JP5915566B2 - Engine control device - Google Patents

Description

  The present invention relates to an engine control device.

  It has a pair of motor generators and is equipped with an output adjustment device that receives the output of the engine and generates an output for driving the vehicle. The output adjustment device is formed so that the output torque of the engine is distributed to each motor generator. In the hybrid type vehicle, the engine is composed of an engine having a variable compression ratio mechanism and a variable valve timing mechanism, and the engine torque and the engine speed that optimize the fuel efficiency while maintaining the engine output at the required engine output. A vehicle in which a combination is set is known (see, for example, Patent Document 1).

  However, for example, when the engine load becomes low in a high-speed operation state, it is necessary to keep the engine speed low while the other motor generator has a high speed. In this case, it is necessary to generate a reverse driving force with one motor generator, that is, it is necessary to operate one motor generator as an electric motor. At this time, in order to obtain electric power necessary to drive one motor generator, it is necessary to generate electric power with the other motor generator, that is, it is necessary to operate the other motor generator as a generator. Such power flow is called power circulation, which is inefficient and undesirable. Therefore, a hybrid vehicle is also known in which a non-locking state in which one motor generator is allowed to rotate is temporarily switched from a non-locking state in which one motor generator is prevented from rotating (for example, Patent Documents). 2).

International Publication No. 2010/113332 JP 2010-274735 A

  However, when the vehicle described in Patent Document 1 is provided with the lock mechanism, the combination of the engine torque and the engine speed that optimizes the fuel consumption in the locked state is the combination of the engine torque and the engine speed that optimizes the fuel consumption in the unlocked state. It does not necessarily match the combination. Therefore, there is a possibility that the fuel consumption is deteriorated in the locked state.

  An object of the present invention is to provide an engine control device that can obtain better fuel efficiency in a locked state while preventing generation of power circulation.

  According to the present invention, there is provided an output adjusting device that has a pair of motor generators and that receives an output of the engine and generates an output for driving the vehicle, and the output adjusting device has an output torque of the engine applied to each motor generator. It is configured to distribute torque, and the output adjustment device further switches temporarily from an unlocked state where one motor generator is allowed to rotate to a locked state where one motor generator is prevented from rotating. A variable compression ratio mechanism capable of changing the mechanical compression ratio, a variable valve timing mechanism capable of controlling the closing timing of the intake valve, and an intake air amount A throttle valve arranged in the engine intake passage, and the mechanical compression ratio of the engine is equal to or higher than a predetermined compression ratio. A first ultra-high expansion ratio cycle is performed in which the intake air amount is controlled by controlling the throttle opening while maintaining the closing timing of the intake valve on the side away from the intake bottom dead center, The engine further includes an exhaust gas recirculation mechanism capable of performing an exhaust gas recirculation operation by connecting the exhaust passage and the intake air passage of the engine by an exhaust gas recirculation passage, and the exhaust gas recirculation operation is performed in a locked state. An engine control device is provided.

  It is possible to obtain better fuel efficiency in the locked state while preventing the generation of power circulation.

1 is an overall view of an engine and an output adjustment device. It is a figure for demonstrating an effect | action of an output adjustment apparatus. It is a figure which shows the relationship between an engine output, engine torque Te, and engine speed Ne. It is a flowchart for performing driving control of a vehicle. It is a figure for demonstrating charging / discharging control of a battery. FIG. 2 is an overall view of the engine shown in FIG. 1. It is a disassembled perspective view of a variable compression ratio mechanism. It is side surface sectional drawing of the engine represented schematically. It is a figure which shows a variable valve timing mechanism. It is a figure which shows the lift amount of an intake valve and an exhaust valve. It is a figure for demonstrating a mechanical compression ratio, an actual compression ratio, and an expansion ratio. It is a figure which shows the relationship between theoretical thermal efficiency and an expansion ratio. It is a figure for demonstrating a normal cycle and a super-high expansion ratio cycle. It is a figure which shows changes, such as a mechanical compression ratio according to an engine torque. It is a figure which shows an equal output line and each operation line. It is a figure for demonstrating the effect | action of a locking mechanism. It is a figure which shows area | region AEGR. It is a figure which shows the map of the opening degree VEGR of an EGR control valve. It is a flowchart for performing cycle control. It is a flowchart for performing lock control. It is a flowchart for performing EGR control. It is a figure which shows allowable upper limit TeU. It is a figure which shows the allowable lower limit NeL. It is a flowchart for performing EGR control in another Example by this invention. It is a time chart explaining another Example by this invention. It is a flowchart for performing cycle control in still another embodiment according to the present invention. It is a time chart explaining another Example by this invention. It is a flowchart for performing exhaust gas temperature control in another Example by this invention. It is a flowchart for performing cycle control in still another embodiment according to the present invention. It is a flowchart for performing EGR control in still another embodiment according to the present invention. It is a time chart explaining another Example by this invention. It is a time chart explaining another Example by this invention. It is a flowchart for performing cycle control in still another embodiment according to the present invention.

  FIG. 1 shows an overall view of a spark ignition engine 1 and an output adjustment device 2 mounted on a hybrid vehicle.

  First, the output adjustment device 2 will be briefly described with reference to FIG. In the embodiment shown in FIG. 1, the output adjustment device 2 includes a pair of motor generators MG1 and MG2 that operate as an electric motor and a generator and a planetary gear mechanism 3. The planetary gear mechanism 3 includes a sun gear 4, a ring gear 5, a planetary gear 6 disposed between the sun gear 4 and the ring gear 5, and a planetary carrier 7 that carries the planetary gear 6. Sun gear 4 is connected to a rotating shaft 8 of motor generator MG 1, and planetary carrier 7 is connected to an output shaft 9 of engine 1. Ring gear 5 is connected on the one hand to rotating shaft 10 of motor generator MG2 and on the other hand to output shaft 12 connected to the drive wheel via belt 11. Therefore, it can be seen that when the ring gear 5 rotates, the output shaft 12 is rotated accordingly.

  Each motor generator MG1, MG2 is mounted on a corresponding rotating shaft 8, 10 and has rotors 13, 15 having a plurality of permanent magnets mounted on the outer peripheral surface thereof, and a stator 14 wound with an exciting coil for forming a rotating magnetic field. , 16 and an AC synchronous motor. Excitation coils of stators 14 and 16 of motor generators MG1 and MG2 are connected to corresponding motor drive control circuits 17 and 18, respectively, and these motor drive control circuits 17 and 18 are connected to a battery 19 that generates a DC high voltage. . In the embodiment shown in FIG. 1, the motor generator MG2 mainly operates as an electric motor, and the motor generator MG1 mainly operates as a generator.

  The output adjustment device 2 further includes a lock mechanism LM that temporarily switches from an unlocked state in which the rotation of the motor generator MG1 is allowed to a locked state in which the rotation of the motor generator MG1 is blocked. In the embodiment shown in FIG. 1, the lock mechanism LM includes a clutch. This clutch is fixed to the rotor 13 of the motor generator MG1 and is rotatable with the rotor 13. The clutch plate CL1 is fixed to the housing and cannot be rotated. Plate CL2. The clutch plate CL2 is movable between a position where it is disengaged from the clutch plate CL1 and a position where it is engaged with the clutch plate CL1. When switching to the unlocked state, the clutch plate CL2 is moved to a position where it is disengaged from the clutch plate CL1. As a result, rotation of motor generator MG1 is allowed. On the other hand, when switching to the locked state, the clutch plate CL2 is moved to a position where it is engaged with the clutch plate CL1. As a result, rotation of motor generator MG1 is prevented.

  The electronic control unit 20 comprises a digital computer and is connected to each other by a bidirectional bus 21. A ROM (read only memory) 22, a RAM (random access memory) 23, a CPU (microprocessor) 24, an input port 25 and an output port 26 are connected. It comprises. A load sensor 28 that generates an output voltage proportional to the depression amount L of the accelerator pedal 27 is connected to the accelerator pedal 27, and the output voltage of the load sensor 28 is input to the input port 25 via the corresponding AD converter 25a. . The input port 25 is connected to a crank angle sensor 29 that generates an output pulse every time the crankshaft rotates, for example, 15 °. Furthermore, a signal representing the charging / discharging current of the battery 19 and other various signals are input to the input port 25 via the corresponding AD converter 25a. On the other hand, the output port 26 is connected to each of the motor drive control circuits 17 and 18 and is connected to an element to be controlled of the engine 1, such as a fuel injection valve, through a corresponding drive circuit 26a.

  When driving the motor generator MG2, the DC high voltage of the battery 19 is converted into a three-phase alternating current having a frequency of fm and a current value of Im in the motor drive control circuit 18, and this three-phase alternating current is supplied to the exciting coil of the stator 16. . This frequency fm is a frequency necessary for rotating the rotating magnetic field generated by the exciting coil in synchronization with the rotation of the rotor 15, and this frequency fm is calculated by the CPU 24 based on the rotational speed of the output shaft 10. In the motor drive control circuit 18, this frequency fm is a three-phase AC frequency. On the other hand, the output torque of motor generator MG2 is substantially proportional to current value Im of the three-phase AC. The current value Im is calculated by the CPU 24 based on the required output torque of the motor generator MG2, and the motor drive control circuit 18 sets the current value Im as a three-phase AC current value.

  Further, when motor generator MG2 is driven by an external force, motor generator MG2 operates as a generator, and the electric power generated at this time is regenerated in battery 19. The CPU 24 calculates the required drive torque when the motor generator MG2 is driven by an external force, and activates the motor drive control circuit 18 so that the required drive torque acts on the rotary shaft 10.

  Such drive control for motor generator MG2 is similarly performed for motor generator MG1. That is, when the motor generator MG1 is driven, the DC high voltage of the battery 19 is converted into a three-phase alternating current having a frequency of fm and a current value of Im in the motor drive control circuit 17, and this three-phase alternating current is supplied to the exciting coil of the stator 14. Is done. When motor generator MG1 is driven by an external force, motor generator MG1 operates as a generator, and the electric power generated at this time is regenerated in battery 19. At this time, the motor drive control circuit 17 is operated so that the required drive torque calculated on the rotary shaft 8 acts.

  Next, the relationship between the torque acting on the shafts 8, 9, and 10 and the relationship between the rotational speeds of the shafts 8, 9, and 10 will be described with reference to FIG.

