JP2005248783A - Control device of internal combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To prevent torque fluctuation caused by introduction of evaporating fuel in connection with a control device of an internal combustion engine. <P>SOLUTION: At the time of introduction of evaporating fuel into an intake passage, torque corresponding values such as indicated torque corresponding to generated torque of respective cylinders are calculated from operation data of the internal combustion engine such as crank angle signals. As variation among distribution ratios of evaporating fuel appears as variation among torque of the respective cylinders at the time of introduction of evaporating fuel, a distribution ratio corresponding value corresponding to a distribution ratio of evaporating fuel for every cylinder is calculated from the torque corresponding values of the respective cylinders at the time of introduction of evaporating fuel. A correction value of an engine control parameter such as a fuel injection amount is calculated based on the distribution ratio corresponding value. <P>COPYRIGHT: (C)2005,JPO&NCIPI

Description

  The present invention relates to a control technique for an internal combustion engine, and more particularly to a control technique for an internal combustion engine that is suitable for use in suppressing torque fluctuations when evaporative fuel is introduced.

  An internal combustion engine for a vehicle is provided with a canister that adsorbs and stores evaporated fuel generated in a fuel tank. During the operation of the internal combustion engine, the evaporated fuel stored in the canister is purged from the canister using the negative pressure in the intake passage, is discharged into the intake passage together with fresh air, and is combusted in the combustion chamber.

  Further, the internal combustion engine determines a target air-fuel ratio for realizing a desired combustion state, and controls the fuel supply amount so that the actual air-fuel ratio becomes this target air-fuel ratio. When evaporative fuel is introduced into the combustion chamber for combustion processing as described above, it is necessary to adjust engine control parameters such as the fuel injection amount in consideration of fluctuations in the air-fuel ratio accompanying the introduction of evaporative fuel.

Conventionally, various techniques have been proposed for controlling an internal combustion engine when performing purge of evaporated fuel. For example, in Patent Document 1, in an internal combustion engine that performs air-fuel ratio feedback control by setting an air-fuel ratio feedback correction value for each cylinder or cylinder group, the amount of evaporated fuel is determined from the ratio of the air-fuel ratio feedback correction value for each cylinder or cylinder group. A technique is disclosed in which a distribution rate is obtained and fuel is supplied so that the air-fuel ratio feedback correction value for each cylinder or cylinder group becomes equal based on the obtained distribution rate of evaporated fuel.
JP-A-9-88659 JP 2003-214265 A

  However, since the ratio of the air-fuel ratio feedback correction value for each cylinder or each cylinder group does not necessarily match the distribution ratio of the evaporated fuel, it is difficult to accurately estimate the distribution ratio of the evaporated fuel by the method as in the above prior art. . Further, in an internal combustion engine in which an air-fuel ratio sensor is mounted for each bank according to the above-described prior art, the estimation accuracy of the evaporated fuel distribution rate is further reduced unless the same output can be ensured between the air-fuel ratio sensors. For this reason, in the above prior art, it is not possible to accurately correct the fuel injection amount according to the distribution variation of the evaporated fuel, and as a result, there is a risk of inducing variation in the generated torque between the cylinders.

  Further, in the above prior art, even if the evaporated fuel distribution rate can be estimated with a certain degree of accuracy from the air-fuel ratio feedback correction value, the evaporated fuel distribution rate is obtained in the region where the concentration of evaporated fuel is unstable immediately after the start of the purge. I can't. The distribution ratio of the evaporated fuel is calculated after the concentration of the evaporated fuel is stabilized and the air-fuel ratio feedback correction value is stabilized. Furthermore, there is a time delay corresponding to the exhaust gas transport delay from the stabilization of the evaporated fuel concentration to the stabilization of the air-fuel ratio feedback correction value. For this reason, in the above prior art, the correction of the fuel injection amount is delayed with respect to the introduction of the evaporated fuel, and there is a possibility that the torque change caused by the introduction of the evaporated fuel cannot be suppressed for a while after the purge starts.

  Variations in torque generated between cylinders and torque changes in specific cylinders appear as torque fluctuations in the internal combustion engine. For this reason, in the prior art, the torque fluctuation of the internal combustion engine occurs with the introduction of the evaporated fuel, and the drivability of the vehicle may be deteriorated.

  The present invention has been made to solve the above-described problems, and an object of the present invention is to provide a control device for an internal combustion engine that can suppress torque fluctuations of the internal combustion engine caused by the introduction of evaporated fuel.

In order to achieve the above object, a first invention is a control device for an internal combustion engine,
Evaporative fuel introduction means for introducing evaporative fuel generated in the fuel tank into the intake passage of the internal combustion engine;
Torque corresponding value calculating means for obtaining a torque corresponding value corresponding to the generated torque of each cylinder from the operation data of the internal combustion engine;
A distribution rate corresponding value calculating means for obtaining a distribution rate corresponding value corresponding to the distribution rate of each evaporated fuel cylinder based on the torque corresponding value of each cylinder at the time of evaporative fuel introduction;
Control means for controlling the internal combustion engine in accordance with engine control parameters set for each cylinder;
Correction value calculation means for obtaining a correction value of the engine control parameter based on the distribution rate corresponding value;
It is characterized by having.

