CN116964312A

CN116964312A - Control device for internal combustion engine

Info

Publication number: CN116964312A
Application number: CN202280019882.9A
Authority: CN
Inventors: 助川义宽; 赤城好彦; 押领司一浩; 熊野贤吾
Original assignee: Hitachi Astemo Ltd
Current assignee: Hitachi Astemo Ltd
Priority date: 2021-04-12
Filing date: 2022-03-02
Publication date: 2023-10-27
Also published as: DE112022000672T5; WO2022219952A1; JP2022162284A

Abstract

The internal combustion engine control device of the present invention includes a combustion state estimating section that estimates a combustion state of the engine based on a first in-cylinder pressure acquired near an ignition timing and a second in-cylinder pressure acquired near a combustion end timing. The combustion state estimating unit calculates a difference between the first in-cylinder pressure and the second in-cylinder pressure in a plurality of cycles, and estimates a magnitude of torque fluctuation of the internal combustion engine based on a fluctuation rate of the difference in the plurality of cycles.

Description

Control device for internal combustion engine

Technical Field

The present invention relates to an internal combustion engine control device.

Background

In recent years, in vehicles such as automobiles, restrictions on fuel consumption (fuel consumption) and harmful components in exhaust gas are strengthened, and there is a trend that such restrictions are also strengthened in the future. In this case, a technique of estimating a state in a combustion chamber of the engine and controlling the engine based on the estimation result is known. The air-fuel ratio, ignition timing, and the like are appropriately controlled according to the current combustion state, so that the thermal efficiency of the engine can be improved or the emission of harmful gas can be reduced.

The method of determining the heat generation distribution and the combustion torque from the detection result of the in-cylinder pressure by the pressure sensor is widely used for estimation of the combustion state of the engine. Such a technique for estimating the combustion state is disclosed in, for example, patent document 1 and patent document 2.

Patent document 1 describes the following technique: the engine torque of the internal combustion engine is calculated using the graphic torque and the pump loss torque.

Patent document 2 describes the following technique: the in-cylinder pressure is sampled at a crank angle position of 60 ° before the top dead center and at a crank angle position of 60 ° after the top dead center, and a detected pressure ratio is employed as a combustion state parameter.

Prior art literature

Patent literature

Patent document 1: japanese patent laid-open publication No. 2017-57803

Patent document 2: japanese patent laid-open No. 2002-97996

Disclosure of Invention

Technical problem to be solved by the invention

The method of estimating the combustion state from the measurement result (in-cylinder pressure) of the in-cylinder pressure sensor described in patent document 1 requires a time-series pressure value in a predetermined section of the engine cycle, for example, a section in which the crank angle around the compression top dead center is 360 °. Further, in order to estimate the combustion state with sufficient accuracy, it is necessary to sample the pressure at intervals at which the crank angle is substantially 1 ° or less. Therefore, the control device for estimating the combustion state requires a large storage capacity and high-speed calculation capability. In addition, a pressure sensor corresponding to high-speed sampling is also required. This causes a problem that the cost of the engine system increases.

Further, although patent document 2 describes detecting in-cylinder pressure at a crank angle position of 60 ° before the top dead center and at a crank angle position of 60 ° after the top dead center, it is known that the combustion state of the internal combustion engine (engine) cannot be accurately estimated from the in-cylinder pressure detected at the positions.

The present invention has been made in view of the above circumstances, and an object thereof is to provide an internal combustion engine control device capable of estimating a combustion state of an internal combustion engine at low cost.

Technical means for solving the technical problems

The internal combustion engine control device according to the present invention includes a combustion state estimating unit that estimates a combustion state of the engine based on a first in-cylinder pressure acquired near an ignition timing and a second in-cylinder pressure acquired near a combustion end timing.

Effects of the invention

According to the present invention, it is possible to provide an internal combustion engine control device capable of estimating the combustion state of an internal combustion engine at low cost.

Other problems, configurations and effects than those described above will become more apparent from the following description of the embodiments.

Drawings

Fig. 1 is a cross-sectional view showing an example of the overall structure of an engine according to embodiment 1 of the present invention.

Fig. 2 is a block diagram showing a configuration example of a controller according to embodiment 1 of the present invention.

Fig. 3 is a flowchart showing an example of the estimation process of the magnitude of the engine torque fluctuation, which is performed by the torque fluctuation estimating unit of the controller according to embodiment 1 of the present invention.

Fig. 4 is a characteristic diagram showing a correlation between the cyclic fluctuation ratio of the differential pressure and the magnitude of the engine torque fluctuation according to embodiment 1 of the present invention.

Fig. 5 is a graph showing an example of the detection timing of the in-cylinder pressure according to embodiment 1 of the present invention.

Fig. 6 is a characteristic diagram showing an example of a presumption error of torque fluctuation with respect to a difference between the acquisition timing of the first in-cylinder pressure and the ignition timing according to embodiment 1 of the present invention.

Fig. 7 is a diagram showing a preferred variation of the acquisition timing of the first in-cylinder pressure with respect to the variation of the ignition timing according to embodiment 1 of the present invention.

Fig. 8 is a characteristic diagram showing an example of an estimation error of the engine torque fluctuation according to embodiment 1 of the present invention with respect to the difference between the acquisition timing of the second in-cylinder pressure and ca90+20°.

Fig. 9 is a characteristic diagram showing a relationship between the combustion timing and the integrated value of the heat generation rate according to embodiment 1 of the present invention.

Fig. 10 is a diagram showing a preferred variation of the second in-cylinder pressure acquisition timing according to embodiment 1 of the present invention with respect to a change of ca90+20°.

Fig. 11 is a diagram showing an example of a control module of a controller for performing EGR control according to embodiment 1 of the present invention.

Fig. 12 is a diagram showing an example of control of an actuator based on the deviation δcov in the EGR system according to embodiment 1 of the present invention.

Fig. 13 is a diagram showing an example in which the controller according to embodiment 1 of the present invention controls the amount of ignition energy, the gas flow intensity, the compression ratio, and the intake air temperature based on the deviation δcov.

Fig. 14 is a diagram showing an example of control of an actuator based on the deviation δcov in the lean burn system according to embodiment 1 of the present invention.

Fig. 15 is a block diagram showing a configuration example of a controller according to embodiment 2 of the present invention.

Fig. 16 is a characteristic diagram showing the correlation between the reference combustion center of gravity and the ratio of the first in-cylinder pressure to the second in-cylinder pressure according to embodiment 2 of the present invention.

Fig. 17 is a diagram showing an example of the correction amount from the reference combustion center of gravity against the change in the volumetric efficiency according to embodiment 2 of the present invention.

Fig. 18 is a diagram showing an example of the correction amount from the reference combustion center of gravity against the change in the engine speed according to embodiment 2 of the present invention.

Fig. 19 is a flowchart showing an example of a combustion center of gravity estimation procedure performed by the combustion center of gravity estimation unit of the controller according to embodiment 2 of the present invention.

Fig. 20 is a diagram showing a general relationship between the optimum combustion center of gravity and the fuel consumption rate of the engine according to embodiment 2 of the present invention.

Fig. 21 is a diagram showing an example of a control module of a controller for performing EGR control according to embodiment 2 of the present invention.

Fig. 22 is a characteristic diagram showing a correlation between a difference between a first in-cylinder pressure and a second in-cylinder pressure and a reference combustion center of gravity according to embodiment 2 of the present invention.

Fig. 23 is a diagram showing an example of a time history of an ignition signal, primary and secondary voltages of an ignition coil, and a secondary current according to embodiment 2 of the present invention.

Fig. 24 is a characteristic diagram showing the correlation between the maximum value of the primary voltage immediately after the start of discharge, the maximum value of the secondary voltage, and the in-cylinder pressure at the time of discharge according to embodiment 2 of the present invention.