In FIG. 2A, r ₁ indicates the radius of the pitch circle of the sun gear 4, and r ₂ indicates the radius of the pitch circle of the ring gear 5. Assume that torque Te is applied to the output shaft 9 of the engine 1 in the state shown in FIG. 2A to generate a force F in the rotation center portion of the planetary gear 6 in the rotational direction of the output shaft 9. At this time, a force F / 2 in the same direction as the force F acts on the sun gear 4 and the ring gear 5 at the meshing portion with the planetary gear 6. As a result, torque Tes (= (F / 2) · r ₁ ) acts on the rotating shaft 8 of the sun gear 4, and torque Ter (= (F / 2) · r ₂ ) acts on the rotating shaft 10 of the ring gear 5. Will act. On the other hand, since the torque Te acting on the output shaft 9 of the engine 1 is expressed by F · (r ₁ + r ₂ ) / 2, the torque Tes acting on the rotating shaft 8 of the sun gear 4 is represented by r ₁ , r ₂ , Te. Is expressed as Tes = (r ₁ / (r ₁ + r ₂ )) · Te, and the torque Ter acting on the rotating shaft 10 of the ring gear 5 is expressed as r ₁ , r ₂ , and Te, Ter = (r ₂ / ( r ₁ + r ₂ )) · Te.

That is, the torque Te generated on the output shaft 9 of the engine 1 is distributed at a ratio of r ₁ : r ₂ between the torque Tes acting on the rotating shaft 8 of the sun gear 4 and the torque Ter acting on the rotating shaft 10 of the ring gear 5. It will be. In this case, since r ₂ > r ₁ , the torque Ter acting on the rotating shaft 10 of the ring gear 5 is necessarily greater than the torque Tes acting on the rotating shaft 8 of the sun gear 4. Note that if the pitch circle radius r ₁ of the sun gear 4 / the radius r ₂ of the pitch circle of the ring gear 5, that is, the number of teeth of the sun gear 4 / the number of teeth of the ring gear 5 is ρ, Tes is Tes = (ρ / (1 + ρ) ) · Te, and Ter is expressed as Ter = (1 / (1 + ρ)) · Te.

On the other hand, when the rotation direction of the output shaft 9 of the engine 1, that is, the direction of action of the torque Te indicated by the arrow in FIG. 2A is the forward rotation direction, the sun gear 4 is rotated forward while the planetary carrier 7 stops rotating. When rotated in the direction, the ring gear 5 rotates in the opposite direction. At this time, the ratio of the rotational speeds of the sun gear 4 and the ring gear 5 is r ₂ : r ₁ . Dashed Z ₁ in FIG. 2 (B) represents the rotational speed of the relationship between the time schematically. In FIG. 2B, the vertical axis indicates the normal rotation direction and the lower direction indicates the reverse rotation direction with respect to zero 0. In FIG. 2B, S indicates the sun gear 4, C indicates the planetary carrier 7, and R indicates the ring gear 5. As shown in FIG. 2B, the rotation of the sun gear S, the planetary carrier C, and the ring gear R is set with the interval between the planetary carrier C and the ring gear R as r ₁ and the interval between the planetary carrier C and the sun gear S as r _2. When notation number by black circles that indicate the respective number of revolutions will be positioned on a straight line indicated by a broken line Z _1.

On the other hand, when the relative rotation among the sun gear 4, the ring gear 5 and the planetary gear 6 is stopped and the planetary carrier 7 is rotated in the forward direction, the sun gear 4, the ring gear 5 and the planetary carrier 7 are rotated at the same rotational speed in the forward direction. Rotate. Rotational speed of the relationship at this time is shown by a broken line Z _2. Therefore the actual rotational speed of the relationship is represented by a solid line Z obtained by Ju疊dashed Z ₁ in broken lines Z _2, indicated by the solid line Z sun gear S, planetary carrier C, points representing the rotation speed of the ring gear R is thus It will be on a straight line. Therefore, when any two of the sun gear S, the planetary carrier C, and the ring gear R are determined, the remaining one is automatically determined. If the relationship of r ₁ / r ₂ = ρ described above is used, the interval between the sun gear S and the planetary carrier C and the interval between the planetary carrier C and the ring gear R are 1: ρ as shown in FIG. It becomes.

  FIG. 2C schematically shows the rotational speeds of the sun gear S, the planetary carrier C, and the ring gear R, and the torque that acts on the sun gear S, the planetary carrier C, and the ring gear R. The vertical and horizontal axes in FIG. 2C are the same as those in FIG. 2B, and the solid line shown in FIG. 2C corresponds to the solid line shown in FIG. On the other hand, in FIG. 2C, the torque acting on the corresponding rotating shaft is shown at each black dot representing the rotational speed. In each torque, when the direction in which the torque acts and the rotation direction are the same, the case where the driving torque is applied to the corresponding rotation shaft is shown, and the direction in which the torque acts is opposite to the rotation direction. In this case, the corresponding rotating shaft gives torque.

In the example shown in FIG. 2C, the engine torque Te is acting on the planetary carrier C, and this engine torque Te is distributed to the torque Ter applied to the ring gear R and the torque Tes applied to the sun gear S. . Ring to the rotary shaft 10 of the gear R and acts with the vehicle drive torque Tr for driving the torque Tm ₂ and a vehicle engine torque Ter distributed and motor generator MG2, these torque _Ter, Tm 2, Tr Are in balance. In the case shown in FIG. 2C, the torque Tm ₂ has the same direction of rotation as that of the torque, and the rotational direction is the same. Therefore, the torque Tm ₂ gives a driving torque to the rotating shaft 10 of the ring gear R. Accordingly, at this time, motor generator MG2 operates as a drive motor. Figure 2 is the sum of the driving torque Tm ₂ by the engine torque Ter and the motor generator MG2 that is dispensed at this time in the case shown in (C) becomes equal to the vehicle drive torque Tr, therefore at this time the vehicle has an engine 1 It is driven by motor generator MG2.

On the other hand, the torque Tm ₁ of the engine torque Tes distributed and motor generator MG1 are acting, it is balanced to these torques Tes and Tm ₁ to the rotary shaft 8 of the sun gear 5. In the case shown in FIG. 2 (C), the torque Tm ₁ is reverse in the direction in which the torque acts and the direction of rotation, so that the torque Tm ₁ is given drive torque from the rotating shaft 10 of the ring gear R. Accordingly, at this time, the motor generator MG1 operates as a generator. That is, the engine torque Tes distributed at this time is equal to the torque for driving the motor generator MG1, and therefore the motor generator MG1 is driven by the engine 1 at this time.

  In FIG. 2C, Nr, Ne, and Ns indicate the rotational speeds of the rotating shaft 10 of the ring gear R, the rotating shaft of the planetary carrier C, that is, the driving shaft 9 and the rotating shaft 8 of the sun gear S, respectively. From (C), the relationship between the rotational speeds of the shafts 8, 9, and 10 and the relationship between the torques acting on the shafts 8, 9, and 10 can be understood at a glance. 2C is called a collinear diagram, and the solid line shown in FIG. 2C is called an operation collinear.

As shown in FIG. 2C, when the vehicle drive torque is Tr and the rotation speed of the ring gear 5 is Nr, the vehicle drive output Pr for driving the vehicle is represented by Pr = Tr · Nr. Is done. Further, the output Pe of the engine 1 at this time is represented by a product Te · Ne of the engine torque Te and the engine speed Ne. On the other hand, at this time, the power generation energy of motor generator MG1 is similarly represented by the product of torque and rotation speed, and therefore the power generation energy of motor generator MG1 is Tm ₁ · Ns. Further, the driving energy of motor generator MG2 is also expressed by the product of torque and rotational speed, and therefore the driving energy of motor generator MG2 is Tm ₂ · Nr. Here, assuming that power generation energy Tm ₁ · Ns of motor generator MG ₁ is made equal to drive energy Tm ₂ · Nr of motor generator MG 2 and motor generator MG 2 is driven by electric power generated by motor generator MG 1, all of engine 1 The output Pe is used as the vehicle drive output Pr. At this time, Pr = Pe, and therefore Tr · Nr = Te · Ne. That is, the engine torque Te is converted into the vehicle driving torque Tr. Therefore, the output adjusting device 2 performs a torque conversion action. In reality, since there is a power generation loss and a gear transmission loss, it is not possible to use all the output Pe of the engine 1 as the vehicle drive output Pr. However, the output adjustment device 2 is performing a torque conversion action. There is no.

FIG. 3A shows the equal output lines Pe _{1 to} Pe ₉ of the engine 1, and Pe ₁ <Pe ₂ <Pe ₃ <Pe ₅ <Pe ₅ <Pe ₆ <Pe between the magnitudes of the outputs. ₇ <Pe ₈ <Pe ₉ 3A shows the engine torque Te, and the horizontal axis in FIG. 3A shows the engine speed Ne. As can be seen from FIG. 3A, there are innumerable combinations of the engine torque Te and the engine speed Ne that satisfy the required output Pe of the engine 1 required for driving the vehicle. In this case, what kind of engine torque Te Even if the combination of the engine speed Ne and the engine speed Ne is selected, the engine torque Te can be converted into the vehicle driving torque Tr in the output adjusting device 2. Therefore, when this output adjusting device 2 is used, it is possible to set a combination of a desired engine torque Te and an engine speed Ne at which the same engine output Pe can be obtained. In the present invention, as will be described later, a combination of the engine torque Te and the engine speed Ne that can obtain the best fuel consumption while ensuring the required output Pe of the engine 1 is set. The relationship shown in FIG. 3A is stored in the ROM 22 in advance.

  FIG. 3B shows an equal accelerator opening line of the accelerator pedal 27, that is, an equal depression amount line L, and the depression amount L is shown as a percentage with respect to each equal depression amount line L. The vertical axis in FIG. 3B indicates the required vehicle driving torque TrX required for driving the vehicle, and the horizontal axis in FIG. 3B indicates the rotational speed Nr of the ring gear 5. Yes. It can be seen from FIG. 3B that the required vehicle driving torque TrX is determined from the depression amount L of the accelerator pedal 27 and the rotation speed Nr of the ring gear 5 at that time. The relationship shown in FIG. 3B is stored in the ROM 22 in advance.

  Next, a basic control routine for driving the vehicle will be described with reference to FIG. This routine is executed by interruption every predetermined time.

  Referring to FIG. 4, first, at step 100, the rotational speed Nr of the ring gear 5 is detected. Next, at step 101, the depression amount L of the accelerator pedal 27 is read. Next, at step 102, the required vehicle drive torque TrX is calculated from the relationship shown in FIG. Next, at step 103, the required vehicle drive output Tr (= TrX · Nr) is calculated by multiplying the required vehicle drive torque TrX by the rotational speed Nr of the ring gear 5. Next, at step 104, the engine 1 is added to the required vehicle drive output Pr by adding the engine output Pd to be increased or decreased for charging / discharging of the battery 19 and the engine output Ph required to drive the auxiliary machine. The required output Pn is calculated. The engine output Pd for charging / discharging the battery 19 is calculated by a routine shown in FIG.

  Next, at step 105, the final required output Pe (= Pn / ηt) of the engine 1 is calculated by dividing the output Pr required of the engine 1 by the torque conversion efficiency ηt in the output adjusting device 2. Next, at step 106, a required engine torque TeX, a required engine speed NeX, etc. that satisfy the required output Pe of the engine and obtain the minimum fuel consumption are set from the relationship shown in FIG. The setting of the required engine torque TeX and the required engine speed NeX will be described later. In the present invention, the minimum fuel consumption means the minimum fuel consumption when not only the efficiency of the engine 1 but also the gear transmission efficiency of the output adjusting device 2 is taken into consideration.

