JP2007040219A - Control device of internal combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a control device of an internal combustion engine containing a period when expansion processes of a plurality of cylinders overlap one another which can effectively control a torque level difference. <P>SOLUTION: The torque of current cycle and the torque carried over to the next cycle are calculated by using a combustion model (a step 102). The generation of a torque level difference is predicted from the total torque determined by adding the torque carried over from the previous cycle to the torque of current cycle (a step 104). When the generation of torque level difference is predicted, optimal ignition timing, which generates no torque level difference, is calculated by using the combustion model. <P>COPYRIGHT: (C)2007,JPO&INPIT

Description

  The present invention relates to an internal combustion engine control apparatus, and more particularly to an internal combustion engine control apparatus suitable as an apparatus for controlling an internal combustion engine that includes a period in which expansion strokes of a plurality of cylinders overlap in a cycle.

  A spark ignition type internal combustion engine in which fuel is directly injected into a cylinder is known. Some internal combustion engines of this type operate by switching between a homogeneous combustion mode and a stratified combustion mode.

  When switching between homogeneous combustion and stratified combustion, a torque step is likely to occur. Japanese Patent Laid-Open No. 2004-60504 discloses a technique for controlling the ignition timing so as to absorb a torque step when switching between homogeneous combustion and stratified combustion.

JP 2004-60504 A

  However, according to the knowledge of the inventors of the present invention, even when the ignition timing control as disclosed in the above publication is performed, in the case of a six-cylinder engine or the like, the torque step is changed when switching between homogeneous combustion and stratified combustion. Was easy to grow.

  The present invention has been made to solve the above-described problems, and an object thereof is to provide a control device for an internal combustion engine that can effectively suppress a torque step.

In order to achieve the above object, a first invention is an apparatus for controlling an internal combustion engine that includes a period in which the expansion strokes of a plurality of cylinders overlap in a cycle,
For each cylinder, a cylinder-specific torque prediction means that predicts the relationship between the cylinder-specific torque and the crank angle generated by the combustion pressure of the cylinder based on the combustion model;
An overall torque predicting means for predicting a relationship between the overall torque of the internal combustion engine and a crank angle by adding up the cylinder specific torque according to the overlap of expansion strokes of the cylinders;
Torque step prediction means for predicting whether or not the torque step falls below an allowable level based on the relationship between the overall torque and the crank angle;
Parameter setting means for setting a predetermined engine control parameter so that the torque step is less than or equal to the allowable level when it is predicted that the torque step exceeds the allowable level;
It is characterized by providing.

The second invention is the first invention, wherein
The engine control parameter is an ignition timing.

  According to the first invention, for each cylinder, the relationship between the cylinder-specific torque generated by the combustion pressure of the cylinder and the crank angle can be predicted. Such cylinder specific torque reflects where the combustion pressure peak is relative to the crank angle. Then, by adding up the cylinder specific torque according to the overlap of the expansion strokes of the cylinders, it is possible to accurately predict the relationship between the overall torque of the internal combustion engine and the crank angle. The overall torque predicted in this way is faithfully influenced by the fact that the combustion pressure peaks of multiple cylinders, where the expansion strokes partially overlap, approach or leave each other due to a significant change in the ignition timing. Reflected. That is, this total torque faithfully reflects the torque concentration and the generation of torque valleys during the period in which the expansion strokes of a plurality of cylinders overlap. Therefore, according to the present invention, by predicting the degree of the torque step based on the relationship between the overall torque and the crank angle, the ignition timing is greatly increased, for example, when switching between stratified combustion and homogeneous lean combustion. It is possible to accurately predict the degree of torque step when changed. According to the present invention, when it is predicted that the torque step exceeds the allowable level, the predetermined engine control parameter is set so that the torque step is equal to or less than the allowable level, so that the excessive torque step is reduced. Occurrence can be avoided in advance.

  According to the second invention, when the torque step is predicted to exceed the allowable level, it is possible to reliably avoid the occurrence of the torque step exceeding the allowable level by adjusting the ignition timing.