Fig. 25 is a block diagram showing a configuration example from an ignition coil to a controller in the case where the controller according to embodiment 2 of the present invention obtains an in-cylinder pressure based on a secondary voltage of the ignition coil.

Fig. 26 is a characteristic diagram showing a correlation between the discharge timing and the in-cylinder pressure at the time of discharge according to embodiment 2 of the present invention.

Fig. 27 is a block diagram showing a configuration example from a crank angle sensor to a controller in the case where the controller according to embodiment 2 of the present invention obtains an in-cylinder pressure based on an angular acceleration of a crank shaft.

Detailed Description

The mode for carrying out the present invention will be described below with reference to the accompanying drawings. In the present specification and the drawings, constituent elements having substantially the same function or structure are denoted by the same reference numerals, and repetitive description thereof will be omitted.

< embodiment 1>

[ structural example of Engine ]

First, an example of an engine to which the present invention is applied is described with reference to fig. 1.

Fig. 1 is a cross-sectional view showing an example of the overall structure of an engine 1 according to embodiment 1 of the present invention.

The engine 1 is an example of a spark-ignition four-cycle gasoline engine. The combustion chamber of the engine 1 is formed by an engine head, a cylinder 13, a piston 14, an intake valve 15, and an exhaust valve 16. In the engine 1, the fuel injection valve 18 is provided on the intake port 21, and the injection nozzle of the fuel injection valve 18 penetrates into the intake port 21, thereby constituting a so-called port injection type internal combustion engine.

A spark plug 17b is provided on the engine cylinder head, and an ignition coil 17a is provided at an upper portion of the spark plug 17 b. The pressure sensor 10 is disposed in parallel to the engine head. The pressure sensor 10 detects the in-cylinder pressure in the cylinder 13 by capturing deformation of the metal diaphragm caused by a pressure difference, for example, by a piezoelectric resistance element.

Air for combustion is drawn into the combustion chamber through an air filter 19, a throttle valve 20, and an intake port 21. Then, the burned gas (exhaust gas) discharged from the combustion chamber is discharged to the atmosphere through the exhaust port 24 and the catalytic converter 25.

The amount of air taken into the combustion chamber is measured by an air flow sensor 22 provided on the upstream side of the throttle valve 20. Further, the air-fuel ratio of the gas (exhaust gas) discharged from the combustion chamber is detected by an air-fuel ratio sensor 27 provided on the upstream side of the catalytic converter 25.

The exhaust port 24 and the intake port 21 communicate through an EGR pipe 28, and constitute a so-called exhaust gas recirculation system (EGR system) in which a part of the exhaust gas flowing through the exhaust port 24 returns to the inside of the intake port 21. The amount of exhaust gas flowing through the EGR pipe 28 is regulated by the EGR valve 29.

Further, a timing rotor 26 (signal rotor) is provided at the shaft portion of the crank shaft. The crank angle sensor 11 (detecting portion) disposed in the vicinity of the timing rotor 26 (detected portion) detects the rotation of the timing rotor 26 to detect the rotation and the phase of the crank shaft, that is, the crank speed (engine speed). Detection signals of the pressure sensor 10, the crank angle sensor 11, the air flow sensor 22, and the air-fuel ratio sensor 27 are acquired into the controller 12.

The controller 12 is an example of a control device of the engine 1, and uses an ECU (Engine Control Unit: engine control unit), for example. The controller 12 outputs instructions of the opening degree of the throttle valve 20, the opening degree of the EGR valve 29, the fuel injection timing or the fuel injection amount of the fuel injection valve 18, the ignition timing of the ignition plug 17b, and the like based on the detection values of the various sensors, thereby controlling the engine 1 in a prescribed operating state.

In fig. 1, only a single cylinder is shown for illustrating the structure of the combustion chamber of the engine 1, but the engine according to the embodiment of the present invention may be a multi-cylinder engine including a plurality of cylinders.

[ structural example of controller ]

Fig. 2 is a block diagram showing a configuration example of the controller 12 according to embodiment 1 of the present invention.

The controller 12 includes an input/output unit 121, a control unit 122, and a storage unit 123 electrically connected to each other via a system bus not shown.

The input/output unit 121 includes an input port and an output port, which are not shown, and performs processing for inputting and outputting various signals to and from various devices and sensors in the vehicle on which the engine 1 is mounted. For example, the input/output unit 121 reads a signal of the pressure sensor 10, and transmits the signal to the control unit 122. The input/output unit 121 outputs control signals to the respective devices according to the command of the control unit 122.

The control unit 122 controls the operation of the engine 1. For example, the control unit 122 controls the throttle opening, the EGR opening, the fuel injection amount, and the ignition timing according to the combustion steady state of the engine 1. The control unit 122 according to embodiment 1 includes a torque fluctuation estimating unit 122a and an engine control unit 122b. The torque fluctuation estimating section 122a serves as an example of a combustion state estimating section.

The torque fluctuation estimating unit 122a estimates the magnitude of the engine torque fluctuation based on the in-cylinder pressure detected by the pressure sensor 10. The combustion state of the engine 1 estimated in the present embodiment is the magnitude of the engine torque fluctuation. The combustion state estimating unit (torque fluctuation estimating unit 122 a) estimates the combustion state of the internal combustion engine (engine 1) based on the first in-cylinder pressure P1 acquired near the ignition timing and the second in-cylinder pressure P2 acquired near the combustion end timing. Therefore, the combustion state estimating unit (torque fluctuation estimating unit 122 a) calculates the difference between the first in-cylinder pressure P1 and the second in-cylinder pressure P2 in a plurality of cycles, and estimates the magnitude of the engine torque fluctuation of the internal combustion engine (engine 1) based on the fluctuation rate of the difference in the plurality of cycles.

The engine control unit (engine control unit 122 b) controls the EGR rate or the air-fuel ratio based on the magnitude of the engine torque fluctuation. For example, the engine control unit 122b can control the EGR rate or the air-fuel ratio by changing the EGR opening degree, the throttle opening degree, and the fuel injection amount of the engine 1 based on the magnitude of the engine torque fluctuation obtained by the torque fluctuation estimating unit 122 a. The engine control unit (engine control unit 122 b) can also control at least one of ignition energy, in-cylinder flow strength, compression ratio, and intake air temperature based on the magnitude of the engine torque variation.

The storage unit 123 is a volatile Memory such as a RAM (Random Access Memory: random access Memory) or a nonvolatile Memory such as a ROM (Read Only Memory). The storage unit 123 stores a control program executed by an arithmetic processing device (not shown) provided in the controller 12. The arithmetic processing device reads and executes the control program from the storage unit 123, thereby realizing the functions of the respective modules of the control unit 122. For example, a CPU (Central Processing Unit: central processing unit) or an MPU (Micro Processing Unit: micro processing unit) is used as the arithmetic processing device. The controller 12 has a nonvolatile auxiliary storage device constituted by a semiconductor memory or the like, and the control program may be stored in the auxiliary storage device.

[ estimation Process of the Engine Torque Change size ]

Next, an example of the estimation process of the magnitude of the engine torque variation performed by the controller 12 will be described with reference to fig. 3.

Fig. 3 is a flowchart showing an example of the estimation process of the magnitude of the engine torque fluctuation, which is performed by the torque fluctuation estimating unit 122a of the controller 12.

First, the torque fluctuation estimating unit 122a initializes the variable S and the variable Q to zero (S1). Then, the torque fluctuation estimating section 122a acquires a first in-cylinder pressure P1 (referred to as "in-cylinder pressure P1" in the figure) in the vicinity of the ignition timing detected by the pressure sensor 10 (S2). Then, the torque fluctuation estimating unit 122a obtains a second in-cylinder pressure P2 (referred to as "in-cylinder pressure P2" in the figure) near the combustion end timing detected by the pressure sensor 10 (S3).