In order to achieve the above object, a second invention is a control device for an internal combustion engine,
Evaporative fuel introduction means for introducing evaporative fuel generated in the fuel tank into the intake passage of the internal combustion engine;
Torque corresponding value calculating means for obtaining a torque corresponding value corresponding to the generated torque of each cylinder from the operation data of the internal combustion engine;
A concentration corresponding value corresponding to the concentration of the evaporated fuel introduced into the specific cylinder based on a difference between the torque corresponding value of the specific cylinder at the time of evaporating fuel introduction and the torque corresponding value of the specific cylinder before the evaporative fuel introduction. Concentration corresponding value calculating means for obtaining
Control means for controlling the internal combustion engine in accordance with engine control parameters set for each cylinder;
Correction value calculation means for obtaining a correction value of the engine control parameter for the specific cylinder based on the concentration correspondence value;
It is characterized by having.

  In the first invention, the variation in the distribution ratio of the evaporated fuel appears as a variation in the torque of each cylinder when the evaporated fuel is introduced. Accordingly, by obtaining the torque corresponding value corresponding to the generated torque of each cylinder from the operation data of the internal combustion engine when evaporating fuel is introduced, the distribution ratio of the evaporated fuel for each cylinder can be estimated from the torque corresponding value of each cylinder. . According to the first aspect of the invention, since the correction value of the engine control parameter is calculated based on the distribution rate corresponding value corresponding to the distribution rate of the evaporated fuel for each cylinder, it is accurately determined according to the variation in the distribution rate of the evaporated fuel. It is possible to adjust the engine control parameter for each cylinder, and to effectively suppress the variation in the torque generated between the cylinders when the evaporated fuel is introduced. As a result, it is possible to suppress torque fluctuations of the internal combustion engine due to variations in torque generated between the cylinders.

  In the second invention, the concentration of the evaporated fuel introduced into the specific cylinder appears as a torque difference between the generated torque when the evaporated fuel is not introduced and the generated torque when the evaporated fuel is introduced. This torque difference fluctuates according to the fluctuation of the concentration in the region where the concentration of the evaporated fuel is unstable, such as immediately after the introduction of the evaporated fuel. Therefore, the vapor corresponding value corresponding to the generated torque of the specific cylinder is calculated when the evaporated fuel is introduced, and is compared with the torque corresponding value detected before the vaporized fuel is introduced. Can be estimated accurately and promptly. According to the second invention, since the correction value of the engine control parameter for the specific cylinder is calculated based on the concentration corresponding value corresponding to the concentration of the evaporated fuel, the concentration immediately after the introduction of the evaporated fuel is in an unstable region. Even in such a case, the engine control parameter of the specific cylinder can be accurately adjusted according to the concentration of the evaporated fuel, and the torque change of the specific cylinder caused by the introduction of the evaporated fuel can be quickly suppressed. Thereby, it becomes possible to suppress the torque fluctuation of the internal combustion engine due to the torque change of the specific cylinder.

Embodiment 1 FIG.
Hereinafter, Embodiment 1 of the present invention will be described with reference to FIGS.

  FIG. 1 is a schematic configuration diagram of an engine system to which a control device according to Embodiment 1 of the present invention is applied. The internal combustion engine 2 according to the present embodiment is a spark ignition type four-stroke engine, and has a plurality of cylinders (not shown). An intake passage 4 and an exhaust passage 6 are connected to the combustion chamber 16 of each cylinder. An intake valve 8 for controlling the communication state is provided at a connection portion between the combustion chamber 16 and the intake passage 4, and an exhaust valve 10 for controlling the communication state is provided at a connection portion between the combustion chamber 16 and the exhaust passage 6. It has been. A spark plug 12 is attached to the top of the combustion chamber 16. A crank angle sensor 32 that outputs a signal at a predetermined crank angle position is attached in the vicinity of the crankshaft 22.

  The exhaust passage 6 is provided with an A / F sensor 34 that detects the air-fuel ratio of the exhaust gas. Further, a catalyst device (not shown) for purifying harmful components in the exhaust gas is provided downstream thereof.

  The intake passage 4 is provided with an air cleaner 20 at one end thereof. An air flow meter 36 for measuring the intake air flow rate is disposed immediately downstream of the air cleaner 20. Further, an electronically controlled throttle valve 18 that adjusts the amount of fresh air flowing into the combustion chamber 16 is disposed downstream of the air flow meter 36. The throttle valve 18 incorporates a throttle sensor 38 that outputs a signal corresponding to its opening. The front end of the intake passage 4 is branched to introduce air into the combustion chamber 16 of each cylinder, and a fuel injection valve 14 for supplying fuel to the combustion chamber 16 is attached to each passage.