Next, at step 107, the required torque Tm ₂ X (= TrX−Ter = TrX−TeX / (1 + ρ)) of the motor generator MG2 is calculated from the required vehicle driving torque TrX and the required engine torque TeX. Next, at step 108, the required rotational speed NsX of the sun gear 4 is calculated from the rotational speed Nr of the ring gear 5 and the required engine rotational speed NeX. Since (NeX−Ns) :( Nr−NeX) = 1: ρ from the relationship shown in FIG. 2C, the required rotation speed NsX of the sun gear 4 is Nr− ( Nr−NeX) · (1 + ρ) / ρ.

Next, at step 109, the motor generator MG1 is controlled so that the rotational speed of the motor generator MG1 becomes the required rotational speed NsX. When the rotational speed of motor generator MG1 reaches required rotational speed NsX, engine rotational speed Ne becomes required engine rotational speed NeX, and therefore engine rotational speed Ne is controlled to required engine rotational speed NeX by motor generator MG1. Next, at 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, at step 111, the fuel injection amount necessary for obtaining the required engine torque TeX, the target opening degree of the throttle valve, and the like are calculated. At step 112, the engine 1 is controlled based on them.

By the way, in the hybrid type vehicle, it is necessary to always maintain the charge amount of the battery 19 at a predetermined amount or more. Therefore, in the embodiment according to the present invention, the charge amount SOC is set to the lower limit value SC ₁ as shown in FIG. It is to be maintained between the upper limit value SC ₂ and. That is, in the embodiment according to the present invention the engine output is increased forcibly in order to increase the power generation amount and the charge amount SOC is lower than the lower limit value SC _1, the motor-generator when the charge amount SOC exceeds the upper limit value SC ₂ In order to increase the power consumption, the engine output is forcibly reduced. The charge amount SOC is calculated, for example, by integrating the charge / discharge current I of the battery 19.

  FIG. 5B shows a charging / discharging control routine of the battery 19, and this routine is executed by interruption every predetermined time.

Referring to FIG. 5B, first, at step 120, charge / discharge current I is added to battery 19 to charge amount SOC. This current value I is positive during charging and negative during discharging. Then whether forced or is being charged is determined on the battery 19 at step 121, whether it charge SOC proceeds to step 122 becomes lower than the lower limit value SC ₁ when forced not in charge determination Is done. When SOC <SC ₁ , the routine proceeds to step 124, where the engine output Pd at step 104 in FIG. 4 is set to a predetermined value Pd ₁ . At this time, the engine output is forcibly increased and the battery 19 is forcibly charged. When the battery 19 is forcibly charged, the routine proceeds from step 121 to step 123, where it is determined whether or not the forcible charging operation has been completed, and the routine proceeds to step 124 until the forcible charging operation is completed.

On the other hand, it is decided whether or not forcibly discharged from the battery 19 is determined the routine proceeds to step 125 when it is judged that SOC ≧ SC ₁ in step 122. Charge SOC proceeds to step 126 when not in forcibly discharged whether exceeds the upper limit value SC ₂ is discriminated. Engine power Pd is a value -Pd ₂ which is predetermined in step 104 of FIG. 4 proceeds to step 128 becomes a SOC> SC _2. At this time, the engine output is forcibly decreased and the battery 19 is forcibly discharged. When the battery 19 is forcibly discharged, the routine proceeds from step 125 to step 127, where it is determined whether or not the forcible discharging action has been completed, and the routine proceeds to step 128 until the forcible discharging action is completed.

  Next, the spark ignition engine shown in FIG. 1 will be described with reference to FIG.

  Referring to FIG. 6, 30 is a crankcase, 31 is a cylinder block, 32 is a cylinder head, 33 is a piston, 34 is a combustion chamber, 35 is a spark plug disposed in the center of the top surface of the combustion chamber 34, and 36 is an intake air. Reference numeral 37 denotes an intake port, reference numeral 38 denotes an exhaust valve, and reference numeral 39 denotes an exhaust port. The intake port 37 is connected to a surge tank 41 via an intake branch pipe 40, and each intake branch pipe 40 is provided with a fuel injection valve 42 for injecting fuel into the corresponding intake port 37. The fuel injection valve 42 may be disposed in each combustion chamber 34 instead of being attached to each intake branch pipe 40. On the other hand, a water temperature sensor 31 a for detecting the temperature of engine cooling water is attached to the cylinder block 31.

  The surge tank 41 is connected to an air cleaner 44 via an intake duct 43, and a throttle valve 46 driven by an actuator 45 and an intake air amount detector 47 using, for example, heat rays are arranged in the intake duct 43. On the other hand, the exhaust port 39 is connected via an exhaust manifold 48 to, for example, a catalytic converter 49 containing a three-way catalyst, and an air-fuel ratio sensor 49 a is disposed in the exhaust manifold 48. An exhaust pipe 48a is connected to the catalytic converter 49, and a temperature sensor 49b for detecting the temperature of the exhaust gas flowing out from the three-way catalyst is attached to the exhaust pipe 48a. The temperature of the exhaust gas represents the temperature of the three-way catalyst.

  On the other hand, in the embodiment shown in FIG. 6, the piston 33 is positioned at the compression top dead center by changing the relative position of the crankcase 30 and the cylinder block 31 in the cylinder axial direction at the connecting portion between the crankcase 30 and the cylinder block 31. A variable compression ratio mechanism A is provided that can change the volume of the combustion chamber 34 when the intake valve 36 is closed. Further, in order to control the amount of intake air actually supplied into the combustion chamber 34, the closing timing of the intake valve 36 is set. A controllable variable valve timing mechanism B is provided.

  7 shows an exploded perspective view of the variable compression ratio mechanism A shown in FIG. 6, and FIG. 8 shows a side sectional view of the engine 1 schematically. Referring to FIG. 7, a plurality of protrusions 50 spaced from each other are formed below both side walls of the cylinder block 31, and cam insertion holes 51 each having a circular cross section are formed in each protrusion 50. Has been. On the other hand, a plurality of protrusions 52 are formed on the upper wall surface of the crankcase 30 so as to be fitted between the corresponding protrusions 50 spaced apart from each other. Cam insertion holes 53 each having a circular cross section are formed.

  As shown in FIG. 7, a pair of camshafts 54 and 55 are provided, and on each camshaft 54 and 55, a circular cam 56 is rotatably inserted into each cam insertion hole 51. It is fixed. These circular cams 56 are coaxial with the rotational axes of the camshafts 54 and 55. On the other hand, an eccentric shaft 57 arranged eccentrically with respect to the rotation axis of each camshaft 54, 55 extends between the circular cams 56 as shown by hatching in FIG. A cam 58 is eccentrically mounted for rotation. As shown in FIG. 7, the circular cams 58 are disposed between the circular cams 56, and the circular cams 58 are rotatably inserted into the corresponding cam insertion holes 53.

  When the circular cams 56 fixed on the camshafts 54 and 55 are rotated in opposite directions as shown by solid arrows in FIG. 8A from the state shown in FIG. In order to move toward the lower center, the circular cam 58 rotates in the direction opposite to the circular cam 56 in the cam insertion hole 53 as shown by the broken arrow in FIG. 8A, and is shown in FIG. 8B. As described above, when the eccentric shaft 57 moves to the lower center, the center of the circular cam 58 moves below the eccentric shaft 57.

  8A and 8B, the relative positions of the crankcase 30 and the cylinder block 31 are determined by the distance between the center of the circular cam 56 and the center of the circular cam 58. As the distance between the center and the center of the circular cam 58 increases, the cylinder block 31 moves away from the crankcase 30. When the cylinder block 31 moves away from the crankcase 30, the volume of the combustion chamber 34 when the piston 33 is located at the compression top dead center is increased. Therefore, by rotating the camshafts 54 and 55, the piston 33 is compressed at the top dead center. The volume of the combustion chamber 34 when it is located at can be changed.

  As shown in FIG. 7, in order to rotate the camshafts 54 and 55 in opposite directions, a pair of worm gears 61 and 62 having screw directions opposite to each other are attached to the rotation shaft of the drive motor 59, respectively. Gears 63 and 64 that mesh with the worm gears 61 and 62 are fixed to the end portions of the camshafts 54 and 55, respectively. In this embodiment, by driving the drive motor 59, the volume of the combustion chamber 34 when the piston 33 is located at the compression top dead center can be changed over a wide range. The variable compression ratio mechanism A shown in FIGS. 6 to 8 shows an example, and any type of variable compression ratio mechanism can be used.

  On the other hand, FIG. 9 shows the variable valve timing mechanism B attached to the end of the camshaft 70 for driving the intake valve 36 in FIG. Referring to FIG. 9, the variable valve timing mechanism B includes a timing pulley 71 that is rotated in the arrow direction by the output shaft 9 of the engine 1 via a timing belt, a cylindrical housing 72 that rotates together with the timing pulley 71, A rotating shaft 73 that rotates together with the intake valve driving camshaft 70 and that can rotate relative to the cylindrical housing 72, and a plurality of partitions that extend from the inner peripheral surface of the cylindrical housing 72 to the outer peripheral surface of the rotating shaft 73. A wall 74 and a vane 75 extending from the outer peripheral surface of the rotating shaft 73 to the inner peripheral surface of the cylindrical housing 72 between the partition walls 74 are provided. 76 and a retarding hydraulic chamber 77 are formed.

  The hydraulic oil supply control to the hydraulic chambers 76 and 77 is performed by a hydraulic oil supply control valve 78. The hydraulic oil supply control valve 78 includes hydraulic ports 79 and 80 connected to the hydraulic chambers 76 and 77, a hydraulic oil supply port 82 discharged from the hydraulic pump 81, a pair of drain ports 83 and 84, And a spool valve 85 for controlling communication between the ports 79, 80, 82, 83, and 84.

  When the cam phase of the intake valve driving camshaft 70 should be advanced, the spool valve 85 is moved to the right in FIG. 9 and the hydraulic oil supplied from the supply port 82 is advanced via the hydraulic port 79. The hydraulic oil in the retard hydraulic chamber 77 is discharged from the drain port 84 while being supplied to the hydraulic chamber 76. At this time, the rotary shaft 73 is rotated relative to the cylindrical housing 72 in the direction of the arrow.

  On the other hand, when the cam phase of the intake valve driving camshaft 70 should be retarded, the spool valve 85 is moved to the left in FIG. 9, and the hydraulic oil supplied from the supply port 82 causes the hydraulic port 80 to move. The hydraulic oil in the advance hydraulic chamber 76 is discharged from the drain port 83 while being supplied to the retard hydraulic chamber 77. At this time, the rotating shaft 73 is rotated relative to the cylindrical housing 72 in the direction opposite to the arrow.