Embodiment 1 FIG.
[Description of system configuration]
FIG. 1 is a diagram for explaining a system configuration according to the first embodiment of the present invention. As shown in FIG. 1, the system of the present embodiment includes an internal combustion engine 10. The internal combustion engine 10 is a six-cylinder engine, and FIG. 1 shows a cross section of one of the cylinders. Each cylinder of the internal combustion engine 10 is provided with a piston 11, an intake valve 12, an exhaust valve 14, an ignition plug 16, and an intake port 18 and an exhaust port 20 that communicate with the inside of the cylinder. The intake valve 12 opens and closes so that the combustion chamber 22 and the intake port 18 are in a conduction state or a cutoff state. The exhaust valve 14 opens and closes so that the combustion chamber 22 and the exhaust port 20 are in a conductive state or a shut-off state.

  Each cylinder of the internal combustion engine 10 is provided with an in-cylinder injector 24 that directly injects fuel into the cylinder (inside the combustion chamber 22). The in-cylinder injector 24 is fed with fuel pressurized by a pump (not shown).

  The intake port 18 communicates with the intake passage 30. An air cleaner 32 is provided at the upstream end of the intake passage 30. Air is taken into the intake passage 30 via the air cleaner 32. An air flow meter 33 is disposed downstream of the air cleaner 32. The air flow meter 33 is a sensor that detects an intake air amount GA flowing in the intake passage 30. A downstream portion of the intake passage 30 is branched for each cylinder (for each intake port 18), and a surge tank 34 is provided at the branched portion.

  A throttle valve 36 is disposed upstream of the surge tank 34 in the intake passage 30. The throttle valve 36 is an electronically controlled throttle valve that opens and closes by driving a motor. The throttle valve 36 is provided with a throttle position sensor 37 for detecting the opening degree. An intake pressure sensor 38 that detects the intake pipe pressure PM is provided downstream of the throttle valve 36.

  Connected to the exhaust port 20 is an exhaust passage 40 for discharging the combustion gas generated by the combustion in the combustion chamber 22 as an exhaust gas. A catalyst 42 for purifying exhaust gas is provided in the exhaust passage 40. An air-fuel ratio sensor 44 for detecting the air-fuel ratio of the exhaust gas is disposed upstream of the catalyst 42 in the exhaust passage 40.

  A crank angle sensor 46 is installed in the vicinity of the crankshaft 45 of the internal combustion engine 10. The crank angle sensor 46 is a sensor that reverses the Hi output and the Lo output each time the crankshaft 45 rotates by a predetermined rotation angle. According to the output of the crank angle sensor 46, it is possible to detect the rotational position and rotational speed of the crankshaft 45, and further the engine rotational speed NE and the like.

  The system of this embodiment includes an ECU (Electronic Control Unit) 50. Connected to the ECU 50 are various sensors such as the air flow meter 33, throttle position sensor 37, intake pressure sensor 38 and air-fuel ratio sensor 44 described above, and various actuators such as the ignition plug 16, in-cylinder injector 24 and throttle valve 36 described above. Has been. The ECU 50 can control the operating state of the internal combustion engine 10 by appropriately driving each actuator based on the output of each sensor.

[Outline of Operation of Embodiment 1]
(Switching combustion mode)
The ECU 50 switches between operation near the theoretical air-fuel ratio and operation at a leaner air-fuel ratio depending on the operation state detected by each sensor. Further, the ECU 50 switches between homogeneous lean combustion and stratified combustion according to the operating state when performing lean air-fuel ratio operation.

  When performing homogeneous lean combustion, fuel is injected from the in-cylinder injector 24 in the intake stroke. The injected fuel is sufficiently mixed with the air in the cylinder during the compression stroke, so that a homogeneous lean air-fuel mixture is formed. In homogeneous lean combustion, such a homogeneous lean mixture is burned.

  On the other hand, when stratified combustion is performed, fuel is injected from the in-cylinder injector 24 in the compression stroke. The injected fuel bounces up in the cavity of the piston 11 and forms a rich mixture layer near the stoichiometric air-fuel ratio in the vicinity of the spark plug 16. In the stratified combustion, the air-fuel ratio is extremely lean as a whole by burning in a state where this relatively rich mixture layer is formed only around the spark plug 16.

  The ECU 50 switches between the homogeneous lean combustion and the stratified combustion by controlling the fuel injection timing from the in-cylinder injector 24, the opening degree of the throttle valve 36, and the like.