  The fuel injected from the fuel injection valve 14 is supplied from the fuel tank 50 through a fuel passage (not shown). The fuel tank 50 is connected to a vapor passage 44 for extracting evaporated fuel generated inside. One end of the vapor passage 44 is connected to a canister 40 filled with activated carbon for adsorbing evaporated fuel. Therefore, the evaporated fuel generated inside the fuel tank 50 reaches the canister 40 through the vapor passage 44 and is adsorbed inside the canister 40.

  An atmospheric air introduction passage 46 is connected to the canister 40, and a purge passage 42 communicating with the intake passage 4 is connected downstream of the throttle valve 18. The purge passage 42 is provided with a purge control valve 48 for controlling the flow rate of gas flowing through the purge passage 42. The purge control valve 48 is a control valve that realizes an arbitrary opening degree by duty control.

  The internal combustion engine 2 includes an ECU (Electronic Control Unit) 30 as its control device. The ECU 30 comprehensively controls various devices related to the operation of the internal combustion engine 2 based on the operation data of the internal combustion engine 2 detected by a plurality of sensors. A crank angle sensor 32, an A / F sensor 34, an air flow meter 36, and a throttle sensor 38 are connected to the input side of the ECU 30, and a spark plug 12, a fuel injection valve 14, and a purge control valve 48 are connected to the output side. Yes. The ECU 30 receives operation data from the sensors 32, 34, 36, and 38 and supplies drive signals to the devices 12, 14, and 48. In addition to the sensors 32, 34, 36, and 38 and the devices 12, 14, and 48, a plurality of sensors and devices are connected to the ECU 30, but the description thereof is omitted here.

  One of the controls of the internal combustion engine 2 performed by the ECU 30 is a purge control for purging the evaporated fuel adsorbed by the canister 40. This purge control is executed by appropriately driving the purge control valve 48 when the predetermined purge condition is satisfied during operation of the internal combustion engine 2. When the purge control valve 48 is driven by a duty, the negative pressure in the intake passage 4 of the internal combustion engine 2 is guided to the canister 40. As a result, the evaporated fuel in the canister 40 is discharged into the purge passage 42 along with the fresh air sucked from the air introduction passage 46. The discharged evaporated fuel is introduced into the intake passage 6 through the purge passage 42 and is combusted in the combustion chamber 16.

  When the purge control is executed, the air-fuel ratio of the air-fuel mixture sucked into the combustion chamber 16 varies due to the influence of the evaporated fuel introduced into the intake passage 6. At this time, if the distribution ratio of the evaporated fuel for each cylinder is equal, the air-fuel ratio of each cylinder can be brought close to the target air-fuel ratio by the air-fuel ratio feedback control based on the output signal of the A / F sensor 34. However, in reality, the distribution ratio of the evaporated fuel varies due to intake air interference between cylinders and the like. For this reason, even if the exhaust air-fuel ratio is made to coincide with the target air-fuel ratio by air-fuel ratio feedback control, the mixture air-fuel ratio varies among the cylinders, and a torque difference occurs between the cylinders due to the variation. FIG. 2 is a graph showing the torque behavior of each cylinder during purge execution. A torque difference between the cylinders as shown in FIG. 2 appears as a torque fluctuation of the internal combustion engine 2 and deteriorates the drivability of the vehicle. For this reason, the ECU 30 is configured to also execute torque fluctuation suppression control described below when executing the purge control.

  The torque fluctuation suppression control executed in the present embodiment suppresses the variation in generated torque between cylinders due to the variation in the distribution ratio of the evaporated fuel by correcting the fuel injection amount in advance in accordance with the execution of the purge control. It is intended to suppress the torque fluctuation of the internal combustion engine accompanying the introduction of the evaporated fuel. The fuel injection amount is corrected by multiplying the fuel injection amount by a special correction factor for purge control separately from the correction factor used in the air-fuel ratio feedback.

  FIG. 3 is a flowchart for explaining a correction coefficient calculation method used in torque fluctuation suppression control. In the routine shown in FIG. 3, it is first determined whether or not purge control is being executed (step 100). Whether or not the purge control is being executed can be determined, for example, by whether or not a drive signal for duty-controlling the purge control valve 48 is being output.

  If it is determined in step 100 that the purge control is being executed, then in step 102, a torque corresponding value corresponding to the generated torque for each cylinder is calculated (step 102). For example, as described below, the torque correspondence value can be calculated from the operation data (crank angle signal) supplied from the crank angle sensor 32 according to the equation of motion.

The following equations (1) and (2) are equations for calculating torque from the crank angle signal supplied from the crank angle sensor 32.
Ti = J × (dω / dt) + Tf + Tl (1)
Ti = Tgas + Tinertia (2)
In the above equations (1) and (2), Ti is the indicated torque generated on the crankshaft by engine combustion. Here, the right side of equation (2) indicates the torque that generates the indicated torque Ti, and the right side of equation (1) indicates the torque that consumes the indicated torque Ti.