  If the spool valve 85 is returned to the neutral position shown in FIG. 9 while the rotation shaft 73 is rotated relative to the cylindrical housing 72, the relative rotation operation of the rotation shaft 73 is stopped, and the rotation shaft 73 The relative rotation position at that time is held. Therefore, the variable valve timing mechanism B can advance and retard the cam phase of the intake valve driving camshaft 70 by a desired amount.

  In FIG. 10, the solid line shows the time when the cam phase of the intake valve driving camshaft 70 is most advanced by the variable valve timing mechanism B, and the broken line shows the cam phase of the intake valve driving camshaft 70 being the most advanced. It shows when it is retarded. Therefore, the valve opening period of the intake valve 36 can be arbitrarily set between the range indicated by the solid line and the range indicated by the broken line in FIG. 10, and therefore the closing timing of the intake valve 36 is also the range indicated by the arrow C in FIG. Any crank angle can be set.

  The variable valve timing mechanism B shown in FIGS. 6 and 9 shows an example, and for example, a variable valve timing that can change only the closing timing of the intake valve while keeping the opening timing of the intake valve constant. Various types of variable valve timing mechanisms, such as mechanisms, can be used.

  Referring again to FIG. 6, the engine 1 includes an exhaust gas recirculation (hereinafter referred to as EGR) mechanism 90. The EGR mechanism 90 includes an EGR passage 91 that connects the surge tank 41 and the exhaust manifold 48 to each other. An electric control type EGR control valve 92 for controlling the amount of EGR gas supplied to the engine 1 is arranged in the EGR passage 91, and a cooling device 93 for cooling the EGR gas flowing in the EGR passage 91 is disposed around the EGR passage 91. Be placed. When the EGR action for supplying EGR gas to the engine 1 is to be performed, the EGR control valve 92 is opened. On the other hand, when the EGR action should be stopped, the EGR control valve 92 is closed. Further, when the EGR gas amount is to be increased, the opening degree of the EGR control valve 92 is increased, and when the EGR gas amount is to be decreased, the opening degree of the EGR control valve 92 is decreased. The EGR passage 91 may be connected to the intake port 37.

  Next, the meanings of terms used in the present application will be described with reference to FIG. 11A, 11B, and 11C show an engine having a combustion chamber volume of 50 ml and a piston stroke volume of 500 ml for the sake of explanation. , (B), (C), the combustion chamber volume represents the volume of the combustion chamber when the piston is located at the compression top dead center.

  FIG. 11A explains the mechanical compression ratio. The mechanical compression ratio is a value mechanically determined only from the stroke volume of the piston and the combustion chamber volume during the compression stroke, and this mechanical compression ratio is expressed by (combustion chamber volume + stroke volume) / combustion chamber volume. In the example shown in FIG. 11A, this mechanical compression ratio is (50 ml + 500 ml) / 50 ml = 11.

  FIG. 11B explains the actual compression ratio. This actual compression ratio is a value determined from the actual piston stroke volume and the combustion chamber volume from when the compression action is actually started until the piston reaches top dead center, and this actual compression ratio is (combustion chamber volume + actual (Stroke volume) / combustion chamber volume. That is, as shown in FIG. 11B, even if the piston starts to rise in the compression stroke, the compression action is not performed while the intake valve is open, and the actual compression is performed after the intake valve is closed. The action begins. Therefore, the actual compression ratio is expressed as described above using the actual stroke volume. In the example shown in FIG. 11B, the actual compression ratio is (50 ml + 450 ml) / 50 ml = 10.

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

  Next, the ultra-high expansion ratio cycle used in the present invention will be described with reference to FIGS. FIG. 12 shows the relationship between the theoretical thermal efficiency, the expansion ratio, and the actual compression ratio ε. FIG. 13 shows the relationship between the normal cycle and the ultra-high expansion ratio cycle that are selectively used according to the required engine torque Te in the present invention. A comparison is shown.

  FIG. 13A shows a normal cycle in which the intake valve closes near the bottom dead center and the compression action by the piston is started from the vicinity of the intake bottom dead center. In the example shown in FIG. 13A, the combustion chamber volume is set to 50 ml, and the stroke volume of the piston is set to 500 ml, as in the example shown in FIGS. 11A, 11B, and 11C. As can be seen from FIG. 13A, in a normal cycle, the mechanical compression ratio is (50 ml + 500 ml) / 50 ml = 11, the actual compression ratio is almost 11, and the expansion ratio is also (50 ml + 500 ml) / 50 ml = 11. That is, in a normal internal combustion engine, the mechanical compression ratio, the actual compression ratio, and the expansion ratio are substantially equal.

  The solid line in FIG. 12 shows the change in theoretical thermal efficiency when the actual compression ratio ε and the expansion ratio are substantially equal, that is, in a normal cycle. In this case, it can be seen that the theoretical thermal efficiency increases as the expansion ratio increases, that is, as the actual compression ratio increases. Therefore, in order to increase the theoretical thermal efficiency in a normal cycle, it is only necessary to increase the actual compression ratio. However, the actual compression ratio ε, that is, the expansion ratio can be increased only to a maximum of about 12 due to the restriction of the occurrence of knocking during engine high-load operation, and thus the theoretical thermal efficiency is sufficiently increased in a normal cycle. You can't do that.

  On the other hand, under such circumstances, it has been studied to increase the theoretical thermal efficiency while strictly separating the mechanical compression ratio and the actual compression ratio ε. As a result, the theoretical thermal efficiency is governed by the expansion ratio, and the actual compression ratio ε is certain. It has been found that the actual compression ratio ε has little influence on the theoretical thermal efficiency at higher levels. That is, if the actual compression ratio ε is increased, the explosive force is increased, but a large amount of energy is required for compression, and thus even if the actual compression ratio ε is increased, the theoretical thermal efficiency is hardly increased. On the other hand, when the expansion ratio is increased, the period during which the pressing force acts on the piston during the expansion stroke becomes longer, and thus the period during which the piston applies the rotational force to the crankshaft becomes longer. Therefore, the larger the expansion ratio, the higher the theoretical thermal efficiency. The broken lines in FIG. 12 indicate the theoretical thermal efficiency when the expansion ratio is increased while the actual compression ratio ε is fixed to 5, 6, 7, 8, 9, and 10, respectively. In FIG. 12, black circles indicate the positions of the theoretical thermal efficiency peaks when the actual compression ratio ε is 5, 6, 7, 8, 9, 10. From FIG. 12, the increase in the theoretical thermal efficiency when the expansion ratio is increased while maintaining the actual compression ratio ε at a low value such as 10, and the actual compression ratio ε increases with the expansion ratio as shown by the solid line in FIG. It can be seen that there is no significant difference from the increase in theoretical thermal efficiency when

  Thus, if the actual compression ratio ε is maintained at a low value, knocking does not occur. Therefore, if the expansion ratio is increased while the actual compression ratio ε is maintained at a low value, the theory prevents knocking from occurring. The thermal efficiency can be greatly increased. FIG. 13B shows an example of using the variable compression ratio mechanism A and the variable valve timing mechanism B to increase the expansion ratio while maintaining the actual compression ratio ε at a low value.

  Referring to FIG. 13B, in this example, the variable compression ratio mechanism A reduces the combustion chamber volume from 50 ml to 20 ml. On the other hand, the variable valve timing mechanism B delays the closing timing of the intake valve until the actual piston stroke volume is reduced from 500 ml to 200 ml. As a result, in this example, the actual compression ratio is (20 ml + 200 ml) / 20 ml = 11, and the expansion ratio is (20 ml + 500 ml) / 20 ml = 26. In the normal cycle shown in FIG. 13A, the actual compression ratio is almost 11 and the expansion ratio is 11, as described above. Compared to this case, only the expansion ratio is shown in FIG. 13B. It can be seen that it has been increased to 26. This is why it is called an ultra-high expansion ratio cycle.

  As described above, when the expansion ratio is increased, the theoretical thermal efficiency is improved and the fuel efficiency is improved. Therefore, it is preferable to increase the expansion ratio in the widest possible operating range. However, as shown in FIG. 13B, in the ultra-high expansion ratio cycle, since the actual piston stroke volume during the compression stroke is reduced, the amount of intake air that can be sucked into the combustion chamber 34 is reduced. The high expansion ratio cycle can be adopted only when the amount of intake air supplied into 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. 13B is set, and when the required engine torque Te is high, the normal cycle shown in FIG. 13A is set.

  Next, how the engine 1 is controlled according to the required engine torque Te will be described with reference to FIG.

  FIG. 14 shows changes in the mechanical compression ratio, the expansion ratio, the closing timing of the intake valve 36, the actual compression ratio, the intake air amount, the opening degree of the throttle valve 46, and the fuel consumption according to the required engine torque Te. . The fuel consumption indicates the amount of fuel consumed when the vehicle has traveled a predetermined travel distance in a predetermined travel mode. Therefore, the value indicating the fuel consumption decreases as the fuel consumption improves. In the embodiment according to the present invention, the average air-fuel ratio in the normal combustion chamber 34 is determined by the air-fuel ratio sensor 49a so that unburned HC, CO and NOx in the exhaust gas can be simultaneously reduced by the three-way catalyst in the catalytic converter 49. Feedback control is performed to the theoretical air-fuel ratio based on the output signal. FIG. 12 shows the theoretical thermal efficiency when the average air-fuel ratio in the combustion chamber 34 is thus the stoichiometric air-fuel ratio.

On the other hand, in the embodiment according to the present invention, since the average air-fuel ratio in the combustion chamber 34 is controlled to the stoichiometric air-fuel ratio, the engine torque Te is proportional to the amount of intake air supplied into the combustion chamber 34. As shown in FIG. 14, the intake air amount is decreased as the required engine torque Te decreases. Accordingly, in order to reduce the intake air amount as the required engine torque Te decreases, the closing timing of the intake valve 36 is delayed as shown by the solid line in FIG. As described above, while the intake air amount is controlled by delaying the closing timing of the intake valve 36, the throttle valve 46 is maintained in a fully opened state. On the other hand, by a required engine torque Te becomes lower than a certain value Te ₁ controls the closing timing of the intake valve 36 is not E and controls the intake air amount required for the intake air amount. Therefore, when the required engine torque Te is lower than this value Te ₁ , that is, the limit value Te ₁ , the valve closing timing of the intake valve 36 is held at the limit valve closing timing at the limit value Te _1. The amount of air is controlled.