Next, the torque fluctuation estimating unit 122a calculates a differential pressure dp=p2—p1 between the second in-cylinder pressure P2 and the first in-cylinder pressure P1 (S4), and adds dP to the variable S (S5). Further, the square value of dP is added to the variable Q (S6). The torque fluctuation estimating unit 122a repeats the processing from step S2 to step S6 for N times (for example, n=100) with a predetermined number of cycles, and obtains the integrated value of the pressure difference dP for N cycles as the variable S. Further, the torque fluctuation estimating unit 122a obtains an integrated value of squares of the pressure difference dP for N cycles as the variable Q.

Next, the torque fluctuation estimating unit 122a divides the integrated value S of the pressure difference dP by a predetermined number of cycles N to obtain a cycle average value dpmeasan of the pressure difference dP (S7).

Next, the torque fluctuation estimating unit 122a obtains a cyclic fluctuation coefficient Cov of dP of the differential pressure dP (S8). The cyclic variation coefficient Cov of dP of the differential pressure dP is a parameter obtained by normalizing the cyclic standard deviation of the differential pressure dP to the cyclic average value dpmeasan of the differential pressure dP and expressing the normalized value as a percentage, and is obtained by the following expression (1). The coefficient of variation is denoted by Cov (Coefficient of variation).

[ mathematics 1]

CoV of dP＝[{Q/N-dPmean ² } ^1/2 /dPmean]×100 ···(1)

Next, the torque fluctuation estimating unit 122a calculates the magnitude Cov of IMEP of the engine torque fluctuation from the cyclic fluctuation ratio Cov of dP of the differential pressure dP (S9). Here, a method of calculating the magnitude of the engine torque variation in step S9 will be described with reference to fig. 4.

Fig. 4 is a characteristic diagram showing the correlation between the cyclic variation rate Cov of dP [% ] of the differential pressure dP and the magnitude of the engine torque variation (Cov of IMEP [% ]). In this characteristic diagram, the actual measurement results of the cycle fluctuation ratio Cov of dP of the differential pressure dP and the magnitude of the engine torque fluctuation (Cov of IMEP) are shown using the spark ignition engine.

The magnitude of the engine torque variation is defined as the normalized standard deviation of the indicated mean effective pressure IMEP (Indicated Mean Effective Pressure: indicated mean effective pressure) over N cycles with the mean value of IMEP over N cycles, i.e., cov of IMEP [% ].

As shown in fig. 4, the inventors of the present application have found that a strong correlation can be obtained between Cov of IMEP, which indicates the magnitude (degree) of the engine torque fluctuation, and the cyclic fluctuation ratio Cov of dP of the differential pressure dP. Therefore, the inventors of the present application can estimate the magnitude of the engine torque variation from the cyclic variation rate Cov of dP of the differential pressure dP.

The correlation between Cov of IMEP and Cov of dP obtained by the calibration in advance is held as correlation or table data in the storage unit 123 of the controller 12. The torque fluctuation estimating unit 122a refers to the correlation or table data from the storage unit 123, and obtains the magnitude Cov of IMEP of the engine torque fluctuation from Cov of dP.

Then, the torque fluctuation estimating unit 122a outputs the magnitude Cov of IMEP of the obtained engine torque fluctuation to the engine control unit 122b (S10). After a predetermined time, the process of the present flow is performed again from step S1.

Here, the timing of detecting in-cylinder pressure will be described with reference to fig. 5.

Fig. 5 is a graph showing an example of detection timing of in-cylinder pressure. In the graph shown in fig. 5, the horizontal axis represents crank angle [ deg ] and the vertical axis represents in-cylinder pressure [ bar ].

Fig. 5 is a graph (1) showing an example of conventional detection timing of in-cylinder pressure. As is clear from the case of the change in-cylinder pressure shown in the graph (1), the crank angle is around 360 ° and the in-cylinder pressure is maximum.

The conventional controller disclosed in patent document 1 acquires in-cylinder pressure detected by an in-cylinder pressure sensor at a high sampling rate (for example, at intervals of 1 degree or less in crank angle) and estimates a combustion state in a combustion chamber. The circles attached along the graph in the figure represent the timing of detection of the in-cylinder pressure. However, the conventional method of acquiring in-cylinder pressure at a high sampling rate by the controller not only imposes an operational load on the controller but also has to increase the storage capacity for storing the in-cylinder pressure acquired by the controller. Therefore, if the conventional method is adopted, the cost of the engine system increases.

In addition, with the conventional controller disclosed in patent document 2, the combustion state of the engine is grasped using the in-cylinder pressures detected at two points, the crank angle position of 60 ° before the top dead center and the crank angle position of 60 ° after the top dead center, but the acquisition timing of the in-cylinder pressure is different from that according to the present embodiment. Therefore, in order to grasp the combustion state by the technique disclosed in patent document 2, complicated calculation is required, and the combustion state cannot be accurately estimated at low cost.

[ pressure detection timing ]

Fig. 5 is a graph (2) showing an example of the detection timing of the in-cylinder pressure according to the present embodiment. The change in-cylinder pressure shown in the graph (2) is the same as the change in-cylinder pressure shown in the graph (1). In the figure, an ignition timing indicated by an ignition mark and a combustion period indicated by a thick line are newly added.

In the estimation processing of the combustion state performed by the controller 12 according to embodiment 1, the in-cylinder pressure is acquired at two timings in one cycle. The timing at both points is the timing at which the first in-cylinder pressure P1 is acquired in the compression stroke ("acquisition of the pressure P1" in the drawing) and the timing at which the second in-cylinder pressure P2 is acquired in the expansion stroke ("acquisition of the pressure P2" in the drawing). The timing at which the controller 12 acquires the first in-cylinder pressure P1 is near the ignition timing. Further, the timing at which the controller 12 acquires the second in-cylinder pressure P2 is near the combustion end timing. The timing of the acquisition of the in-cylinder pressure is thus specified because the amount of combustion energy generated during combustion is reflected as the difference between the in-cylinder pressure before and after combustion.

For example, when the acquisition timing of the first in-cylinder pressure P1 is earlier than the ignition timing, the energy of the compressed in-cylinder gas from the pressure acquisition timing to the ignition timing as the combustion start timing is reflected in the difference between the first in-cylinder pressure P1 and the second in-cylinder pressure P2. Further, for example, in the case where the acquisition timing of the second in-cylinder pressure P2 is later than the combustion end timing, the energy of the in-cylinder gas expansion from the combustion end timing to the pressure acquisition timing is reflected in the difference between the first in-cylinder pressure P1 and the second in-cylinder pressure P2.

The above-described compression energy and expansion energy are supplied in excess of the energy generated by combustion, and cause of the estimation error in the estimation process of the combustion state according to embodiment 1. Therefore, in order for controller 12 to accurately estimate the combustion state, it is preferable to make the acquisition timing of first in-cylinder pressure P1 as close as possible to the ignition timing, and it is preferable to make the acquisition timing of second in-cylinder pressure P2 as close as possible to the combustion end timing. Then, the controller 12 determines the combustion state (ignition timing, etc.) of the engine 1 based on the first in-cylinder pressure P1 and the second in-cylinder pressure P2 acquired at the two timings. Therefore, the computational load of the controller 12 can be reduced, and the memory capacity for storing the in-cylinder pressure can be reduced.

Fig. 6 is a characteristic diagram showing an example of a presumption error of a fluctuation of the engine torque with respect to a difference (ms) between the acquisition timing and the ignition timing of the first in-cylinder pressure P1. The ignition timing refers to the timing at which discharge for ignition is started in the ignition plug 17 b. This characteristic map shows how the error [% ] between the estimated value and the measured value of the magnitude of the engine torque fluctuation according to embodiment 1 changes with respect to the difference between the acquisition timing and the ignition timing of the first in-cylinder pressure P1.