  In the right side of equation (1), J is the moment of inertia of the drive member driven by the combustion of the air-fuel mixture, dω / dt is the angular acceleration of the crankshaft, Tf is the friction torque of the drive unit, and Tl is the load received from the road surface during travel Torque. Here, J × (dω / dt) is a dynamic loss torque (= Tac) resulting from the angular acceleration of the crankshaft. The friction torque Tf is a torque due to mechanical friction of each fitting portion such as friction between the piston and the inner wall of the cylinder, and includes torque due to mechanical friction of auxiliary machinery. The load torque Tl is a torque due to a disturbance such as a road surface condition during traveling. Since the shift gear is in the neutral state during cold first idling, Tl = 0 is assumed in the following description.

  In the right side of the equation (2), Tgas represents torque due to cylinder cylinder gas pressure, and Tinertia represents inertial torque due to reciprocating inertial mass such as a piston. The torque Tgas due to the in-cylinder gas pressure is a torque generated by the combustion of the air-fuel mixture in the cylinder. In order to accurately estimate the combustion state, it is necessary to obtain the torque Tgas by the cylinder gas pressure.

  As shown in the equation (1), the indicated torque Ti can be obtained as the sum of dynamic loss torque J × (dω / dt) due to angular acceleration, friction torque Tf, and load torque Tl. However, as shown in the equation (2), the indicated torque Ti and the torque Tgas due to the in-cylinder gas pressure do not match, so the combustion state cannot be accurately estimated from the indicated torque Ti.

  FIG. 4 is a characteristic diagram showing the relationship between each torque and crank angle in equation (2). In FIG. 4, the vertical axis indicates the magnitude of each torque, the horizontal axis indicates the crank angle, the one-dot chain line in FIG. 4 indicates the indicated torque Ti, the solid line indicates the torque Tgas due to the in-cylinder gas pressure, and the broken line reciprocates. Inertia torque Tinertia due to inertial mass is shown. Here, FIG. 4 shows the characteristics in the case of four cylinders. TDC and BDC in FIG. 4 are the top dead center (TDC) or bottom dead center (TDC) of one cylinder among the four cylinders ( The crank angle (0 °, 180 °) in the position of (BDC) is shown. When the internal combustion engine has four cylinders, an explosion stroke is performed for each cylinder every time the crankshaft rotates 180 °, and torque characteristics from TDC to BDC in FIG.

  As shown by the solid line in FIG. 4, the torque Tgas due to the in-cylinder gas pressure rapidly increases and decreases between TDC and BDC. Here, the rapid increase in Tgas is due to the explosion of the air-fuel mixture in the combustion chamber during the explosion stroke. After the explosion, Tgas decreases and becomes a negative value due to the influence of the cylinders in other compression strokes or exhaust strokes. Thereafter, when the crank angle reaches BDC, the change in volume of the cylinder becomes zero and Tgas becomes zero.

  On the other hand, the inertia torque Tinertia due to the reciprocating inertial mass is an inertial torque generated by the inertial mass of a reciprocating member such as a piston, regardless of the torque Tgas due to the in-cylinder gas pressure. The reciprocating member repeatedly accelerates and decelerates, and Tinertia always occurs as long as the crank rotates, even if the angular velocity is constant. As shown by the broken line in FIG. 4, the member that reciprocates is stopped at the position where the crank angle is TDC, and Tinertia = 0. When the crank angle advances from TDC toward BDC, the reciprocating member starts to move from the stopped state. At this time, Tinertia increases in the negative direction due to the inertia of these members. When the crank angle reaches around 90 °, the reciprocating members are moving at a predetermined speed, so that the crankshaft rotates due to the inertia of these members. Therefore, Tinertia changes from a negative value to a positive value between TDC and BDC. After that, when the crank angle reaches BDC, the reciprocating member stops and Tinertia = 0.

  As shown in the equation (2), the indicated torque Ti is the sum of the torque Tgas caused by the in-cylinder gas pressure and the inertia torque Tinertia caused by the reciprocating inertial mass. For this reason, as shown by the one-dot chain line in FIG. 4, between TDC and BDC, the indicated torque Ti increases due to an increase in Tgas due to the explosion of the mixture, once decreases, and then increases again due to Tinertia. The behavior is shown.

  However, when attention is paid to a section with a crank angle of 180 ° from TDC to BDC, the average value of inertia torque Tinertia due to reciprocating inertia mass in this section is zero. This is because a member having a reciprocating inertia mass moves in the opposite direction at a crank angle of 0 ° to 90 ° and a crank angle of 90 ° to 180 °. Therefore, when the torques in the equations (1) and (2) are calculated as average values from TDC to BDC, the inertia torque Tinertia = 0 by the reciprocating inertia mass can be calculated. As a result, the influence of the inertia torque Tinertia due to the reciprocating inertia mass on the indicated torque Ti can be eliminated, and an accurate combustion state can be easily estimated.

