line of work with high torque after the degree mechanical compression is reduced to the target degree of mechanical compression. 6. The system according to claim 5, in which the target degree of mechanical compression is the minimum value of the degree of mechanical compression, and the line of work with high torque becomes a curve where fuel consumption becomes minimal when the engine is running in a state where the degree of mechanical compression is reduced to a minimum values. 7. The system according to claim 5, in which after the control is performed to increase the output power, and the engine torque and engine speed reach values on the high-torque line, the engine torque and engine speed are changed along the line work with high torque, while the required engine output is the boundary output power or more. 8. The system according to claim 5, in which, when expressed in two-dimensional coordinates as a function of engine torque and engine speed, the ratio between engine torque and engine speed, expressed as a line with a full load, is pre-installed on the side further more high engine torque from the high torque line, and when additional high torque is required, the engine torque and engine speed are changed by two The above mentioned line of work with full load. 9. The system according to claim 4, wherein, when expressed in two-dimensional coordinates as a function of engine torque and engine speed, the relationship between engine torque and engine speed, expressed as a line of high torque, is pre-set on the high side torque from the work line with a minimum level of fuel consumption, the target degree of mechanical compression when the engine torque and engine speed are on the work line with high twisting it is pre-set by the moment, and when the required engine output increases above the boundary output power, first the engine torque and the engine speed are found on said high-torque line, satisfying the required output power, then the engine speed becomes the target engine speed, then the engine torque increases from a value on the mentioned line of operation with a minimum level of flow fuel to said target torque to said line with high torque after the mechanical compression ratio falls to the target mechanical compression ratio.

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