As shown in fig. 6, as the difference between the acquisition timing and the ignition timing of the first in-cylinder pressure P1 becomes larger, the estimation error of the magnitude of the engine torque variation increases. Therefore, an allowable error that allows the estimation error of the magnitude of the engine torque variation is set. In order to reduce the estimation error of the magnitude of the engine torque fluctuation to an error required for engine control, the difference between the acquisition timing of the first in-cylinder pressure P1 and the ignition timing is preferably 1ms or less. Therefore, the combustion state estimating unit (torque fluctuation estimating unit 122 a) sets the difference between the acquisition timing of the first in-cylinder pressure P1 and the ignition timing of the cylinder to 1ms or less.

Fig. 7 is a diagram showing a preferred variation of the acquisition timing of the first in-cylinder pressure P1 with respect to the variation of the ignition timing. In fig. 7, the broken line indicates that the ignition timing and the acquisition timing of the first in-cylinder pressure P1 are the same. Further, the solid line indicates that the difference between the ignition timing and the acquisition timing of the first in-cylinder pressure P1 is 1ms. That is, the range sandwiched by the two solid lines is the desired acquisition timing of the first in-cylinder pressure P1. Therefore, the combustion state estimating unit (torque fluctuation estimating unit 122 a) changes the acquisition timing of the first in-cylinder pressure P1 in accordance with the ignition timing of the cylinder.

Fig. 8 is a characteristic diagram showing an example of a presumption error of a difference (°) between the acquisition timing of the engine torque variation with respect to the second in-cylinder pressure P2 and a timing delayed by 20 ° from the 90% combustion timing CA90 (hereinafter simply referred to as "ca90+20°"). From this characteristic map, it is shown how the error [% ] between the estimated and measured value of the magnitude of the engine torque variation changes with respect to the difference between the acquisition timing of the second in-cylinder pressure P2 and ca90+20°. Therefore, the combustion state estimating section (torque fluctuation estimating section 122 a) changes the acquisition timing of the second in-cylinder pressure P2 in correspondence with the 90% combustion timing CA90 of the cylinder.

As shown in fig. 8, as the difference between the acquisition timing of the second in-cylinder pressure P2 and ca90+20° becomes larger, the estimation error of the magnitude of the engine torque variation increases. Therefore, an allowable error that allows the estimation error of the magnitude of the engine torque variation is set. In order to reduce the estimation error of the magnitude of the engine torque fluctuation to an error required for engine control, it is preferable that the difference between the acquisition timing of the second in-cylinder pressure P2 and ca90+20° is set to 10 ° or less.

Fig. 9 is a characteristic diagram showing a relationship between the combustion timing and the integrated value of the heat generation rate.

The 90% combustion timing CA90 shown in the drawing is defined as a crank angle (°) at which the heat generation rate integrated value is 90% when the integrated value of the heat generation rate at the end of combustion is set to 100%. Also, 50% combustion timing CA50 is defined as a crank angle at which the integrated value of the heat generation rate is 50%. The combustion timing CA50 is also referred to as a reference combustion center of gravity CA50ref.

Since the combustion speed varies in each cycle, the 90% combustion timing also varies in each cycle. Therefore, the 90% combustion timing CA90 according to embodiment 1 shown in fig. 9 represents the 90% combustion timing averaged over a predetermined number of cycles (for example, 100 cycles).

Fig. 10 is a diagram showing a preferred variation of the acquisition timing of the second in-cylinder pressure P2 with respect to the variation of ca90+20°. In fig. 10, the broken line indicates that the timing of acquisition of the ca90+20° and the second in-cylinder pressure P2 is the same. Further, the solid line indicates that the difference between ca90+20° and the acquisition timing of the second in-cylinder pressure P2 is 10 °. That is, the range sandwiched by the two solid lines is the desired acquisition timing of the second in-cylinder pressure P2. Therefore, the combustion state estimating unit (torque fluctuation estimating unit 122 a) sets the difference between the acquisition timing of the second in-cylinder pressure P2 and the 20 ° retard timing (90% combustion timing ca90+20°) of the 90% combustion timing CA90 of the cylinder to within 10 °.

In general, the 90% combustion timing CA90 varies with engine operating state parameters such as engine speed, load, air-fuel ratio, EGR rate, ignition timing, cooling water temperature, etc. In the storage 123 of the controller 12, 90% combustion timing CA90 with respect to the engine operation state parameter, which is previously determined by calibration, is stored as map data. The torque fluctuation estimating unit 122a may determine the 90% combustion timing CA90 based on the current engine operation state parameter by referring to the map data.

[ EGR System ]

Next, an example of engine control by the engine control unit 122b will be described.

For example, in an EGR system constituting the engine 1, the controller 12 needs to appropriately control the EGR rate in order to improve the thermal efficiency of the engine 1. In general, when the EGR rate is increased in the partial load, the pumping loss is reduced and the thermal efficiency is increased. In addition, since the combustion temperature is lowered by increasing the EGR rate, the cooling loss and the emission of NOx can also be reduced. In addition, knocking can be suppressed and exhaust loss can be reduced by increasing the EGR rate under high load.

On the other hand, if the EGR rate is too high, the ignitability of the mixture gas becomes low or the flame propagation property is lowered, and thus the possibility of occurrence of misfire becomes high. Therefore, in a range where no misfire occurs or in a range where a misfire is allowable, it is important for the controller 12 to increase the EGR rate as much as possible to increase the thermal efficiency of the engine 1.

If there is a cycle of misfire during operation of the engine 1, the engine torque variation becomes large. Therefore, the controller 12 estimates the magnitude of the engine torque variation, and changes the EGR rate based on the magnitude of the engine torque variation, so that the thermal efficiency of the engine can be improved while suppressing misfire.

Fig. 11 shows a diagram of an example of control modules of the controller 12 that performs EGR control. The torque fluctuation estimating unit 122a estimates the current magnitude Cov of IMEP of the engine torque fluctuation based on the output of the pressure sensor 10 of the engine 1.

The deviation calculating unit 122c included in the engine control unit 122b calculates a deviation δcov obtained by subtracting the magnitude of the target engine torque variation (target Cov) from the magnitude Cov of the current engine torque variation Cov of IMEP. The target CoV is a value read from the ROM of the storage unit 123 by the engine control unit 122 b.

The operation amount calculation unit 122d included in the engine control unit 122b calculates the operation amount of the actuator of the engine 1 based on the deviation δcov. The actuator is, for example, a device provided in the engine 1 for adjusting the opening degree of the throttle valve 20, the EGR valve 29, or adjusting the ignition timing of the ignition plug 17 b. The operation amount calculation unit 122d is constituted by a PID controller, for example. The operation amount calculation unit 122d obtains the operation amount of the actuator of the engine 1 so that the current magnitude Cov of IMEP of the engine torque variation approaches the target magnitude of the engine torque variation (target Cov). Then, the engine control portion 122b transmits the actuator operation amount of the engine 1 to the engine 1, and controls the operating state of the engine 1.

(example of control of an actuator in an EGR System)

Fig. 12 is a diagram showing a control example of an actuator based on the deviation δcov in the EGR system. The horizontal axis represents the deviation δcov [% ], and the vertical axis represents the state of the actuator or the like.

When the actuator is controlled based on the deviation δcov, the controller 12 of the EGR system suppresses the magnitude of the engine torque variation with an increase in the deviation δcov, for example. Therefore, the controller 12 controls such that the opening degree of the EGR valve 29 (broken line) and the opening degree of the throttle valve 20 (solid line) are reduced. By this control, the EGR rate becomes low, and therefore the ignition delay time becomes short, and the combustion speed becomes high. Therefore, in order to bring the combustion to an appropriate timing (fuel consumption optimum timing), the controller 12 controls such that the amount of ignition advance (one-dot chain line) becomes small.