  On the other hand, as described above, when the required engine torque Te is low, an ultra-high expansion ratio cycle is set. Therefore, as shown in FIG. 14, when the required engine torque Te is low, the expansion ratio is increased by increasing the mechanical compression ratio. As shown in FIG. 12, for example, when the actual compression ratio ε is 10, the theoretical thermal efficiency peaks when the expansion ratio is about 35. Therefore, when the required engine torque Te is low, it is preferable to increase the mechanical compression ratio until the expansion ratio becomes about 35. However, it is difficult to increase the mechanical compression ratio until the expansion ratio reaches about 35 due to structural limitations. Therefore, in the embodiment according to the present invention, the mechanical compression ratio is set to the maximum structurally possible mechanical compression ratio so that the highest possible expansion ratio can be obtained when the required engine torque Te is low.

On the other hand, if the closing timing of the intake valve 36 is advanced so as to increase the intake air amount while maintaining the mechanical compression ratio at the maximum mechanical compression ratio, the actual compression ratio increases. However, the actual compression ratio needs to be maintained at 12 or less at the maximum. Accordingly, when the required engine torque Te is increased and the intake air amount is increased, the mechanical compression ratio is lowered so that the actual compression ratio is maintained at the optimum actual compression ratio. Machine as the required engine torque Te, as shown in FIG. 14 the required engine torque Te so that the actual compression ratio is maintained at the optimum actual compression ratio when exceeding the limit value Te ₂ is increased in this embodiment of the present invention The compression ratio is lowered.

  When the required engine torque Te increases, the mechanical compression ratio is lowered to the minimum mechanical compression ratio, and at this time, the normal cycle shown in FIG.

  Incidentally, in the embodiment according to the present invention, when the engine speed Ne is low, the actual compression ratio ε is between 9 and 11. However, when the engine speed Ne increases, the air-fuel mixture in the combustion chamber 34 is disturbed, so that knocking does not easily occur. Therefore, in the embodiment according to the present invention, the actual compression ratio ε increases as the engine speed Ne increases. It will be lost.

  On the other hand, in the embodiment according to the present invention, the expansion ratio when the ultra high expansion ratio cycle is set is 26 to 30. On the other hand, the actual compression ratio ε = 5 in FIG. 12 indicates the lower limit of the actual compression ratio that can be used practically. In this case, the theoretical thermal efficiency peaks when the expansion ratio is approximately 20. The expansion ratio at which the theoretical air-fuel ratio reaches a peak becomes higher than 20 as the actual compression ratio ε becomes larger than 5, and therefore the expansion ratio is 20 or more in view of the actual compression ratio ε that may be practically used. It can be said that it is preferable. 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 the example shown in FIG. 14, the mechanical compression ratio is continuously changed according to the required engine torque Te. However, the mechanical compression ratio can be changed stepwise according to the required engine torque Te.

  On the other hand, as shown by the broken line in FIG. 14, the intake air amount can also be controlled by advancing the closing timing of the intake valve 36 as the required engine torque Te decreases. Accordingly, if the case shown by the solid line in FIG. 14 and the case shown by the broken line can be included, in the embodiment according to the present invention, the required engine torque Te becomes low at the closing timing of the intake valve 36. As a result, it is moved in a direction away from the intake bottom dead center BDC until the limit valve closing timing at which the amount of intake air supplied into the combustion chamber 34 can be controlled.

By the way, when the expansion ratio becomes high, the theoretical thermal efficiency becomes high and the fuel efficiency becomes good, that is, the fuel efficiency becomes small. Thus the required engine torque Te in FIG. 14, the fuel consumption is minimized when: the limit value Te _2. However, between the limit values Te ₁ and Te ₂ , the actual compression ratio decreases as the required engine torque Te decreases, so the fuel efficiency becomes slightly worse, that is, the fuel efficiency increases. Further, fuel consumption is even higher for the throttle valve 46 is made to closed the required engine torque Te is lower than the limit value Te ₁ region. On the other hand, when the required engine torque Te becomes higher than the limit value Te ₂ , the expansion ratio decreases, so that the fuel efficiency increases as the required engine torque Te increases. Accordingly, when the required engine torque Te is the limit value Te ₂ , that is, the fuel consumption is reduced at the boundary between the region where the mechanical compression ratio is lowered due to the increase in the required engine torque Te and the region where the mechanical compression ratio is maintained at the maximum mechanical compression ratio. The smallest.

Limit value Te ₂ of the engine torque Te fuel consumption is minimized is slightly changed according to the engine speed Ne, but the minimum fuel consumption if it is possible to hold the engine torque Te Anyway the limit value Te ₂ Will be obtained. In the present invention, the output adjusting device 2 is used to maintain the engine torque Te at the limit value Te ₂ even if the required output Pe of the engine 1 changes.

  As described with reference to FIG. 14, in the ultra-high expansion ratio cycle, when the engine torque Te is lower than the limit value Te1, the mechanical compression ratio is maintained at a predetermined mechanical compression ratio, for example, 20 or more. The intake air amount is controlled by controlling the opening degree of the throttle valve 46 while maintaining the closing timing of the intake valve 36 on the side away from the intake bottom dead center. This will be referred to as the first ultra-high expansion ratio cycle. On the other hand, when the engine torque Te is higher than the limit value Te1, the mechanical compression ratio is maintained at a predetermined mechanical compression ratio, and the closing timing of the intake valve 36 is delayed while the throttle valve 45 is kept fully open. The amount of intake air is controlled. This will be referred to as a second ultra-high expansion ratio cycle. Accordingly, in the ultra-high expansion ratio cycle, either the first ultra-high expansion ratio cycle or the second ultra-high expansion ratio cycle is performed.

  Next, a control method of the engine 1 will be described with reference to FIG.

  In FIG. 15, a solid line P1 indicates the relationship between the engine torque Te and the engine rotational speed Ne at which the fuel consumption is minimized when the ultra-high expansion ratio cycle is performed. Accordingly, if the engine torque Te and the engine speed Ne are set to the engine torque Te and the engine speed Ne on the solid line P1 when the ultra-high expansion ratio cycle is performed, the fuel consumption is minimized.

  On the other hand, the solid line P2 in FIG. 15 shows the relationship between the engine torque Te and the engine speed Ne at which the fuel consumption is minimized when the normal cycle is performed. Therefore, if the engine torque Te and the engine speed Ne are set to the engine torque Te and the engine speed Ne on the solid line P2 when the normal cycle is performed, the fuel consumption is minimized.

  In the embodiment according to the present invention, the super high expansion ratio cycle is performed when the required output Pe of the engine 1 is lower than the boundary output PY indicated by the broken line in FIG. In other words, when the required output Pe of the engine 1 is lower than the boundary output PY, the mechanical compression ratio is maintained to be equal to or higher than a predetermined compression ratio and the closing timing of the intake valve 36 is away from the intake bottom dead center. The intake air amount is controlled by controlling the throttle opening or the closing timing of the intake valve 36 while being maintained at the same value. In this case, the required engine torque TeX and the required engine speed NeX are set to the engine torque Te and the engine speed Ne on the solid line P1 in FIG. Accordingly, the engine torque Te and the engine speed Ne are changed along the solid line P1 in accordance with the required output Pe of the engine 1.

  On the other hand, when the required output Pe of the engine 1 is higher than the boundary output PY, a normal cycle is performed. That is, the intake air amount is controlled by controlling the throttle opening while the mechanical compression ratio is maintained below a predetermined compression ratio and the closing timing of the intake valve 36 is maintained close to the intake bottom dead center. Is done. In this case, the required engine torque TeX and the required engine speed NeX are set to the engine torque Te and the engine speed Ne on the solid line P2 in FIG. Accordingly, the engine torque Te and the engine speed Ne are changed along the solid line P2 in accordance with the required output Pe of the engine 1.

  Incidentally, as indicated by a broken line in FIG. 16, when the rotation speed Ns of the sun gear S becomes a negative value, that is, when the sun gear S rotates in the reverse direction, an undesirable power circulation occurs. When power circulation occurs, fuel consumption deteriorates. Therefore, in the embodiment according to the present invention, the fuel consumption in the unlocked state is compared with the fuel consumption in the locked state, and when the fuel consumption in the locked state is smaller than the fuel consumption in the unlocked state, the lock mechanism LM switches to the locked state. . When switched to the locked state, the rotational speed Ns of the sun gear S is maintained at zero while the rotational speed Nr of the ring gear R is maintained as shown by a solid line in FIG. On the other hand, when the fuel consumption in the unlocked state is smaller than the fuel consumption in the locked state, the unlocked state is maintained, or the unlocked state is restored.

  In the embodiment according to the present invention, the fuel consumption in the unlocked state is determined by the fuel consumption in the engine 1 and the battery 19 required to generate the required output of the vehicle when the engine 1 is operated in the unlocked state. It is estimated from the power consumption. Similarly, the fuel consumption in the locked state is estimated from the fuel consumption in the engine 1 and the power consumption in the battery 19 necessary for generating the required output of the vehicle when the engine 1 is operated in the locked state. Is done.

  However, when the normal cycle is performed, the fuel consumption is not reduced even if the lock state is switched. Therefore, in the embodiment according to the present invention, it is switched to the locked state when the ultra-high expansion ratio cycle is performed, and is maintained in the unlocked state when the normal cycle is performed.

  In FIG. 15, a solid line P3 shows the relationship between the engine torque Te and the engine speed Ne at which the fuel consumption is minimized when the ultra-high expansion ratio cycle is performed in the locked state. Therefore, when the engine torque Te and the engine speed Ne are set to the engine torque Te and the engine speed Ne on the solid line P3 when the super high expansion ratio cycle is performed in the locked state, the fuel consumption is minimized.

  When the super-high expansion ratio cycle is to be performed under the locked state, the required engine torque TeX and the required engine speed NeX are set to the engine torque Te and the engine speed Ne on the solid line P3 in FIG. Accordingly, the engine torque Te and the engine speed Ne are changed along the solid line P3 in accordance with the required output Pe of the engine 1.

  Note that the solid lines P1, P2, and P3 shown in FIG. 15 are referred to as operation lines.

  By the way, as described with reference to FIG. 14, when the engine torque Te is lower than the limit value Te1 in the ultra-high expansion ratio cycle, the closing timing of the intake valve 36 is maintained on the side away from the intake bottom dead center. A first ultra-high expansion ratio cycle in which the intake air amount is controlled by controlling the opening degree of the throttle valve 46 is performed. On the other hand, when the engine torque Te is higher than the limit value Te1, the second ultra-high expansion ratio cycle in which the intake air amount is controlled by controlling the closing timing of the intake valve 36 while the throttle valve 45 is kept fully open. Is done. As a result, the pumping loss increases in the first ultra-high expansion ratio cycle.

  On the other hand, considering only the fuel consumption of the engine 1, the fuel consumption in the locked state is larger than the fuel consumption in the non-locked state. In this case, the fuel consumption is considerably increased during the first ultra-high expansion ratio cycle and in the locked state.

  On the other hand, if EGR gas is supplied to the engine 1 to perform the EGR action, the pumping loss can be reduced, and thus the fuel consumption can be reduced.

  Therefore, in the embodiment according to the present invention, the EGR action is performed in the locked state during the first ultra-high expansion ratio cycle. As a result, it is possible to suppress an increase in fuel consumption in the locked state during the first ultra-high expansion ratio cycle.