  Also, the ignition timing differs between homogeneous lean combustion and stratified combustion. In the homogeneous lean combustion, the ignition timing is set early in order to reliably burn a homogeneous lean air-fuel mixture having a property of being difficult to burn. On the other hand, although stratified combustion is lean as a whole, since the fuel concentration is relatively high only around the spark plug 16, it is not difficult to burn. Therefore, it is not necessary to advance the ignition timing so much. On the other hand, in stratified combustion, it is necessary to perform ignition in accordance with the timing when the fuel soars around the spark plug 16. For this reason, the ignition timing suitable for homogeneous lean combustion is often at a position that is greatly advanced from the ignition timing suitable for stratified combustion. For this reason, when switching between homogeneous lean combustion and stratified combustion, the ignition timing is usually accompanied by a significant change.

(Torque step)
Since homogeneous lean combustion and stratified combustion differ greatly in the manner of combustion as described above, a torque step is likely to occur at the time of switching. Conventionally, particularly in a six-cylinder engine or the like, a torque step at the time of switching between homogeneous lean combustion and stratified combustion tends to be large. The present inventors have found that the cause is as follows.

  FIG. 2 is a diagram showing the relationship between the combustion pressure (in-cylinder pressure) and the crank angle for each of three cylinders that explode in the order of a six-cylinder engine. An engine having a plurality of cylinders is usually configured such that the cylinders explode at equal intervals. For this reason, in the case of a six-cylinder engine, the crank angles of the six cylinders are shifted by 120 ° CA.

  Referring to the example of cylinder # 1 and cylinder # 2 in FIG. 2, when cylinder # 2 starts the expansion stroke, that is, when it is at the compression top dead center, cylinder # 1 that has previously exploded is compressed top dead. It is in a state of 120 ° CA after the point. That is, the 60 ° CA period at the end of the expansion stroke of the # 1 cylinder overlaps the 60 ° CA period at the beginning of the expansion stroke of the # 2 cylinder. Therefore, the total torque of the internal combustion engine 10 during the 60 ° CA period at the beginning of the expansion stroke of the # 2 cylinder is the sum of the torque generated by the # 2 cylinder combustion pressure and the torque generated by the # 1 cylinder combustion pressure. It becomes a thing.

  As described above, the torque generated by the combustion pressure of 120 ° CA to 180 ° CA after the compression top dead center is integrated with the torque generated by the combustion pressure of the cycle performed by the next explosion cylinder. In the present embodiment, the torque generated by the combustion pressure of 120 ° CA to 180 ° CA after compression top dead center is referred to as torque carryover to the next cycle. Taking the # 1 cylinder in FIG. 2 as an example, the hatched portion corresponds to the combustion pressure causing the torque carryover.

  Now, assume that stratified combustion is switched to homogeneous lean combustion between the # 1 cylinder and the # 2 cylinder in FIG. In this case, the ignition timing is greatly advanced between the # 1 cylinder and the # 2 cylinder in accordance with the switching from the stratified combustion to the homogeneous lean combustion. In the # 1 cylinder where the ignition timing was delayed, the peak of the combustion pressure is at the rear position. On the other hand, in the # 2 cylinder where the ignition timing is greatly advanced, the peak of the combustion pressure moves forward. As a result, the torque is concentrated when the peaks of the combustion pressures of both cylinders approach. In this way, excessive torque is instantaneously generated at the position where the expansion stroke of the # 1 cylinder and the expansion stroke of the # 2 cylinder overlap, and a large torque step is likely to occur.

  Conversely, when switching from homogeneous lean combustion to stratified combustion, the ignition timing is significantly retarded. The significant advance of the ignition timing keeps the combustion pressure peaks of the two cylinders across the switchover. As a result, a torque valley is generated, and a large torque step is likely to occur.

  As described above, when the ignition timing is significantly changed, a large torque step is likely to occur at the portion where the expansion strokes of the two cylinders overlap due to the effect of the torque carryover. Therefore, in this embodiment, the torque at the portion where the expansion strokes of the two cylinders overlap is predicted in advance after taking into account the torque carryover. Based on the predicted torque, the ignition timing is adjusted so that the torque step does not exceed the allowable level.