When the magnitude of the engine torque fluctuation is equal to or greater than a predetermined value under the control of the controller 12, the EGR rate is set low to suppress the torque circulation fluctuation. Thereby, the combustion of the engine 1 is controlled in a stable direction. When the magnitude of the engine torque fluctuation is smaller than the predetermined value, the controller 12 can set the EGR rate high, thereby improving the thermal efficiency of the engine 1.

In addition, it is also conceivable to have a structure in which the amount of ignition energy supplied to the ignition plug 17b, the intensity of gas flow in the cylinder, the compression ratio, the intake air temperature can be adjusted, and the controller 12 can control these adjustment amounts based on the deviation δcov.

Fig. 13 is a diagram showing an example in which the controller 12 controls the amount of ignition energy, the intensity of gas flow, the compression ratio, and the intake air temperature based on the deviation δcov. The horizontal axis of fig. 13 represents the deviation δcov [% ], and the vertical axis represents the ignition energy and the like.

The higher the ignition energy amount, the strength of the gas flow in the cylinder, the compression ratio, and the intake air temperature are, the more the ignition and flame propagation are promoted, and the torque fluctuation is suppressed. Therefore, as shown in fig. 13, the controller 12 preferably controls the increase of the deviation δcov in the direction in which the values of these items increase.

For example, the amount of ignition energy may be adjusted by controlling the amount of current provided to the spark plug 17b by the controller 12. Further, the intensity of the gas flow in the cylinder may be adjusted by controlling the flow rate of air in the intake port 21 by the controller 12. Further, controller 12 may adjust the compression ratio by controlling the position of top dead center of piston 14. Further, the controller 12 may adjust the intake air temperature by controlling the switching of the heater provided in the intake port 21.

In addition, in the above control, the controller 12 may control any one of the intensity of the gas flow, the compression ratio, and the intake air temperature individually, or may control in combination of several. In addition to these controls, the controller 12 may combine the controls of the EGR valve opening degree, the throttle valve opening degree, or the amount of ignition advance described above.

[ lean burn System ]

In addition, in the lean-burn system constituting the engine 1, it is also necessary to appropriately control the air-fuel ratio in order to improve the thermal efficiency of the engine 1. In general, when the air-fuel ratio is increased in the partial load, the pumping loss decreases and the thermal efficiency increases. In addition, since the combustion temperature is lowered by increasing the air-fuel ratio, the cooling loss and the emission of NOx can also be reduced. On the other hand, if the air-fuel ratio is too high, the combustibility of the mixture gas becomes low or the flame propagation property decreases, and thus the possibility of occurrence of misfire becomes high. Therefore, in a range where no misfire occurs or in a range where a misfire is allowable, it is important for the controller 12 to perform control to raise the air-fuel ratio as much as possible to raise the thermal efficiency of the engine 1.

(example of control of an actuator in a lean burn System)

Fig. 14 is a diagram showing a control example of an actuator based on the deviation δcov in the lean burn system. The horizontal axis of fig. 14 represents the deviation δcov [% ], and the vertical axis represents the state of the actuator or the like.

The controller 12 in the lean-burn system controls the actuators based on the deviation δcov. For example, in order to suppress the cyclical variation of the torque, the controller 12 controls the actuator such that the opening degree (solid line) of the throttle valve 20 becomes smaller as the deviation δcov increases. By this control, the air-fuel ratio becomes low, so the ignition delay time becomes short, and the combustion speed becomes high. Therefore, in order to bring the combustion to an appropriate timing (fuel consumption optimum timing), the controller 12 controls the actuator so that the amount of ignition advance (one-dot chain line) becomes small.

By this control, when the magnitude of the engine torque variation is equal to or greater than a predetermined value, the air-fuel ratio is set low to suppress the cyclical variation of the torque. By setting the air-fuel ratio low, the combustion of the engine 1 is controlled in a stable direction. When the magnitude of the engine torque fluctuation is smaller than a predetermined value, the air-fuel ratio is set high, and therefore the thermal efficiency can be improved.

The control of the amount of ignition energy, the intensity of gas flow in the cylinder, the compression ratio, and the intake air temperature shown in fig. 13 can also be applied to the lean burn system in the same manner as the EGR system described above.

< embodiment 2>

[ estimation of Combustion barycenter ]

Next, a description will be given of a controller according to embodiment 2 of the present invention that estimates a combustion center of gravity CA50 based on a first in-cylinder pressure P1 acquired near an ignition timing and a second in-cylinder pressure P2 acquired near a combustion end timing.

As shown in fig. 9, the combustion center of gravity CA50 is defined as a crank angle at which the integrated value of the heat generation rate at the end of combustion is 50% when the integrated value of the heat generation rate is set to 100%. The combustion centroid CA50 according to embodiment 2 shows the combustion centroid obtained by averaging the combustion centroid at a predetermined number of cycles (for example, 100 cycles), although the combustion centroid changes for each cycle according to the cycle variation of the combustion speed.

Fig. 15 is a block diagram showing a configuration example of controller 12A according to embodiment 2.

The controller 12A includes an input/output unit 121, a control unit 124, and a storage unit 123 electrically connected to each other via a system bus not shown.

The input/output unit 121 and the storage unit 123 are similar to the input/output unit 121 and the storage unit 123 of the controller 12 in embodiment 1, and therefore, detailed description thereof is omitted.

The control unit 124 controls the operation of the engine 1. For example, the control unit 124 controls the throttle opening, the EGR opening, the fuel injection amount, and the ignition timing according to the combustion steady state of the engine 1. The control unit 124 according to embodiment 2 includes a combustion center of gravity estimating unit 124a and an engine control unit 124b.

The combustion centroid estimating unit 124a estimates the combustion centroid CA50 based on the in-cylinder pressure detected by the pressure sensor 10. The combustion state of the engine 1 estimated in the present embodiment is the combustion center of gravity CA50.

The engine control portion (engine control portion 122 b) controls the ignition timing based on the ratio of the first in-cylinder pressure P1 to the second in-cylinder pressure P2. For example, the engine control unit 124b controls the ignition timing of the engine 1 based on the combustion center of gravity CA50 obtained by the combustion center of gravity estimating unit 124 a.

[ principle of estimation of Combustion barycenter ]

The principle of estimating the combustion center of gravity will be described below with reference to fig. 16 to 18.

Fig. 16 is a characteristic diagram showing the correlation between the reference combustion center of gravity CA50ref (°) and the ratio of the first in-cylinder pressure P1 to the second in-cylinder pressure P2. The characteristic map shows the actual measurement results using the pressure ratio P2/P1 and the reference combustion center of gravity CA50ref of the spark ignition engine. The combustion center of gravity estimating portion 124a acquires the first in-cylinder pressure P1 near the ignition timing and acquires the second in-cylinder pressure P2 near the combustion end timing by the pressure sensor 10. Fig. 16 shows the measurement results of the volumetric efficiency of the engine 1, the pressure ratio P2/P1 after the engine speed is fixed, and the reference combustion center of gravity CA50ref. Here, the volumetric efficiency is defined as a reference volumetric efficiency ηref, and the rotational speed is defined as a reference rotational speed Nref.

According to the new findings of the inventors of the present application, it is found that there is a strong correlation between the reference combustion center of gravity CA50ref and the ratio P2/P1 of the first in-cylinder pressure P1 acquired near the ignition timing and the second in-cylinder pressure P2 acquired near the combustion end timing, as shown in fig. 16. Therefore, the inventors of the present application know that the reference combustion center of gravity CA50ref can be estimated from the pressure ratio P2/P1.

Further, according to the new findings of the inventors of the present application, it is found that when the volumetric efficiency is different from the reference volumetric efficiency ηref and the engine speed is different from the reference speed Nref, the combustion centroid estimating unit 124a corrects the reference combustion centroid CA50ref estimated by using the reference volumetric efficiency ηref and the reference speed Nref, thereby estimating the combustion centroid CA50 at the current speed and the volumetric efficiency.