  In the first ultra-high expansion ratio cycle, the throttle opening increases as the engine torque Te increases. If the EGR action is performed when the throttle opening is large, the exhaust gas may flow backward in the intake duct 43 and pass through the throttle valve 46 to reach the intake air amount detector 47. In this case, there is a possibility that the intake air amount detector 47 cannot accurately detect the intake air amount.

  Further, in the ultra-high expansion ratio cycle, the closing timing of the intake valve 36 is greatly delayed, so that combustion is not always stable. If the EGR action is performed in this state, the combustion may become more unstable. In particular, combustion may become more unstable at low loads when the first ultra-high expansion ratio cycle is performed.

  Therefore, in the embodiment according to the present invention, even in the locked state, when the engine torque Te representing the throttle opening is higher than the allowable upper limit TeU, the EGR action is stopped, and when the engine torque Te is lower than the allowable upper limit TeU. EGR action is performed. Even in the locked state, the EGR action is stopped when the engine speed Ne is lower than the allowable lower limit NeL, and the EGR action is performed when the engine speed Ne is higher than the allowable lower limit.

  That is, as shown in FIG. 17, the engine operating state determined by the engine torque Te and the engine speed Ne is within an area AEGR where the engine torque Te is lower than the allowable upper limit TeU and the engine speed Ne is higher than the allowable lower limit NeL. The EGR action is performed when the engine is in the state, and the EGR action is stopped when the engine operating state is outside the region AEGR. In this way, stable combustion can be obtained while accurately detecting the intake air amount.

  The opening degree VEGR of the EGR control valve 92 when the EGR action is performed is stored in advance in the ROM 22 in the form of a map as a function of the engine torque Te and the engine speed Ne as shown in FIG.

  On the other hand, during the first ultra-high expansion ratio cycle and in the unlocked state, the EGR action is not performed. As a result, combustion is prevented from becoming unstable due to the EGR action.

  FIG. 19 shows a routine for executing the above-described cycle control. This routine is executed by interruption every predetermined time.

  Referring to FIG. 19, first, at step 200, it is judged if the output Pe of the engine 1 is lower than the boundary output PY. When Pe <PY, the routine proceeds to step 201 where an ultra-high expansion ratio cycle is performed. On the other hand, when Pe ≧ PY, the routine proceeds to step 202 where a normal cycle is performed.

  Note that the output Pe of the engine 1 is maintained when the super high expansion ratio cycle is switched to the normal cycle or when the normal cycle is switched to the super high expansion ratio cycle. The output Pe of the engine 1 is also maintained when switching from the unlocked state to the locked state or when switching from the locked state to the unlocked state. That is, as shown in FIG. 15, when switching to the normal cycle when the super high expansion ratio cycle is being performed in the engine operating state represented by the point X1 on the curve P1, the output Pe of the engine 1 is The engine operating state is changed to a point X2 on the curve P2 so as to be maintained at Pe4. On the other hand, when the engine operating state represented by the point X1 should be switched to the locked state, the engine operating state is changed to a point X3 on the curve P3 so that the output Pe of the engine 1 is maintained at Pe4.

  FIG. 20 shows a routine for executing the above-described lock control. This routine is executed by interruption every predetermined time.

  Referring to FIG. 20, first, at step 300, it is judged if the engine 1 is performing an ultra-high expansion ratio cycle. When the super high expansion ratio cycle is being performed, the routine proceeds to step 301 where the fuel consumption FbUL in the unlocked state and the fuel consumption FbL in the locked state are estimated. In the subsequent step 302, it is determined whether or not the fuel consumption FbL in the locked state is smaller than the fuel consumption FbL in the non-locked state. When FbL <FbUL, the routine proceeds to step 303 where the lock mechanism LM is activated and switched to the locked state. On the other hand, when FbL ≧ FbUL, the routine proceeds to step 304 where the lock mechanism LM is stopped and returned to the unlocked state.

  FIG. 21 shows a routine for executing the above-described EGR control. This routine is executed by interruption every predetermined time.

  Referring to FIG. 21, first, at step 400, it is judged if the first ultra-high expansion ratio cycle is being performed. When the first ultra-high expansion ratio cycle is being performed, the routine proceeds to step 401, where it is determined whether or not the lock state is established. When it is in the locked state, the routine proceeds to step 402, where it is determined whether or not the engine operating state is in the region AEGR. When the engine operating state is within the area AEGR, the routine proceeds to step 403 where the EGR action is performed. On the other hand, when the first ultra-high expansion ratio cycle is not performed in step 400, when the unlocked state is detected in step 401, and when the engine operating state is outside the region AEGR in step 402, the process proceeds to step 404. Advance and EGR action is stopped.

  In the example shown in FIG. 17, the region AEGR is defined by the engine torque Te and the engine speed Ne. In another embodiment, the region AEGR is defined only by the engine torque Te or only by the engine speed Ne.

  Next, another embodiment of EGR control will be described.

  In this other embodiment, as shown in FIG. 22, when the opening degree VEGR of the EGR control valve 92 is small, the allowable upper limit TeU is set larger than when the opening degree VEGR is large. That is, the region AEGR is enlarged as the amount of EGR gas decreases. This is because, when the EGR gas amount is small, the possibility of erroneous detection of the intake air amount is low even if the engine torque Te is large, that is, the throttle opening is large.

  In this other embodiment, as shown in FIG. 23, when the opening degree VEGR of the EGR control valve 92 is small, the allowable lower limit NeL is set smaller than when the opening degree VEGR is large. That is, the region AEGR is enlarged as the amount of EGR gas decreases. This is because when the amount of EGR gas is small, the influence of the EGR gas on the combustion stability is small.

  Further, as shown in FIG. 23, when the engine coolant temperature THW is high, the allowable lower limit NeL is set smaller than when the engine coolant temperature THW is low. That is, the region AEGR is enlarged as the engine coolant temperature THW increases. This is because when the engine coolant temperature THW is high, the influence of the EGR gas on the combustion stability is small.

  Further, when the engine coolant temperature THW is lower than the allowable lower limit THWL, the EGR action is stopped. This is because when the engine coolant temperature THW is excessively low, the influence of the EGR gas on the combustion stability is great.

  FIG. 24 shows a routine for executing the above-described EGR control. This routine is executed by interruption every predetermined time.

  Referring to FIG. 24, first, at step 400, it is judged if the first ultra-high expansion ratio cycle is being performed. When the first ultra-high expansion ratio cycle is being performed, the routine proceeds to step 401, where it is determined whether or not the lock state is established. When in the locked state, the routine proceeds to step 401a, where it is determined whether the engine coolant temperature THW is equal to or higher than the allowable lower limit THWL. When THW ≧ THWL, the routine proceeds to step 401b where the area AEGR is determined. That is, the allowable upper limit TeU and the allowable lower limit NeL are determined from FIG. 22 and FIG. In the subsequent step 402, it is determined whether or not the engine operating state is within the area AEGR. When the engine operating state is within the area AEGR, the routine proceeds to step 403 where the EGR action is performed. On the other hand, when the first ultra-high expansion ratio cycle is not performed in step 400, in the non-lock state in step 401, in TH401 <THWL in step 401a, and in the engine operation state in step 402 When it is outside the AEGR, the routine proceeds to step 404 where the EGR action is stopped.

  Next, still another embodiment according to the present invention will be described with reference to FIG.

  Immediately after the EGR action is stopped, EGR gas remains in the intake duct 43, the surge tank 41, the intake branch pipe 40, and the intake port 37, and this EGR gas is sequentially drawn into each cylinder. However, at this time, the amount of EGR gas sucked into each cylinder may vary. As a result, the combustion in each cylinder varies and the torque fluctuation may increase.

  On the other hand, if an ultra-high expansion ratio cycle is performed, combustion may become unstable. Further, in the second ultra-high expansion ratio cycle, the throttle valve 46 is held fully open, so that the EGR gas may flow back to the intake duct 43 upstream of the throttle valve 46. As a result, the EGR gas in the intake passage remains for a long time. This means that the period during which torque fluctuations may occur becomes longer.

  Therefore, in the embodiment shown in FIG. 25, even if an ultrahigh expansion ratio cycle should be performed, when the EGR action is stopped, the mechanical compression ratio is maintained below a predetermined compression ratio and the intake valve A normal cycle in which the intake air amount is controlled by temporarily controlling the throttle opening while maintaining the valve closing timing close to the intake bottom dead center is temporarily performed.

  That is, as shown in FIG. 25, when the EGR action is stopped at time ta1, the super high expansion ratio cycle is stopped and the normal cycle is performed even if the super high expansion ratio cycle should be performed. Next, at time ta2, that is, when a certain time dta elapses, the cycle is returned to the ultrahigh expansion ratio cycle. As a result, combustion is prevented from becoming unstable. Further, the EGR gas remaining in the intake passage is quickly removed. The fixed time dta is set to a time necessary for reducing the EGR gas remaining in the intake passage to almost zero.

  FIG. 26 shows a routine for executing the cycle control of the embodiment shown in FIG. This routine is executed by interruption every predetermined time.

  Referring to FIG. 26, first, at step 200, it is judged if the output Pe of the engine 1 is lower than the boundary output PY. Next, when Pe <PY, the routine proceeds to step 200a, where it is determined whether or not a fixed time dta has elapsed since the EGR action was stopped. When the fixed time dta has elapsed since the EGR action was stopped, that is, not immediately after the EGR action was stopped, the routine proceeds to step 201, where an ultra-high expansion ratio cycle is performed. On the other hand, when Pe ≧ PY in step 200 and when the fixed time dta has not elapsed since the EGR action was stopped in step 200a, that is, immediately after the EGR action was stopped, the routine proceeds to step 202, where Cycle is performed.

  Next, still another embodiment according to the present invention will be described with reference to FIG.

  When the EGR action is performed, the temperature of the exhaust gas is lowered. For this reason, for example, when the EGR action is performed for a long time, the temperature of the catalyst built in the catalytic converter 49 may be lower than its activation temperature.

  Therefore, in the embodiment shown in FIG. 27, when the exhaust gas temperature falls below a predetermined allowable lower limit during the EGR operation in the locked state, the EGR operation is stopped while being maintained in the locked state. If the exhaust gas temperature is lower than the allowable lower limit even when the EGR action is stopped, the locked state is switched to the unlocked state and the super high expansion ratio cycle is switched to the normal cycle. If the exhaust gas temperature is lower than the allowable lower limit even after switching to the normal cycle, temperature rise control is performed.