  Then, when the average value of each torque is obtained in the section from TDC to BDC, the average value of Tinertia becomes 0. From the equation (2), the average value of the indicated torque Ti and the average of the torque Tgas due to the in-cylinder gas pressure are obtained. The value becomes equal. Therefore, the combustion state can be accurately estimated based on the indicated torque Ti.

  Further, when the average value of the angular acceleration of the crankshaft in the section from TDC to BDC is obtained, the average value of Tinertia in this section is 0, so that the influence of the reciprocating inertia mass on the angular acceleration is eliminated. Can be requested. Therefore, it is possible to calculate the angular acceleration caused only by the combustion state, and it is possible to accurately estimate the combustion state based on the angular acceleration.

  Next, a method for calculating each torque on the right side of the equation (1) will be described. First, a method of calculating dynamic loss torque Tac = J × (dω / dt) resulting from angular acceleration will be described. FIG. 5 is a schematic diagram illustrating a method for obtaining the angular acceleration of the crankshaft, and is a diagram illustrating a crank angle signal and torque calculation timing. As shown in FIG. 5, in the present embodiment, a crank angle signal is supplied from the crank angle sensor 32 every 10 ° of rotation of the crankshaft.

The ECU 2 calculates the dynamic loss torque Tac caused by the angular acceleration as an average value from TDC to BDC. For this purpose, the apparatus of this embodiment obtains angular velocities ω ₀ (k) and ω ₀ (k + 1) at two crank angle positions of TDC and BDC, and the crankshaft rotates from TDC to BDC at the same time. Time Δt (k) is obtained.

When obtaining the angular velocity ω ₀ (k), for example, as shown in FIG. 5, the times Δt ₀ (k) and Δt ₁₀ (k) during which the crank angle is rotated by 10 ° forward and backward from the TDC position are calculated. It is detected from the crank angle sensor 32. Since the crankshaft rotates 20 ° during the time Δt ₀ (k) + Δt ₁₀ (k), ω ₀ (k) = (20 / (Δt ₀ (k) + Δt ₁₀ (k)) × ( π / 180) can be calculated to calculate ω ₀ (k) [rad / s] Similarly, when calculating ω ₀ (k + 1), the crank angle is 10 ° forward and backward from the BDC position. Times Δt ₀ (k + 1) and Δt ₁₀ (k + 1) during rotation are detected, and ω ₀ (k + 1) = (20 / (Δt ₀ (k + 1) + Δt ₁₀ ( k + 1))) × (π / 180) can be calculated to calculate ω ₀ (k + 1) [rad / s], and angular velocities ω ₀ (k) and ω ₀ (k + 1) were obtained. After that, (ω ₀ (k + 1) −ω ₀ (k)) / Δt (k) is calculated, and the average value of the angular acceleration during the rotation of the crankshaft from TDC to BDC is calculated.

  After the average value of angular acceleration is obtained, the average value of angular acceleration and the moment of inertia J are multiplied according to the right side of equation (1). As a result, an average value of dynamic loss torque J × (dω / dt) while the crankshaft rotates from TDC to BDC can be calculated. The inertia moment J of the drive unit is obtained in advance from the inertia mass of the drive component.

  Next, a method for calculating the friction torque Tf will be described. FIG. 6 is a map showing the relationship between the friction torque Tf, the engine speed Ne of the internal combustion engine, and the coolant temperature thw. In FIG. 6, the friction torque Tf, the engine speed Ne, and the coolant temperature thw are average values when the crankshaft rotates from TDC to BDC. Further, the cooling water temperature becomes higher in the order of thw1 → thw2 → thw3. As shown in FIG. 6, the friction torque Tf increases as the engine speed (Ne) increases, and increases as the cooling water temperature thw decreases. In the map of FIG. 6, the engine rotational speed Ne and the coolant temperature thw are varied as parameters, the friction torque Tf generated when the crankshaft is rotated from TDC to BDC is measured, and the average value is calculated in advance. Create it. When estimating the combustion state, the average value of the cooling water temperature and the average value of the engine speed in the section from TDC to BDC are applied to the map of FIG. 6 to obtain the average value of the friction torque Tf. At this time, the coolant temperature is detected from a water temperature sensor (not shown), and the engine speed is detected from the crank angle sensor 32.

  The behavior of the friction torque Tf accompanying the variation of the crank angle is very complicated and has a large variation. However, since the behavior of the friction torque Tf mainly depends on the speed of the piston, the average value of the friction torque Tf for each section where the average value of the inertia torque Tinertia due to the reciprocating inertia mass is zero is substantially constant. Therefore, by obtaining the average value of the friction torque Tf for each section (TDC → BDC) where the average value of the inertia torque Tinertia due to the reciprocating inertia mass is 0, the friction torque Tf showing a complex instantaneous behavior can be obtained with high accuracy. it can. Further, by making the friction torque Tf an average value for each section, the map shown in FIG. 6 can be created accurately.