Fig. 17 is a diagram showing an example of the correction amount Δca50_1 (°) from the reference combustion center of gravity ca50ref against the change in the volumetric efficiency (%).

Fig. 18 is a diagram showing an example of the correction amount Δca50—2 (°) from the reference combustion center of gravity CA50ref against a change in the engine speed (1/min).

Based on these correction amounts Δca50_1 and Δca50_2, the combustion center of gravity CA50 at the current rotation speed and volumetric efficiency is found by the following equation (2).

[ math figure 2]

CA 50＝CA 50 ref+ΔCA50_1+ΔCA50_2 ···(2)

The combustion centroid estimating unit 124a obtains the correction amount of the reference combustion centroid CA50ref from each reference value, for example, for not only the volumetric efficiency and the rotation speed, but also the EGR rate, the air-fuel ratio, the ignition timing, the cooling water temperature, and the like, and corrects the reference combustion centroid CA50ref to estimate the combustion centroid CA50 with higher accuracy.

Next, the estimation process of the combustion center of gravity CA50 performed by the controller 12A will be described with reference to fig. 19.

Fig. 19 is a flowchart showing an example of the combustion center of gravity estimation step performed by the combustion center of gravity estimation portion 124a of the controller 12A.

First, the combustion barycenter estimation unit 124a initializes the variable R to zero (S21). Next, the combustion center of gravity estimating portion 124a acquires the first in-cylinder pressure P1 near the ignition timing detected by the pressure sensor 10 (S22). Next, the combustion center of gravity estimating portion 124a acquires a second in-cylinder pressure P2 near the combustion end timing detected by the pressure sensor 10 (S23).

Next, the combustion center of gravity estimation portion 124a obtains a ratio pr=p2/P1 of the second in-cylinder pressure P2 to the first in-cylinder pressure P1 (S24), and adds the ratio PR to the variable R (S25). The combustion center of gravity estimation unit 124a repeats a predetermined cycle number N (for example, n=100) from step S22 to step S25, and obtains an integrated value of the pressure ratio PR for N cycles in R.

After N cycles, the combustion center of gravity estimation unit 124a divides the cumulative value R of PR by a predetermined number of cycles N to obtain a cycle average value PRmean of the pressure ratio PR (S26).

Next, the combustion centroid estimating unit 124a calculates the reference combustion centroid CA50ref from the average value PRmean of the pressure ratios PR (S27). The correlation between the pressure ratio PR and the reference combustion center of gravity CA50ref, which is obtained by the calibration in advance, is held in the memory 123 of the controller 12A as correlation or table data. The combustion centroid estimating unit 124a calculates the reference combustion centroid CA50ref from the average value Prmean of the pressure ratios PR by referring to the correlation or table data.

Next, the combustion centroid estimating unit 124a calculates the combustion centroid CA50 from the reference combustion centroid CA50ref and the correction value Δca50 of the combustion centroid. The memory 123 of the controller 12A stores, as correlation or table data, a correlation between the volumetric efficiency and the combustion center of gravity correction value, a correlation between the rotational speed and the combustion center of gravity correction value, and the like, which are obtained in advance by calibration. The combustion centroid estimating unit 124a obtains the sum of the correction values Δca50 of the combustion centroid by referring to the correlation or table data, and adds them to the reference combustion centroid CA50ref to obtain the combustion centroid CA50 (S28).

Finally, the combustion centroid estimating unit 124a transmits the obtained combustion centroid CA50 to the engine control unit 124b (S29).

[ Engine control ]

Next, an example of engine control by the engine control unit 124b will be described.

In order to improve the thermal efficiency of the engine, the combustion center of gravity needs to be set appropriately.

Fig. 20 is a diagram showing a general relationship between the optimum combustion center of gravity CA50 and the fuel consumption rate of the engine.

In general, if the combustion center of gravity is advanced from the proper timing, the compression work of the piston increases and the thermal efficiency decreases. In addition, if the combustion center of gravity is retarded from the proper timing (target), the exhaust energy increases and the thermal efficiency decreases.

Generally, the engine sets the ignition timing in advance for each load and rotation speed so that the center of gravity of combustion is at the optimum timing. However, it is considered that the combustion center of gravity may shift from the optimum timing due to a change in environmental conditions, a change in engine component characteristics with time, or the like. Therefore, the engine control unit 124b sets the combustion center of gravity at a predetermined optimum timing by controlling the ignition timing, so that the thermal efficiency of the engine can be maintained high even when the environmental conditions change or the engine component characteristics change with time.

Fig. 21 shows an example of a control module of the controller 12A that performs EGR control.

The combustion center of gravity estimation unit 124a estimates the current combustion center of gravity CA50 based on the output (in-cylinder pressure) of the pressure sensor 10 provided in the engine 1.

The deviation calculating unit 124c included in the engine control unit 124b obtains a deviation δca50 between the current combustion centroid CA50 and the target combustion centroid CA50 (referred to as "target CA50" in the drawing).

The ignition timing calculation unit 124d included in the engine control unit 124b calculates the ignition timing of the engine 1 based on the deviation δca50. The ignition timing calculation unit 124d is configured by, for example, a PID controller, and obtains the ignition timing of the engine 1 so that the current combustion center of gravity approaches the target combustion center of gravity. Then, the engine control unit 124b transmits the ignition timing obtained by the ignition timing calculation unit 124d to the engine 1, and operates the engine 1 at the new ignition timing.

In the present embodiment, the estimation method of the combustion center of gravity CA50 is shown, but the present invention is not limited to the combustion center of gravity CA50. The combustion state estimating unit can estimate the combustion phase based on the ratio of the first in-cylinder pressure P1 to the second in-cylinder pressure P2. For example, the combustion state estimating unit can estimate the combustion phase of the 10% combustion timing CA10, the 90% combustion timing CA90, or the like using the ratio P2/P1 of the first in-cylinder pressure P1 acquired near the ignition timing to the second in-cylinder pressure P2 acquired near the combustion end timing.

The pressure ratio is not limited to P2/P1, and the pressure ratio may be P1/P2. Further, even if the pressure difference Δp=p2-p1 or Δp=p1-p2 is used instead of the pressure ratio, the combustion phase can be estimated in the same manner. An example of the pressure difference Δp=p2-P1 is described with reference to fig. 22.

Fig. 22 is a characteristic diagram showing a correlation between the difference between the first in-cylinder pressure P1 and the second in-cylinder pressure P2 and the reference combustion center of gravity CA50ref (°).

The characteristic map shows the actual measurement results using the pressure difference P2-P1 of the spark ignition engine and the reference combustion center of gravity CA50 ref. Here, the first in-cylinder pressure P1 is acquired near the ignition timing and the second in-cylinder pressure P2 is acquired near the combustion end timing by the pressure sensor 10. Fig. 22 shows the result of measuring the pressure difference P2-P1 and the reference combustion center of gravity CA50ref by keeping the volumetric efficiency and the engine rotational speed of the engine 1 constant. Here, the volumetric efficiency is defined as a reference volumetric efficiency ηref, and the rotational speed is defined as a reference rotational speed Nref.

According to the new findings of the inventors of the present application, it is known that, as shown in fig. 22, there is a strong correlation between the reference combustion center of gravity CA50ref and the pressure difference P2-P1 between the first in-cylinder pressure P1 acquired near the ignition timing and the second in-cylinder pressure P2 acquired near the combustion end timing. Therefore, the inventors of the present application have known that the reference combustion center of gravity CA50ref can be estimated from the pressure difference P2-P1.

< example of Using discharge Voltage to find in-Cylinder pressure >

Although the controllers 12 and 12A according to the above-described embodiments have shown an example in which the pressure sensor 10 is used to detect the in-cylinder pressure, a method other than the pressure sensor 10 may be used to determine the in-cylinder pressure.