  That is, as shown in FIG. 27, when the temperature TEG of the exhaust gas falls below the allowable lower limit TEGL at time tb1, the EGR action is stopped while maintaining the locked state even if the EGR action should be performed. The

  When the EGR action is stopped, the exhaust gas temperature TEG should rise. Nevertheless, at time tb2, when the exhaust gas temperature TEG is lower than the allowable lower limit TEGL even if a certain time dtb1 has elapsed since the EGR action was stopped, the locked state To the unlocked state. Further, even if the super high expansion ratio cycle should be maintained, the super high expansion ratio cycle is switched to the normal cycle.

  When switched to the normal cycle, the exhaust gas temperature TEG should rise. Nevertheless, at time tb3, if the exhaust gas temperature TEG is lower than the allowable lower limit TEGL even after a certain time dtb2 has elapsed since switching to the normal cycle, the temperature increase control for increasing the exhaust gas temperature TEG is performed. Is done. This temperature increase control is performed, for example, by retarding the ignition timing. In another embodiment, the battery 19 is forcibly charged in the charge / discharge control of the battery 19 shown in FIG. As a result, the output Pe of the engine 1 is increased and the exhaust gas temperature TEG is increased.

  When the temperature increase control is performed, the exhaust gas temperature TEG increases. In the example shown in FIG. 27, the exhaust gas temperature TEG reaches the allowable lower limit TEGL at time tb4, and at this time, the temperature rise control is stopped. Further, while the EGR action is stopped, the normal cycle is returned to the ultra-high expansion ratio cycle and the unlocked state is returned to the locked state. Next, when a certain time dtb3 has elapsed, the EGR action is resumed.

  That is, in a normal cycle, the throttle opening is controlled in order to control the intake air amount. For this reason, as shown in FIG. 27, when switching to the normal cycle, the intake pressure Pin, which is the pressure in the intake passage downstream of the throttle valve 46, decreases. When the EGR action is resumed when the intake pressure Pin is low, that is, when the EGR control valve 92 is opened, the EGR gas suddenly flows into the surge tank 41, and the amount of EGR gas supplied to the engine 1 greatly increases. There is a fear. On the other hand, when the ultra-high expansion ratio cycle is performed, the intake pressure Pin becomes higher than when the normal cycle is performed.

  Therefore, in the embodiment shown in FIG. 27, when the exhaust gas temperature TEG becomes equal to or higher than the allowable lower limit TEGL, the super high expansion ratio cycle is first restored, and then the EGR action is resumed. As a result, a rapid increase in the amount of EGR gas supplied to the engine 1 is suppressed. The fixed time dtb3 is set to a time necessary for the intake pressure Pin to rise.

  FIG. 28 shows a routine for executing the exhaust gas temperature control of the embodiment shown in FIG. This routine is executed by interruption every predetermined time.

  Referring to FIG. 28, first, at step 500, it is judged if the exhaust gas temperature TEG is lower than the allowable lower limit TEGL. When TEG <TEGL, the routine proceeds to step 501 where it is judged if the flag XSEGR is set. This flag XSEGR is set when the EGR action should be stopped (XSEGR ← 1), and is reset otherwise (XSEGR ← 0). When the flag XSEGR is reset, the routine proceeds to step 502 where the flag XSEGR is set. Therefore, the EGR action is stopped while maintaining the locked state.

  When the flag XSEGR is set in step 501, the process proceeds to step 503, where it is determined whether or not a fixed time dtb1 has elapsed since the EGR action was stopped. When the fixed time dtb1 has not elapsed, the routine proceeds to step 502. When the predetermined time dtb1 has elapsed, the routine proceeds to step 504, where it is determined whether or not the flag XSSC is set. This flag XSSC is set when the super-high expansion ratio cycle is to be stopped (XSSC ← 1), and is reset otherwise (XSSC ← 0). When the flag XSSC is reset, the routine proceeds to step 505, where the flag XSSC is set. Therefore, it is switched to the normal cycle.

  When the flag XSSC is set at step 504, the routine proceeds to step 506, where it is determined whether or not a fixed time dtb2 has elapsed since the ultra-high expansion ratio cycle was stopped. When the fixed time dtb2 has not elapsed, the routine proceeds to step 505. When the fixed time dtb2 has elapsed, the routine proceeds to step 507, where the flag XT is set. This flag XT is set when temperature increase control is to be performed (XT ← 1), and is reset otherwise (XT ← 0). Accordingly, the temperature rise control is started.

  When TEG ≧ TEGL at step 500, the routine proceeds to step 508, where the flag XT is reset and the flag XSSC is reset. Accordingly, the temperature raising control is stopped and the operation is returned to the ultra-high expansion ratio cycle. In the subsequent step 509, it is determined whether or not a fixed time dtb3 has elapsed since returning to the ultra-high expansion ratio cycle. When the fixed time dtb3 has not elapsed, the processing cycle is ended. When the fixed time dtb3 has elapsed, the routine proceeds to step 510, where the flag XSEGR is reset. Therefore, the EGR action is resumed.

  FIG. 29 shows a routine for executing the cycle control of the embodiment shown in FIG. This routine is executed by interruption every predetermined time.

  Referring to FIG. 29, first, at step 200, it is judged if the output Pe of the engine 1 is lower than the boundary output PY. When Pe <PY, the routine proceeds to step 200b, where it is determined whether or not the flag XSSC is reset. When the flag XSSC is reset, the routine proceeds to step 201 where an ultra-high expansion ratio cycle is performed. On the other hand, when Pe ≧ PY at step 200 and when flag XSSC is set at step 200b, the routine proceeds to step 202 where a normal cycle is performed. In the following step 203, it is determined whether or not the flag XT is set. When the flag XT is reset, the processing cycle is terminated. When the flag XT is set, the routine proceeds to step 204 where temperature increase control is performed.

  FIG. 30 shows a routine for executing the EGR control of the embodiment shown in FIG. This routine is executed by interruption every predetermined time.

  Referring to FIG. 30, first, at step 400, it is judged if the first ultra-high expansion ratio cycle is being performed. When the first ultra-high expansion ratio cycle is being performed, the routine proceeds to step 401, where it is determined whether or not the lock state is established. When it is in the locked state, the routine proceeds to step 402, where it is determined whether or not the engine operating state is in the region AEGR. When the engine operating state is in the area AEGR, the routine proceeds to step 402a, where it is determined whether or not the flag XSEGR is reset. When the flag XSEGR is reset, the routine proceeds to step 403 where the EGR action is performed. On the other hand, when the first ultra-high expansion ratio cycle is not performed in step 400, when the unlocked state is detected in step 401, when the engine operating state is outside the region AEGR in step 402, and in step 402a When the flag XSEGR is set, the routine proceeds to step 404 where the EGR action is stopped.

  Next, another embodiment according to the present invention will be described.

  The EGR gas contains particulate matter made of, for example, solid carbon or hydrocarbon. When this particulate matter adheres to the EGR control valve 92, the EGR control valve 92 is fixed in a closed state or an opened state. There is a fear.

  When the EGR control valve 92 is fixed in the closed state, the EGR action is not performed even when the EGR action should be performed. In this case, the fuel consumption when the super high expansion ratio cycle is performed under the locked state is smaller than the fuel consumption when the normal cycle is performed.

  Therefore, when it is determined that the EGR control valve 92 is stuck in the closed state, as shown in FIG. 31, the ultra-high expansion ratio cycle is allowed and the locked state is allowed. That is, when the required output Pe of the engine 1 is lower than the boundary output PY, an ultra-high expansion ratio cycle is performed. Further, when the fuel consumption in the locked state is smaller than the fuel consumption in the non-locked state when the super high expansion ratio cycle is performed, the state is switched to the locked state. In other words, the cycle control and the lock control are not different from when the EGR control valve 92 is not fixed.

  On the other hand, when the EGR control valve 92 is stuck in the open state, the EGR action is performed even when the EGR action should be stopped. In this case, the erroneous detection of the intake air amount described above may occur.

  On the other hand, in a normal cycle, the intake air amount is controlled by controlling the throttle opening. For this reason, the throttle valve 46 is partially closed over a wide engine operation region. Therefore, the risk of the EGR gas reaching the intake air amount detector 47 is lower when the normal cycle is performed than when the ultrahigh expansion ratio cycle is performed.

  Therefore, when it is determined that the EGR control valve 92 is stuck in the open state, as shown in FIG. 32, the ultra-high expansion ratio cycle is prohibited and a normal cycle is performed. When a normal cycle is performed, the lock state is maintained.

  Whether or not the EGR control valve 92 is fixed is determined, for example, as follows. That is, the actual EGR gas amount is calculated from the total gas amount taken into the engine determined by the intake pressure and the engine speed and the amount of fresh air detected by the intake air amount detector 47. If the EGR gas is not supplied to the engine 1 when the EGR action is to be performed, it is determined that the EGR control valve 92 is stuck in the closed state. On the other hand, when EGR gas is supplied to the engine 1 when the EGR action should not be performed, it is determined that the EGR control valve 92 is stuck in the open state.

  FIG. 33 shows a routine for executing cycle control according to still another embodiment of the present invention. This routine is executed by interruption every predetermined time.

  Referring to FIG. 33, first, at step 200, it is judged if the output Pe of the engine 1 is lower than the boundary output PY. When Pe <PY, the routine proceeds to step 200c, where it is determined whether or not the EGR control valve 92 is stuck in the closed state. When the EGR control valve 92 is not fixed in the closed state, that is, when the EGR control valve 92 is not fixed or is fixed in the open state, the routine proceeds to step 201 where an ultra-high expansion ratio cycle is performed. . On the other hand, when Pe ≧ PY in step 200 and when the EGR control valve 92 is stuck in the closed state in step 200c, the routine proceeds to step 202 and a normal cycle is performed.