  The above-described circumstances relating to the torque step are common to an engine having a portion where the expansion strokes of a plurality of cylinders overlap, that is, an engine having five or more cylinders. In this embodiment, a six-cylinder engine will be described as an example, but the present invention can be applied not only to a six-cylinder engine but also to an engine having an arbitrary number of cylinders of five or more cylinders.

  Moreover, the situation regarding the torque step described above is not limited to the time of switching between the stratified combustion mode and the homogeneous lean combustion mode as long as the ignition timing is significantly changed. For example, when switching between the stoichiometric combustion mode operated at the stoichiometric air-fuel ratio and the homogeneous lean combustion mode, the ignition timing is generally significantly changed. The present invention can be effectively applied to any case involving a significant change in ignition timing, including such a case.

(Combustion model)
In the present embodiment, torque is predicted by a method of calculating torque (illustrated torque) for each cylinder using a combustion model. In this embodiment, a Wiebe function model described below is used as the combustion model. The Wiebe function model is a heat release rate model. The Wiebe function is expressed by the following equation.

The meaning of each symbol in the above formula (1) is as follows.
Q: Calorific value θ: Crank angle [° CA]
m: Shape factor
k: combustion efficiency θ _b : ignition delay [° CA]
θ _p : Combustion period [° CA]
a = 6.9 (constant)
Among these, the shape factor m, the combustion efficiency k, the ignition delay θ _b , and the combustion period θ _p are parameters that are determined according to operating conditions.

  This Wiebe function can be solved for the calorific value Q as follows. First, the following relationship holds mathematically.

  G (θ) in the above equation (2) is a function defined in parentheses on the middle side of the above equation (2). Using the relationship of the above equation (2), the above equation (1) can be modified as follows.

  By integrating both sides of the above equation (3) with θ, the following equation is obtained.

The above equation (4) is a solution for the calorific value Q of the Wiebe function. The integral constant C in the above equation (4) is determined by the initial value of Q. In this case, the initial value of Q is, when theta = theta _b, a Q = 0.

  Next, a method for calculating torque from the above equation (4) will be described. First, the following energy equation is established thermodynamically.

  In the above equation (5), P represents the combustion pressure (in-cylinder pressure), V represents the cylinder volume, κ represents the specific heat ratio, and dV / dθ represents the change in cylinder volume for each crank angle. V and dV / dθ are geometrically determined according to the crank angle θ. That is, V and dV / dθ are functions of the crank angle θ.

  By substituting Q expressed by the above equation (4) into the above equation (5) and solving for the combustion pressure P, the combustion pressure P for each crank angle can be calculated. In this calculation process, as an initial condition, it is assumed that P = intake pipe pressure is established when θ is θ when the intake valve is closed.

The relationship of the following equation holds between the torque T and the combustion pressure P.
T = P · dV / dθ (6)
By substituting the combustion pressure P for each crank angle into the above equation (6), the torque T for each crank angle can be calculated.

(Torque step prediction)
In the present embodiment, the magnitude of the torque step is predicted as follows based on the cylinder-specific torque calculated as described above.

FIG. 3 is a diagram for explaining a method of predicting the magnitude of the torque step. Here, the average value T ₁ of the total torque of the internal combustion engine 10 during the period of 0 ° CA~120 ° CA after the compression top dead center of ♯1 cylinders, the period leading to it, that is after the compression top dead center of ♯2 cylinder 0 A difference (T ₂ −T ₁ ) from the average value T ₂ of the overall torque of the internal combustion engine 10 in the period of ° to CA 120 ° CA is set as a predicted value of the torque step.

  As shown in the lower graph of FIG. 3, the overall torque of the internal combustion engine 10 in the period from 0 ° CA to 60 ° CA after the compression top dead center of the # 2 cylinder is the cylinder-specific torque (no hatching) of the # 2 cylinder. Portion) and the cylinder specific torque (portion with hatching) of 120 ° CA to 180 ° CA after compression top dead center of the # 1 cylinder. On the other hand, the overall torque of the internal combustion engine 10 in the period from 60 ° CA to 120 ° CA after the compression top dead center of the # 2 cylinder is solely due to the cylinder specific torque of the # 2 cylinder.