  Further, as described above, the friction torque Tf includes torque due to friction of auxiliary machinery. Here, the value of the torque due to the friction of the auxiliary machines varies depending on whether or not the auxiliary machines are operating. For example, the rotation of the engine is transmitted to a compressor of an air conditioner, which is one of the auxiliary machines, by a belt or the like, and torque due to friction is generated even when the air conditioner is not actually operating.

  On the other hand, when the auxiliary machines are operated, for example, when the air conditioner switch is turned on, the torque consumed by the compressor becomes larger than in the state where the air conditioner is not operated. For this reason, the torque due to the friction of the auxiliary machinery increases, and the value of the friction torque Tf also increases. Therefore, in order to accurately determine the friction torque Tf, the operation state of the auxiliary machinery is detected, and when the auxiliary machinery is switched on, the value of the friction torque Tf obtained from the map of FIG. It is desirable to correct.

  It is more preferable to correct the friction torque Tf in consideration of the difference between the temperature of the portion where the friction torque Tf is actually generated and the cooling water temperature at the time of extremely cold start. In this case, it is desirable to perform correction in consideration of the engine start time after the cold start, the in-cylinder inflow fuel amount, and the like.

  In the present embodiment, the above indicated torque (also referred to as estimated indicated torque) Ti is used as a torque corresponding value corresponding to the generated torque of each cylinder. The ECU 30 calculates the indicated torque for a total of n cycles (for example, 10 cycles) for each cylinder by the above calculation method. The calculated indicated torque is stored in a memory provided in the ECU 30. The calculated n cycle indicated torque is averaged for each cylinder, and the average indicated torque for each cylinder is calculated.

  If the average indicated torque for n cycles for each cylinder is calculated in step 102, then the average indicated torque for all cylinders for all cylinders (average indicated torque for all cylinders) is calculated (step 104). That is, the average indicated torque for each cylinder calculated in step 102 is further averaged among the cylinders.

  In the next step 106, the difference (or deviation rate) between the average indicated torque calculated in step 102 and the average indicated torque of all cylinders calculated in step 104 is obtained for each cylinder, and the calculated torque difference ( Alternatively, it is determined for each cylinder whether or not the absolute value of the deviation rate is greater than a predetermined determination value. The variation in torque between cylinders that occurs when evaporative fuel is introduced is due to the variation in the distribution rate of the evaporated fuel, and the tendency of the variation in the distribution rate and the variation in the torque coincide with each other. In this embodiment, the variation in torque is expressed as a torque difference (or divergence rate) for each cylinder, and this is used as a distribution rate corresponding value corresponding to the distribution rate of the evaporated fuel for each cylinder.

  As a result of the determination in step 106, the correction coefficient of the fuel injection amount corresponding to the torque difference (or deviation rate) is calculated from a map (not shown) or the like for the cylinder whose absolute value of torque difference (or deviation rate) is larger than the determination value. (Step 108). On the other hand, the correction coefficient for the fuel injection amount is not calculated for the cylinder whose absolute value of the torque difference (or deviation rate) is equal to or smaller than the determination value.

  The ECU 30 corrects the fuel injection amount of each cylinder using the correction coefficient calculated by the correction coefficient calculation routine. Further, the ECU 30 stores the calculated correction coefficient for each cylinder in a memory (not shown). The ECU 30 learns the value of the correction coefficient for each cylinder, and updates the correction coefficient stored in the memory to a new value every time a new correction coefficient is calculated. When the next purge control is executed, the ECU 30 controls the fuel injection valve 14 so as to supply the cylinder with a fuel injection amount multiplied by a correction coefficient in advance.

  According to the torque fluctuation suppression control described above, the fuel injection amount for each cylinder is corrected based on the torque difference (or deviation rate) between the cylinders corresponding to the distribution ratio of the evaporated fuel. The fuel injection amount for each cylinder can be accurately adjusted according to the variation. Further, since the variation in the distribution ratio of the evaporated fuel is basically caused by the hardware specifications of the internal combustion engine 2, the distribution state of the distribution ratio does not change greatly every purge. According to the torque fluctuation suppression control, since the correction coefficient calculated for correcting the fuel injection amount is learned and used in the next purge control, the fuel corresponding to the variation in the distribution ratio of the evaporated fuel from the start of the purge. Injection can be realized for each cylinder. From the above, according to the control device of the present embodiment, it is possible to effectively suppress the variation in the torque generated between the cylinders when the evaporated fuel is introduced, thereby suppressing the torque fluctuation of the internal combustion engine 2. Is possible.