DESCRIPTION OF SYMBOLS 1 Spark ignition type engine 2 Output adjustment device 30 Crankcase 31 Cylinder block 34 Combustion chamber 36 Intake valve A Variable compression ratio mechanism B Variable valve timing mechanism MG1, MG2 Motor generator LM Lock mechanism

Claims (16)

An output adjustment device having a pair of motor generators and receiving an engine output and generating a vehicle driving output is provided, and the output adjustment device is configured to distribute the engine output torque to each motor generator. The output adjusting device further includes a lock mechanism that can be temporarily switched from an unlocked state in which one motor generator is allowed to rotate to a locked state in which one motor generator is prevented from rotating. A variable compression ratio mechanism capable of changing the mechanical compression ratio of the engine, a variable valve timing mechanism capable of controlling the closing timing of the intake valve, and an engine intake passage for controlling the intake air amount The throttle valve is provided, and the engine compression ratio is maintained at a predetermined compression ratio or higher in the engine. A first ultra-high expansion ratio cycle is performed in which the intake air amount is controlled by controlling the throttle opening while the closing timing of the intake valve is maintained on the side away from the intake bottom dead center. and comprising an exhaust passage and an intake passage capable of executing the exhaust gas recirculation action are connected to each other by an exhaust recirculation passage exhaust recirculation mechanism of the engine, exhaust gas recirculation action is performed when the locked state, an unlocked state An engine controller that sometimes stops the exhaust gas recirculation action . When even during locked state engine torque higher than the allowable upper limit is stopped exhaust gas recirculation action, claim 1 in which the exhaust recirculation action is performed when the engine torque even during the locked state is less than the allowable upper limit The engine control device described in 1. The exhaust gas recirculation mechanism has an exhaust gas recirculation control valve disposed in the exhaust gas recirculation passage, and controls the opening degree of the exhaust gas recirculation control valve to control the amount of exhaust gas recirculation gas. The engine control device according to claim 2 , wherein the allowable upper limit is set larger when the opening of the exhaust gas is smaller than when the opening of the exhaust gas recirculation control valve is large. The exhaust gas recirculation action is stopped when the engine speed is lower than the allowable lower limit even in the locked state, and the exhaust gas recirculation action is performed when the engine speed is higher than the allowable lower limit in the locked state. Item 4. The engine control device according to any one of Items 1 to 3 . The exhaust gas recirculation mechanism has an exhaust gas recirculation control valve disposed in the exhaust gas recirculation passage, and controls the opening degree of the exhaust gas recirculation control valve to control the amount of exhaust gas recirculation gas. The engine control device according to claim 4 , wherein the allowable lower limit is set lower when the opening of the exhaust gas is smaller than when the opening of the exhaust gas recirculation control valve is large. When the exhaust gas recirculation action is stopped, the mechanical compression ratio is maintained below the predetermined compression ratio, and the throttle opening is controlled while the intake valve closing timing is maintained close to the intake bottom dead center. The engine control apparatus according to any one of claims 1 to 5, wherein another cycle in which the intake air amount is controlled by the operation is temporarily performed. Claim 1, exhaust gas recirculation action while being maintained in a locked state when the exhaust gas temperature when the exhaust gas recirculation action even during the locked state is lowered beyond the allowable lower limit of the predetermined is stopped until 6 The engine control device according to one item. When the engine torque is lower than the limit value, the first super-high expansion ratio cycle is performed. When the engine torque is higher than the limit value, the mechanical compression ratio is maintained at a predetermined mechanical compression ratio and the throttle valve is The engine control according to any one of claims 1 to 7 , wherein a second ultra-high expansion ratio cycle in which an intake air amount is controlled by delaying a closing timing of the intake valve while being held in a fully open state is performed. apparatus. It is determined whether or not the exhaust gas recirculation control valve is stuck, and when it is determined that the exhaust gas recirculation control valve is stuck in the closed state, switching to the locked state is permitted and the first When it is determined that the high expansion ratio cycle or the second ultra-high expansion ratio cycle is permitted and the exhaust gas recirculation control valve is stuck in the open state, the engine is switched to the unlocked state and the mechanical compression ratio is Another cycle is performed in which the intake air amount is controlled by controlling the throttle opening while maintaining the compression ratio below a predetermined compression ratio and maintaining the closing timing of the intake valve close to the intake bottom dead center. The engine control device according to claim 8 .   An output adjustment device having a pair of motor generators and receiving an engine output and generating a vehicle driving output is provided, and the output adjustment device is configured to distribute the engine output torque to each motor generator. The output adjusting device further includes a lock mechanism that can be temporarily switched from an unlocked state in which one motor generator is allowed to rotate to a locked state in which one motor generator is prevented from rotating. A variable compression ratio mechanism capable of changing the mechanical compression ratio of the engine, a variable valve timing mechanism capable of controlling the closing timing of the intake valve, and an engine intake passage for controlling the intake air amount The throttle valve is provided, and the engine compression ratio is maintained at a predetermined compression ratio or higher in the engine. A first ultra-high expansion ratio cycle is performed in which the intake air amount is controlled by controlling the throttle opening while the closing timing of the intake valve is maintained on the side away from the intake bottom dead center. The exhaust passage and the intake passage of the engine are connected to each other by an exhaust recirculation passage, and an exhaust gas recirculation mechanism capable of performing an exhaust gas recirculation operation is provided, and the exhaust gas recirculation operation is performed in a locked state.
  Even when locked, the exhaust gas recirculation action is stopped when the engine torque is higher than the allowable upper limit, and when the engine torque is lower than the allowable upper limit, the engine control apparatus is configured to perform the exhaust gas recirculation action. .   The exhaust gas recirculation mechanism has an exhaust gas recirculation control valve disposed in the exhaust gas recirculation passage, and controls the opening degree of the exhaust gas recirculation control valve to control the amount of exhaust gas recirculation gas. The engine control device according to claim 10, wherein the allowable upper limit is set larger when the opening of the exhaust gas recirculation control valve is smaller than when the opening of the exhaust gas recirculation control valve is large.   An output adjustment device having a pair of motor generators and receiving an engine output and generating a vehicle driving output is provided, and the output adjustment device is configured to distribute the engine output torque to each motor generator. The output adjusting device further includes a lock mechanism that can be temporarily switched from an unlocked state in which one motor generator is allowed to rotate to a locked state in which one motor generator is prevented from rotating. A variable compression ratio mechanism capable of changing the mechanical compression ratio of the engine, a variable valve timing mechanism capable of controlling the closing timing of the intake valve, and an engine intake passage for controlling the intake air amount The throttle valve is provided, and the engine compression ratio is maintained at a predetermined compression ratio or higher in the engine. A first ultra-high expansion ratio cycle is performed in which the intake air amount is controlled by controlling the throttle opening while the closing timing of the intake valve is maintained on the side away from the intake bottom dead center. The exhaust passage and the intake passage of the engine are connected to each other by an exhaust recirculation passage, and an exhaust gas recirculation mechanism capable of performing an exhaust gas recirculation operation is provided, and the exhaust gas recirculation operation is performed in a locked state.
  Even in the locked state, the exhaust gas recirculation action is stopped when the engine speed is lower than the allowable lower limit, and the exhaust gas recirculation action is performed when the engine speed is higher than the allowable lower limit value in the locked state. Control device.   The exhaust gas recirculation mechanism has an exhaust gas recirculation control valve disposed in the exhaust gas recirculation passage, and controls the opening degree of the exhaust gas recirculation control valve to control the amount of exhaust gas recirculation gas. The engine control apparatus according to claim 12, wherein the allowable lower limit is set lower when the opening of the exhaust gas recirculation control valve is smaller than when the opening of the exhaust gas recirculation control valve is large.   An output adjustment device having a pair of motor generators and receiving an engine output and generating a vehicle driving output is provided, and the output adjustment device is configured to distribute the engine output torque to each motor generator. The output adjusting device further includes a lock mechanism that can be temporarily switched from an unlocked state in which one motor generator is allowed to rotate to a locked state in which one motor generator is prevented from rotating. A variable compression ratio mechanism capable of changing the mechanical compression ratio of the engine, a variable valve timing mechanism capable of controlling the closing timing of the intake valve, and an engine intake passage for controlling the intake air amount The throttle valve is provided, and the engine compression ratio is maintained at a predetermined compression ratio or higher in the engine. A first ultra-high expansion ratio cycle is performed in which the intake air amount is controlled by controlling the throttle opening while the closing timing of the intake valve is maintained on the side away from the intake bottom dead center. The exhaust passage and the intake passage of the engine are connected to each other by an exhaust recirculation passage, and an exhaust gas recirculation mechanism capable of performing an exhaust gas recirculation operation is provided, and the exhaust gas recirculation operation is performed in a locked state.
  When the exhaust gas recirculation action is stopped, the mechanical compression ratio is maintained below the predetermined compression ratio, and the throttle opening is controlled while the intake valve closing timing is maintained close to the intake bottom dead center. An engine control device in which another cycle in which the intake air amount is controlled is temporarily performed.   An output adjustment device having a pair of motor generators and receiving an engine output and generating a vehicle driving output is provided, and the output adjustment device is configured to distribute the engine output torque to each motor generator. The output adjusting device further includes a lock mechanism that can be temporarily switched from an unlocked state in which one motor generator is allowed to rotate to a locked state in which one motor generator is prevented from rotating. A variable compression ratio mechanism capable of changing the mechanical compression ratio of the engine, a variable valve timing mechanism capable of controlling the closing timing of the intake valve, and an engine intake passage for controlling the intake air amount The throttle valve is provided, and the engine compression ratio is maintained at a predetermined compression ratio or higher in the engine. A first ultra-high expansion ratio cycle is performed in which the intake air amount is controlled by controlling the throttle opening while the closing timing of the intake valve is maintained on the side away from the intake bottom dead center. The exhaust passage and the intake passage of the engine are connected to each other by an exhaust recirculation passage, and an exhaust gas recirculation mechanism capable of performing an exhaust gas recirculation operation is provided, and the exhaust gas recirculation operation is performed in a locked state.
  An engine control device in which the exhaust gas recirculation operation is stopped while being maintained in the locked state when the exhaust gas temperature falls below a predetermined allowable lower limit during the exhaust gas recirculation operation in the locked state.   An output adjustment device having a pair of motor generators and receiving an engine output and generating a vehicle driving output is provided, and the output adjustment device is configured to distribute the engine output torque to each motor generator. The output adjusting device further includes a lock mechanism that can be temporarily switched from an unlocked state in which one motor generator is allowed to rotate to a locked state in which one motor generator is prevented from rotating. A variable compression ratio mechanism capable of changing the mechanical compression ratio of the engine, a variable valve timing mechanism capable of controlling the closing timing of the intake valve, and an engine intake passage for controlling the intake air amount The throttle valve is provided, and the engine compression ratio is maintained at a predetermined compression ratio or higher in the engine. A first ultra-high expansion ratio cycle is performed in which the intake air amount is controlled by controlling the throttle opening while the closing timing of the intake valve is maintained on the side away from the intake bottom dead center. The exhaust passage and the intake passage of the engine are connected to each other by an exhaust recirculation passage, and an exhaust gas recirculation mechanism capable of performing an exhaust gas recirculation operation is provided, and the exhaust gas recirculation operation is performed in a locked state.
  When the engine torque is lower than the limit value, the first super-high expansion ratio cycle is performed. When the engine torque is higher than the limit value, the mechanical compression ratio is maintained at a predetermined mechanical compression ratio and the throttle valve is A second ultra-high expansion ratio cycle is performed in which the intake air amount is controlled by delaying the closing timing of the intake valve while being held in the fully open state;
  It is determined whether or not the exhaust gas recirculation control valve is stuck, and when it is determined that the exhaust gas recirculation control valve is stuck in the closed state, switching to the locked state is permitted and the first When it is determined that the high expansion ratio cycle or the second ultra-high expansion ratio cycle is permitted and the exhaust gas recirculation control valve is stuck in the open state, the engine is switched to the unlocked state and the mechanical compression ratio is Another cycle is performed in which the intake air amount is controlled by controlling the throttle opening while maintaining the compression ratio below a predetermined compression ratio and maintaining the closing timing of the intake valve close to the intake bottom dead center. Engine control device.

Images