Therefore, the average value T ₂ of the overall torque of the internal combustion engine 10 during the period of 0 ° ~CA120 ° CA after the compression top dead center of ♯2 cylinder is in the range of 0 ° CA~120 ° CA in the lower graph in FIG. 3 Can be obtained by averaging including the torque carry-over portion with hatching.

Similarly, the average value T ₁ of the total torque of the internal combustion engine 10 during the period of 0 ° ~CA120 ° CA after the compression top dead center of ♯1 cylinder, 0 ° CA~120 ° CA of the upper graph in FIG. 3 Can be obtained by averaging the range including the torque carry-over portion with hatching.

In the present embodiment, the ignition timing is adjusted so that the absolute value of the torque step (T ₂ −T ₁ ) predicted in this way is below an allowable level.

[Specific Processing in Embodiment 1]
FIG. 4 is a flowchart of a routine executed by the ECU 50 in the present embodiment in order to realize the above function. In the system of the present embodiment, the target ignition timing is set according to the operating conditions by processing of another routine (not shown). The routine shown in FIG. 4 performs processing for resetting the target ignition timing so that a torque step does not occur. This routine is repeatedly executed according to the explosion order of each cylinder of the internal combustion engine 10. In the following description, the cycle of the cylinder targeted by the present routine being executed is referred to as “current cycle”.

  According to the routine shown in FIG. 4, first, the operating conditions and the torque carryover of the previous cycle are read (step 100). Specifically, the target air-fuel ratio and target ignition timing of the current cycle, the current engine speed NE, and the load factor KL are acquired as the operating conditions. The engine speed NE can be acquired based on the output of the crank angle sensor 46. Further, the load factor KL can be acquired based on the output of the intake pressure sensor 38. The torque carryover of the previous cycle will be described later.

Next, the torque of the current cycle and the torque carry-over to the next cycle are calculated by the following processing (step 102). First, the shape factor m, the combustion efficiency k, the ignition delay θ _b , and the combustion period θ _{p that} are parameters of the Wiebe function are calculated based on the following equations.
m = f ₁ (A / F, NE, KL, SA) (7)
k = f ₂ (A / F, NE, KL, SA) (8)
θ _b = f ₃ (A / F, NE, KL, SA) (9)
θ _p = f ₄ (A / F, NE, KL, SA) (10)

In the above equations (7) to (10), the shape factor m, the combustion efficiency k, the ignition delay θ _b , and the combustion period θ _p are respectively expressed as the air-fuel ratio A / F, the engine speed NE, the load factor KL, and the ignition. This function specifies the time SA as a variable. That is, the (7) to (10) is representative of A / F, NE, KL, and changes in SA is, m, k, theta _b, and the characteristics of the changes that result in each of theta _p. Such characteristics have been investigated in advance based on experiments and theories. Note that such characteristics are not limited to the function form as described above, and may be stored in the ECU 50 in the form of a map, for example.

The ECU 50 adds the target air-fuel ratio, engine speed NE, load factor KL, and target ignition timing acquired in step 100 to the variables A / F, NE, KL, and SA in the expressions (7) to (10). Respectively, m, k, θ _b , and θ _p are calculated.

In step 102, the calculated m, k, θ _b , and θ _p are then substituted into the above equation (4). Then, (when _{θ = θ b, Q = 0} ) the initial conditions by substituting, for calculating the integral constant C in equation (4). With the above processing, the above equation (4) is specified as representing the calorific value Q for each crank angle predicted for the current cycle.

  In the step 102, next, the specified Q expressed by the above equation (4) is substituted into the above equation (5), and this is solved for the combustion pressure P by numerical calculation. Thereby, the combustion pressure P for each crank angle is calculated. In this calculation, the specific heat ratio κ is set to a predetermined constant value. In this calculation, the initial condition is that P is the value of the intake pressure detected by the intake pressure sensor 38 when θ is the value when the intake valve is closed.

  Subsequently, the cylinder-specific torque T for each crank angle is calculated by substituting the calculated combustion pressure P for each crank angle into the above equation (6). In step 102 described above, the torque in the range of 0 ° CA to 120 ° CA after the compression top dead center is stored as the torque of the current cycle among the cylinder specific torques T for each crank angle thus calculated. The torque in the range of 120 ° CA to 180 ° CA after compression top dead center is stored as the torque carry-over to the next cycle. The torque carry-over amount of the previous cycle read in step 100 is the torque carry-over amount calculated in step 120 when the previous routine is executed.