  In the above-described embodiment, the “torque corresponding value calculating means” according to the first aspect of the present invention is realized by the execution of the processing of step 102 by the ECU 30. Further, the “distribution rate corresponding value calculating means” of the first invention is realized by the execution of the processing of steps 104 and 106 by the ECU 30. Further, the “correction value calculation means” of the first invention is realized by the execution of the processing of step 108 by the ECU 30.

Embodiment 2.
The second embodiment of the present invention will be described below with reference to FIG.

  The control device as the second embodiment of the present invention is applied to the engine system having the configuration shown in FIG. 1 as in the first embodiment, and the ECU 30 functions as the control device for the internal combustion engine 2. In the present embodiment, when executing the purge control, the ECU 30 executes torque fluctuation suppression control described below instead of the torque fluctuation suppression control described in the first embodiment.

  The torque fluctuation suppression control executed in the present embodiment obtains the concentration of the evaporated fuel introduced into each cylinder when the purge control is executed, and corrects the fuel injection amount of each cylinder according to the concentration of the evaporated fuel. It is intended to quickly suppress a torque change accompanying the introduction of the evaporated fuel, thereby suppressing a torque fluctuation of the internal combustion engine accompanying the introduction of the evaporated fuel. The torque fluctuation suppression control according to the first embodiment aims to eliminate the torque difference between the cylinders, whereas the torque fluctuation suppression control according to the present embodiment aims to eliminate the torque change for each cylinder. As in the first embodiment, correction of the fuel injection amount is performed by multiplying the fuel injection amount by a special correction coefficient for purge control separately from the correction coefficient used in the air-fuel ratio feedback.

  FIG. 7 is a flowchart for explaining a correction coefficient calculation method used in torque fluctuation suppression control according to the present embodiment. In the routine shown in FIG. 7, it is first determined whether or not purge control has been started (step 200). The start of the purge control can be determined by, for example, outputting a drive signal for duty-controlling the purge control valve 48. In the first embodiment, the correction coefficient for each cylinder is calculated in the same routine, whereas in this embodiment, the correction coefficient is calculated in a separate routine for each cylinder. That is, the routine of FIG. 7 is executed for each cylinder (FIG. 7 is a routine for calculating the correction coefficient of the m-th cylinder).

If it is determined in step 200 that the purge control is not started, a torque corresponding value corresponding to the generated torque is calculated for n cycles (for example, 10 cycles) of the m-th cylinder. Here, as the torque corresponding value, as described in the first embodiment, those indicated torque calculated in accordance with the equation of motion from the crank angle signal of the crank angle sensor 32 (estimated indicated torque) Ti _m is used And ECU30 averages them after the indicated torque Ti _m of n cycles is calculated to calculate the average indicated torque Ti _{M_ave} of the m cylinders (or, step 210). The calculated average _indicated torque _{Tim_ave} is stored in a memory (not shown ₎ of the ECU 30. The average _indicated torque _{Tim_ave} is calculated every time the determination in step 200 is not established, and the latest value is stored and updated in the memory of the ECU 30.

When it is determined that the purge control is started is determined in step 200, the time indicated torque of the m cylinders in the (k-th cycle) Ti _m (k) is calculated (step 202). Then, at the next step 204, the torque difference e of indicated torque Ti _m of the calculated present cycle (k) and the average indicated torque Ti _{M_ave} before the purge control is started in step 202 is calculated.

  The torque difference e corresponds to the concentration of the evaporated fuel introduced into the m-th cylinder with the start of purge control. The air-fuel ratio of the air-fuel mixture sucked into the combustion chamber 16 increases due to the introduction of the evaporated fuel, and accordingly, the generated torque of the cylinder also increases. Therefore, the concentration of the evaporated fuel can be estimated by obtaining the torque difference e before and after the introduction of the evaporated fuel. In step 206, the evaporated fuel concentration d corresponding to the torque difference e is calculated from a preset calculation formula or map. Although the concentration d of the evaporated fuel is calculated here, it is not always necessary to calculate the concentration itself, and a value corresponding to the concentration of the evaporated fuel may be calculated.

  The fuel injection amount correction coefficient is calculated from a map or the like (not shown) according to the evaporated fuel concentration d calculated in step 206 (step 208). In the map, the correction coefficient is set so that the fuel injection amount decreases as the concentration d of the evaporated fuel increases. The ECU 30 corrects the fuel injection amount of the mth cylinder in the next cycle using the correction coefficient calculated by the correction coefficient calculation routine. While the purge control is being executed, the ECU 30 repeats the processing of steps 200 to 208, calculates a new correction coefficient for each cycle, and continuously corrects the fuel injection amount for the next cycle.

  According to the torque fluctuation suppression control described above, the fuel injection amount is corrected for each cylinder according to the concentration of the evaporated fuel introduced by the execution of the purge control, so that the concentration immediately after the start of the purge control is unstable. Even so, the fuel injection amount can be accurately adjusted according to the concentration of the evaporated fuel. Further, since the concentration of the evaporated fuel is calculated from the torque difference between the average indicated torque before the start of purging and the indicated torque of the current cycle after the start of purging, the fuel injection amount can be adjusted immediately after the start of the purge control. From the above, according to the control device of the present embodiment, it is possible to quickly suppress the torque change of each cylinder that occurs when the evaporated fuel is introduced, thereby suppressing the torque fluctuation of the internal combustion engine 2. Become.