For example, controller 12 may determine the in-cylinder pressure from the discharge voltage of ignition coil 17 a. Therefore, the combustion state estimating unit (pressure calculating unit 31 a) can calculate the first in-cylinder pressure P1 and the second in-cylinder pressure P2 based on the voltage value of the ignition coil (ignition coil 17 a), the current value of the ignition coil (ignition coil 17 a), or the discharge time of the ignition coil (ignition coil 17 a).

Fig. 23 is a diagram showing an example of a time history of the ignition signal, the primary and secondary voltages of the ignition coil 17a, and the secondary current. In the ignition coil 17a, an ignition signal is generated from the controller 12. The voltage on the primary side (primary voltage), the voltage on the secondary side (secondary voltage), and the current on the secondary side (secondary current) of the ignition coil 17a change with time. At the timing of shifting the ignition signal from the high level to the low level, the secondary current starts to change until the secondary current returns to the original value, and the above period is defined as a discharge period.

Fig. 24 is a characteristic diagram showing the correlation of the maximum value V1max of the primary voltage immediately after the start of discharge, the maximum value V2max of the secondary voltage, and the in-cylinder pressure at the time of discharge. Fig. 24 shows the case of the ignition signal and the primary voltage and the secondary current of the ignition coil 17 a.

When the ignition signal transitions from the high level to the low level, a large potential difference is generated on the secondary side of the ignition coil 17a, and discharge is started by the ignition plug 17 b. As shown in fig. 24, the maximum value V2max (kV) of the secondary voltage immediately after the start of discharge and the in-cylinder pressure at the time of discharge show a high correlation. Therefore, the in-cylinder pressure at the time of discharge can be obtained by measuring the maximum value V2max of the secondary voltage.

Further, a counter electromotive force is generated on the primary side of the ignition coil 17a immediately after the discharge of the ignition plug 17b, and as shown in fig. 24, the maximum value V1max (V) of the primary voltage immediately after the start of the discharge also has a high correlation with the in-cylinder pressure at the time of the discharge. Therefore, the controller 12 can also determine the in-cylinder pressure at the time of discharge by measuring the maximum value V1max of the primary voltage of the ignition coil 17 a.

Fig. 25 is a block diagram showing a configuration example from ignition coil 17a to controller 12 in the case where controller 12 obtains the in-cylinder pressure based on the secondary voltage of ignition coil 17 a.

An internal combustion engine (engine 1) includes: an ignition coil (ignition coil 17 a); and a voltage peak holding section (peak holding circuit 30) that detects the voltage of the ignition coil (ignition coil 17 a) and holds the peak value of the voltage of the ignition coil (ignition coil 17 a). Then, the combustion state estimating section (pressure calculating section 31 a) calculates the first in-cylinder pressure P1 and the second in-cylinder pressure P2 based on the peak values.

The secondary voltage of the ignition coil 17a is output to the peak hold circuit 30, and the peak hold circuit 30 holds the peak value of the secondary voltage measured in a predetermined time, and the peak value of the secondary voltage is detected as the maximum value V2max of the secondary voltage. The peak hold circuit 30 may be provided in a circuit for supplying an ignition signal to the ignition coil 17 a.

The peak hold circuit 30 sends the maximum value V2max of the secondary voltage to the controller 12. The pressure calculation unit 31a in the controller 12 calculates the in-cylinder pressure at the time of discharge using correlation or table data between the secondary voltage maximum value V2max and the in-cylinder pressure, which are obtained by calibration in advance.

The pressure calculation unit 31a performs a process of obtaining the in-cylinder pressure from the primary voltage of the ignition coil 17a in the same manner as the process of obtaining the in-cylinder pressure from the secondary voltage. That is, the maximum value V1max of the primary voltage of the ignition coil 17a is detected by the peak hold circuit 30, and output to the controller 12. The pressure calculation unit 31a provided in the controller 12 calculates the in-cylinder pressure at the time of discharge using correlation or table data between the primary voltage maximum value V1max and the in-cylinder pressure, which are obtained by calibration in advance.

When the in-cylinder pressure detection by the ignition coil 17a is applied to the estimation method of the combustion state according to the present embodiment, first, the pressure calculation unit 31a obtains the first in-cylinder pressure P1 based on the secondary voltage maximum value V2max or the primary voltage maximum value V1max associated with the discharge of the ignition timing. Further, near the combustion end timing, the controller 12 sends an ignition signal to the ignition coil 17a, and the ignition coil 17a discharges. The second discharge is performed at the timing when the second in-cylinder pressure P2 is obtained, but does not ignite at this timing. The pressure calculation unit 31a obtains the second in-cylinder pressure P2 based on the secondary voltage maximum value V2max or the primary voltage maximum value V1max of the ignition coil 17a at the time of the discharge.

When the pressure calculation unit 31a obtains the in-cylinder pressure based on the voltage of the ignition coil 17a, the ignition coil 17a needs to be charged and discharged, and therefore the detection of the in-cylinder pressure needs to be separated by a predetermined interval (for example, 4 ms) or more. However, in the method for detecting a combustion state according to the present embodiment, the pressure calculating portion 31a may detect the in-cylinder pressure in the vicinity of the ignition timing and the vicinity of the combustion end timing. In general, the interval between the ignition timing and the combustion end timing is longer than the charge-discharge time of the ignition coil 17 a. Therefore, it is preferable to apply the method of detecting the in-cylinder pressure based on the voltage of the ignition coil 17a to the combustion state estimating method according to the present embodiment.

Thus, if the pressure calculation unit 31a detects the in-cylinder pressure based on the voltage of the ignition coil 17a, the pressure sensor 10 is not required, and the cost of the engine system can be reduced. In addition, the following advantages are provided: a space for mounting the pressure sensor 10 on the engine is not required, and thus the degree of freedom in designing the engine such as the cooling water passage and the shape of the combustion chamber becomes large.

(example of obtaining in-cylinder pressure Using discharge period)

The in-cylinder pressure may be obtained using not only the maximum voltage of the ignition coil 17a but also the discharge period. In general, when the in-cylinder pressure increases, the discharge period of the ignition coil 17a becomes shorter. Fig. 26 shows a case where the period from the start of discharge to the secondary current of the ignition coil 17a becoming equal to or less than a predetermined value is a discharge period.

Fig. 26 is a characteristic diagram showing a correlation between the in-cylinder pressure during discharge and at the time of discharge. The characteristic diagram shows the actual measurement results of the discharge period [ ms ] and the in-cylinder pressure at the time of discharge using the spark ignition engine.

According to the novel findings of the inventors of the present application, as shown in fig. 26, a strong correlation can be obtained between the discharge period and the in-cylinder pressure at the time of discharge. The pressure calculation unit 31a can determine the in-cylinder pressure by using the correlation between the in-cylinder pressure and the discharge period of the ignition coil 17 a. Therefore, the combustion center of gravity CA50 can be estimated based on the first in-cylinder pressure P1 acquired from the discharge timing in the vicinity of the ignition timing and the second in-cylinder pressure P2 acquired from the vicinity of the combustion end timing.

(example of obtaining in-cylinder pressure Using angular acceleration of crankshaft)

The combustion state estimating unit (pressure calculating unit 31 b) can calculate the first in-cylinder pressure P1 and the second in-cylinder pressure P2 based on the angular acceleration of the crankshaft.

The angular acceleration of the crank shaft is represented by the following formula (3). Omega represents crank rotation speed, J represents moment of inertia, T _e Represents combustion torque, T _L Representing the load torque. Load torque T is estimated based on the rotational speed of engine 1 _L 。

[ math 3]

In addition, combustion torque T _e Expressed by the following equation (4) as a function of the in-cylinder pressure Pc and the crank angle θ.