Next, it is predicted by the method described with reference to FIG. 3 whether or not the torque step falls below an allowable level (step 104). Specifically, first, the value of the total torque of the internal combustion engine 10 is obtained by adding the torque carryover from the previous cycle read in step 100 to the torque of the current cycle calculated in step 102. Then, an average value T ₂ over 0 ° CA~120 ° CA of the entire torque. Next, the difference (T ₂ −T ₁ ) from the total torque average value T ₁ of the previous cycle, which was calculated in step 104 when this routine was executed last time, is calculated as a torque step. Then, it is determined whether or not the absolute value of the torque difference (T ₂ -T ₁₎ is equal to or less than the allowable value alpha.

In the above step 104, when it is recognized that | T ₂ −T ₁ | ≦ α is established, it can be predicted that the magnitude of the torque step falls within the allowable level. In this case, the target ignition timing read in step 100 is used as it is as the actual ignition timing of the current cycle (step 106).

On the other hand, if it is found in step 104 that | T ₂ −T ₁ | ≦ α is not established, it can be predicted that the magnitude of the torque step exceeds the allowable level. In this case, an optimal ignition timing is calculated so that the magnitude of the torque step can be kept below the allowable level (step 108). Specifically, by using the ignition timing value that is advanced or retarded little by little (for example, by 1 ° CA) from the target ignition timing, the same processing as in steps 102 and 104 described above is performed, so that | T ₂ − Find an ignition timing that satisfies T ₁ | ≦ α. The found ignition timing is used as the actual ignition timing of the current cycle. By this process, it is possible to avoid a torque step exceeding an allowable level.

  By the way, in the first embodiment described above, the torque step is suppressed by resetting the ignition timing in the above step 108, but in the present invention, engine control parameters (for example, air-fuel ratio) other than the ignition timing are set. You may make it suppress a torque level | step difference by resetting.

  In the first embodiment described above, the Wiebe function model corresponds to the “combustion model” in the first invention, and the ignition timing corresponds to the “predetermined engine control parameter” in the first invention. . Further, when the ECU 50 executes the process of step 102, the “cylinder-specific torque predicting means” in the first invention executes the process of step 104, so that the “total torque prediction in the first invention”. The “parameter setting means” according to the first aspect of the present invention is realized by the means ”and the“ torque step prediction means ”executing the processing of step 108 described above.

It is a figure for demonstrating the system configuration | structure of Embodiment 1 of this invention. It is the figure which showed the relationship between the combustion pressure (cylinder pressure) according to each cylinder, and the crank angle about three cylinders which explode in order among 6 cylinder engines. It is a figure for demonstrating the method of calculating the magnitude | size of a torque level | step difference. It is a flowchart of the routine performed in Embodiment 1 of the present invention.

Explanation of symbols

DESCRIPTION OF SYMBOLS 10 Internal combustion engine 12 Intake valve 14 Exhaust valve 16 Spark plug 18 Intake port 20 Exhaust port 22 Combustion chamber 24 In-cylinder injector 30 Intake passage 33 Air flow meter 36 Throttle valve 38 Intake pressure sensor 40 Exhaust passage 42 Catalyst 44 Air fuel ratio sensor 46 Crank angle Sensor 50 ECU

Claims (2)

An apparatus for controlling an internal combustion engine including a period in which expansion strokes of a plurality of cylinders overlap in a cycle,
For each cylinder, a cylinder-specific torque prediction means that predicts the relationship between the cylinder-specific torque and the crank angle generated by the combustion pressure of the cylinder based on the combustion model;
An overall torque predicting means for predicting a relationship between the overall torque of the internal combustion engine and a crank angle by adding up the cylinder specific torque according to the overlap of expansion strokes of the cylinders;
Torque step prediction means for predicting whether or not the torque step falls below an allowable level based on the relationship between the overall torque and the crank angle;
Parameter setting means for setting a predetermined engine control parameter so that the torque step is less than or equal to the allowable level when it is predicted that the torque step exceeds the allowable level;
A control device for an internal combustion engine, comprising:   2. The control device for an internal combustion engine according to claim 1, wherein the engine control parameter is an ignition timing.

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