  In the embodiment described above, the “torque corresponding value calculation means” of the second invention is realized by the execution of the processing of steps 202 and 210 by the ECU 30. Further, the “concentration corresponding value calculation means” of the second invention is realized by the execution of the processing of steps 204 and 206 by the ECU 30. Further, the “correction value calculation means” of the second invention is realized by the execution of the processing of step 208 by the ECU 30.

Others.
While some of the embodiments of the present invention have been described above, the present invention is not limited to the above-described embodiments, and various modifications can be made without departing from the spirit of the present invention. For example, the following modifications may be made.

  In the first embodiment, the fuel injection amount of each cylinder is corrected so that the indicated torque of each cylinder matches the average indicated torque of all cylinders. However, the method is not limited to this as long as the torque difference between the cylinders can be eliminated. Not. For example, the fuel injection amount of each cylinder may be corrected so that the indicated torque of the other cylinder is matched with the cylinder having the largest indicated torque.

  Further, in each of the above embodiments, the fuel injection amount is corrected as a means for suppressing the torque difference between the cylinders or the change in torque of each cylinder accompanying the introduction of the evaporated fuel, but other engine control parameters are corrected. You may make it do. For example, the ignition timing of the spark plug 12 may be corrected, and when the intake valve 8 is provided with a variable valve mechanism, the intake angle is corrected by controlling the operating angle and lift amount of the intake valve 8. Good.

  Further, in each of the above embodiments, the indicated torque calculated from the crank angle signal supplied from the crank angle sensor 32 is used as the torque corresponding value, but any other value may be used as long as the value corresponds to the generated torque of the cylinder. May be used. For example, in the case where an in-cylinder pressure sensor for detecting the pressure in the combustion chamber 16 is provided, it can be calculated by a known method based on the operation data supplied from the in-cylinder pressure sensor and the operation data supplied from the crank angle sensor 32. The indicated torque may be used.

1 is a schematic configuration diagram of an engine system to which a control device as Embodiment 1 of the present invention is applied. It is a graph which shows the torque behavior of each cylinder during purge execution. It is a flowchart of the correction coefficient calculation routine performed in Embodiment 1 of this invention. It is a characteristic view showing the relationship between the indicated torque, the torque due to in-cylinder gas pressure, the inertia torque due to reciprocating inertia mass, and the crank angle. It is a schematic diagram which shows a crank angle signal and a torque calculation timing. It is a schematic diagram which shows the map showing the relationship between friction torque, engine speed, and cooling water temperature. It is a flowchart of the correction coefficient calculation routine performed in Embodiment 2 of the present invention.

Explanation of symbols

2 Internal combustion engine 4 Intake passage 6 Exhaust passage 8 Intake valve 10 Exhaust valve 12 Spark plug 14 Fuel injection valve 16 Combustion chamber 18 Throttle valve 22 Crankshaft 30 ECU (Electronic Control Unit)
32 Crank angle sensor 34 A / F sensor 40 Canister 48 Purge control valve 50 Fuel tank

Claims (2)

Evaporative fuel introduction means for introducing evaporative fuel generated in the fuel tank into the intake passage of the internal combustion engine;
Torque corresponding value calculating means for obtaining a torque corresponding value corresponding to the generated torque of each cylinder from the operation data of the internal combustion engine;
A distribution ratio corresponding value calculating means for obtaining a distribution ratio corresponding value corresponding to the distribution ratio of the evaporated fuel based on the torque corresponding value of each cylinder at the time of evaporative fuel introduction;
Control means for controlling the internal combustion engine in accordance with engine control parameters set for each cylinder;
Correction value calculation means for obtaining a correction value of the engine control parameter based on the distribution rate corresponding value;
A control device for an internal combustion engine, comprising: Evaporative fuel introduction means for introducing evaporative fuel generated in the fuel tank into the intake passage of the internal combustion engine;
Torque corresponding value calculating means for obtaining a torque corresponding value corresponding to the generated torque of each cylinder from the operation data of the internal combustion engine;
A concentration corresponding value corresponding to the concentration of the evaporated fuel introduced into the specific cylinder based on a difference between the torque corresponding value of the specific cylinder at the time of evaporating fuel introduction and the torque corresponding value of the specific cylinder before the evaporative fuel introduction. Concentration corresponding value calculating means for obtaining
Control means for controlling the internal combustion engine in accordance with engine control parameters set for each cylinder;
Correction value calculation means for obtaining a correction value of the engine control parameter for the specific cylinder based on the concentration correspondence value;
A control device for an internal combustion engine, comprising:

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