[ mathematics 4]

T _e ＝func.(P _c ,θ) …(4)

The controller 12 substitutes the angular acceleration dω/dt of the crank shaft at the crank angle θ into equation (3), and solves the simultaneous equation with equation (4) to determine the in-cylinder pressure Pc at the crank angle θ.

Fig. 27 is a block diagram showing a configuration example from the crank angle sensor 11 to the controller 12 in the case where the controller 12 obtains the in-cylinder pressure based on the angular acceleration of the crank shaft.

The internal combustion engine (engine 1) includes an angular acceleration calculation portion (angular acceleration calculation portion 32) for calculating an angular acceleration of the crankshaft.

The combustion state estimating unit (pressure calculating unit 31 b) calculates the first in-cylinder pressure P1 and the second in-cylinder pressure P2 based on the angular acceleration of the crankshaft.

The crank angle θ detected by the crank angle sensor 11 is sent to the angular acceleration calculating section 32. The angular acceleration calculating unit 32 calculates the angular acceleration dω/dt of the crankshaft using the crank angle θ. Then, the angular acceleration calculating section 32 sends the angular acceleration dω/dt of the crank shaft and the crank angle θ to the controller 12. The pressure calculation unit 31b in the controller 12 obtains the in-cylinder pressure by solving the simultaneous equations of the equations (3) and (4).

When in-cylinder pressure detection based on angular acceleration of the crankshaft is applied to the combustion state estimation method according to the present embodiment, the angular acceleration calculation unit 32 detects angular acceleration dω/dt of the crankshaft using the crank angle sensor 11 near the ignition timing and near the combustion end timing. Then, the pressure calculation unit 31b obtains the first in-cylinder pressure P1 near the ignition timing and the second in-cylinder pressure P2 near the combustion end timing by solving the simultaneous equations of the equations (3) and (4) using the angular acceleration dω/dt.

Thus, if the in-cylinder pressure Pc is detected based on the angular acceleration of the crankshaft, the pressure sensor 10 is not required, and the cost of the engine system can be reduced.

The controller 12 according to embodiments 1 and 2 described above can estimate the combustion state (torque fluctuation, combustion phase) with high accuracy using the in-cylinder pressure. In embodiment 1 and embodiment 2, 2 in-cylinder pressures acquired in the vicinity of the ignition timing and the vicinity of the combustion end timing for each cycle can be used for estimation of the combustion state, so that the combustion state can be accurately estimated. Further, since the pressure is calculated from 2 in-cylinder pressures, the memory and the calculation load can be reduced.

Further, since high-speed pressure sampling is not required, the controller 12 can detect in-cylinder pressure using existing engine equipment such as the ignition coil 17a, the crank angle sensor 11, and the like, even without using an in-cylinder pressure sensor. By these methods, the system cost required for estimating the combustion state can be suppressed to be low.

< others >

The present invention is not limited to the above embodiments, and various other application examples and modifications can be made without departing from the gist of the present invention described in the patent claims.

For example, the above-described embodiment describes the structure of the controller 12 in detail and in detail for easy understanding of the present invention, but is not necessarily limited to having all the components described. In addition, a part of the structure of one embodiment may be replaced with the constituent elements of another embodiment. In addition, the constituent elements of other embodiments may be added to the structure of one embodiment. In addition, other components may be added, substituted, or deleted to a part of the structure of each embodiment.

Some or all of the respective structures, functions, processing units, and the like of the controller 12 may be realized in hardware by, for example, designing with an integrated circuit. As hardware, an FPGA (Field Programmable Gate Array: field programmable gate array) or an ASIC (Application Specific Integrated Circuit: application specific integrated circuit) or the like may be used.

Control lines and information lines necessary for explanation are shown, but not limited to all control lines and information lines necessary for production. Virtually all structures can be considered interconnected.

In the flowcharts shown in fig. 3 and 19, a plurality of processes may be executed in parallel or the processing order may be changed within a range that does not affect the processing result.

Description of the reference numerals

An engine of 1 …, a pressure sensor of 10 …, a controller of 12 …, an ignition coil of 17a …, a spark plug of 17b …, a peak hold circuit of 30 …, a pressure calculation unit of 31a …, a pressure calculation unit of 31b …, an angular acceleration calculation unit of 32 …, an input/output unit of 121 …, a control unit of 122 …, a torque fluctuation estimation unit of 122a …, a control unit of 122b …, a deviation calculation unit of 122c …, an operation amount calculation unit of 122d …, and a storage unit of 123 ….

Claims

1. A control device for an internal combustion engine is characterized in that,

the engine control device includes a combustion state estimating unit that estimates a combustion state of the internal combustion engine based on a first in-cylinder pressure acquired near an ignition timing and a second in-cylinder pressure acquired near a combustion end timing.

2. The control apparatus for an internal combustion engine according to claim 1, wherein,

The combustion state estimating unit calculates a difference between the first in-cylinder pressure and the second in-cylinder pressure in a plurality of cycles, and estimates a magnitude of torque fluctuation of the internal combustion engine based on a fluctuation rate of the difference in the plurality of cycles.

3. The control apparatus for an internal combustion engine according to claim 2, wherein,

the engine control unit is configured to control an EGR rate or an air-fuel ratio based on the magnitude of the torque fluctuation.

4. The control apparatus for an internal combustion engine according to claim 2, wherein,

the engine control unit is configured to control at least one of ignition energy, in-cylinder flow strength, compression ratio, and intake air temperature based on the magnitude of the torque fluctuation.

5. The control apparatus for an internal combustion engine according to claim 1, wherein,

the combustion state estimating unit changes the timing of acquiring the first in-cylinder pressure in accordance with the ignition timing of the cylinder.

6. The control apparatus for an internal combustion engine according to claim 5, wherein,

the combustion state estimating unit sets a difference between a timing of acquiring the first in-cylinder pressure and an ignition timing of the cylinder to be within 1 ms.

7. The control apparatus for an internal combustion engine according to claim 1, wherein,

the combustion state estimating unit changes the timing of acquiring the second in-cylinder pressure in accordance with the 90% combustion timing of the cylinder.

8. The control apparatus for an internal combustion engine according to claim 7, wherein,

the combustion state estimating unit sets a difference between a timing of acquiring the second in-cylinder pressure and a 20 DEG retarded timing of 90% combustion timing of the cylinder to within 10 deg.

9. The control apparatus for an internal combustion engine according to claim 1, wherein,

the combustion state estimating unit estimates a combustion phase based on a ratio of the first in-cylinder pressure to the second in-cylinder pressure.

10. The control apparatus for an internal combustion engine according to claim 9, wherein,

an engine control portion is included that controls ignition timing based on a ratio of the first in-cylinder pressure to the second in-cylinder pressure.

11. The control apparatus for an internal combustion engine according to claim 1, wherein,

the combustion state estimating unit obtains the first in-cylinder pressure and the second in-cylinder pressure based on a voltage value of an ignition coil, a current value of the ignition coil, or a discharge time of the ignition coil.

12. The control apparatus for an internal combustion engine according to claim 1, wherein,

the internal combustion engine includes:

an ignition coil; and

a voltage peak holding section that detects a voltage of the ignition coil and holds a peak value of the voltage of the ignition coil,

the combustion state estimating section calculates the first in-cylinder pressure and the second in-cylinder pressure based on the peak value.

13. The control apparatus for an internal combustion engine according to claim 1, wherein,

the combustion state estimating unit obtains the first in-cylinder pressure and the second in-cylinder pressure based on an angular acceleration of a crankshaft.

14. The control apparatus for an internal combustion engine according to claim 1, wherein,

the internal combustion engine includes an angular acceleration calculation portion that calculates an angular acceleration of a crankshaft,

the combustion state estimating section calculates the first in-cylinder pressure and the second in-cylinder pressure based on an angular acceleration of the crankshaft.