JP2013204521A - Control device of internal combustion engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a control device of an internal combustion engine that obtains a cooling loss of the internal combustion engine in high accuracy and can optimize a control parameter according to the cooling loss.SOLUTION: An inner-cylinder-surface temperature is estimated based on a fuel integrated injection amount of each of cylinder in a predetermined period of time and a coolant temperature (Step ST2), and a piston top face temperature is estimated based on the fuel integrated injection amount in the predetermined period of time and a lubricant temperature (Step ST3). A cooling loss index reference value of each cylinder is calculated based on these inner-cylinder-surface temperature and piston top face temperature (Step ST4), and a cooling loss index of each cylinder is calculated based on this cooling loss index reference value, an intake gas amount, and an intake gas temperature (Step ST5). A correction of a control parameter of an engine is performed based on this cooling loss index (Step ST6).

Description

  The present invention relates to a control device for an internal combustion engine represented by a diesel engine or the like. In particular, the present invention relates to an improvement for optimizing control parameters of an internal combustion engine.

  As is well known in the art, in the case of a diesel engine used as an automobile engine or the like, in order to realize an improvement in fuel consumption rate, securing a high engine torque, reducing combustion noise, improving exhaust emission, etc. The control amounts (control parameters) of various control devices are adjusted according to the accelerator operation amount, the cooling water temperature, and the intake air temperature. For example, the fuel injection amount and fuel injection timing from a fuel injection valve (hereinafter sometimes referred to as “injector”) are adjusted. Further, the exhaust gas recirculation rate (EGR rate) is adjusted by controlling the opening of an EGR valve provided in an EGR (Exhaust Gas Recirculation) device that recirculates exhaust gas to the intake system for the purpose of reducing NOx emissions. The

  For example, in Patent Document 1 below, an estimated compression end temperature (in-cylinder temperature at the time when the piston reaches compression top dead center) is calculated according to the detected in-cylinder pressure, and is estimated from the target compression end temperature. It is disclosed that the pilot injection amount is corrected according to a value obtained by subtracting the compression end temperature. Patent Document 2 below discloses that when the pilot injection amount is corrected according to the cooling water temperature, the pilot injection amount is increased as the cooling water temperature is lower.

JP 2009-7966 A JP 2001-55947 A

  When the engine control parameters (for example, fuel injection timing, pilot injection amount, etc.) are corrected as described above, the cooling loss (hereinafter referred to as “cooling loss”) caused by the cooling water depriving the cylinder of heat. It is necessary to correct the control parameters in consideration of Further, the optimum value of the control amount of the control parameter varies according to the history of the engine operating state. For this reason, it is difficult to accurately recognize the current temperature environment in the cylinder only with the cooling water temperature, and the control amount is adjusted in advance so as not to deteriorate the combustion state in the cylinder. For example, each control amount is set (for example, in-cylinder compression) in consideration of the fact that engine output can be obtained with high responsiveness when re-acceleration is requested from the vehicle deceleration state (transition). Each control amount is set so that the ignition delay period becomes long when the end temperature is low and the generation timing of the requested engine output is not delayed due to the ignition timing delay). For this reason, during the steady operation, the correction amount of the control parameter may be excessive, which may lead to deterioration of the fuel consumption rate, increase of combustion noise, deterioration of exhaust emission, and the like.

  In view of this point, if the inventors of the present invention can recognize the engine cooling loss with high accuracy, the correction of the control parameter corresponding to the cooling loss is performed, so that the fuel consumption is not corrected. We focused on improving the rate, reducing combustion noise, and improving exhaust emissions.

  And it discovered that this engine cooling loss was determined by the history of the engine operating state, the cooling loss due to cooling water, and the cooling loss due to lubricating oil, leading to the present invention.

  The present invention has been made in view of the above point, and an object of the present invention is to obtain an internal combustion engine with a high degree of accuracy and to optimize the control parameters according to the cold loss. It is to provide an engine control device.

-Solution principle of the invention-
The solution principle of the present invention taken to achieve the above object is an index of the cooling loss for each cylinder of the internal combustion engine by the history of the operating state of the internal combustion engine, the cooling loss due to cooling water, and the cooling loss due to lubricating oil. A value (cooling loss index reference value) is obtained, and the control parameter of the internal combustion engine is adjusted based on this index, so that the control parameter can be optimized.

-Solution-
Specifically, the present invention is premised on a control device for an internal combustion engine that generates an output by combustion of fuel injected from a fuel injection valve into a cylinder. With respect to the control device for the internal combustion engine, the cylinder inner surface temperature estimated from the fuel injection amount toward the cylinder and the coolant temperature within a predetermined time, and the fuel injection amount toward the cylinder within the predetermined time Based on the piston surface temperature estimated from the oil temperature and the lubricating oil temperature, an index of cooling loss in one cylinder is obtained, and the internal combustion engine control parameter is corrected according to the index of cooling loss.

  With this specific matter, the cooling loss due to the cooling water is determined from the estimated value of the cylinder inner surface temperature, and the cooling loss due to the lubricating oil is determined from the estimated value of the piston surface temperature. Then, an index of cooling loss in one cylinder is obtained from these estimated values, and the internal combustion engine control parameters are corrected. For example, based on this index, when the cooling loss is large, the internal combustion engine control parameter is corrected so that the temperature in the cylinder can be increased (combustion efficiency is increased). By correcting the internal combustion engine control parameter in this manner, the cooling loss in one cylinder can be obtained with high accuracy, and the internal combustion engine control parameter reflecting that can be corrected. As a result, the operating state of the internal combustion engine suitable for the current requirements can be obtained without overcorrecting the control parameters, thereby improving the fuel consumption rate, reducing the combustion noise, and improving the exhaust emission. Is possible.

  Specifically, the internal combustion engine has a plurality of cylinders, and fuel is injected from the fuel injection valve into the cylinders individually for each cylinder, and the index of the cooling loss is obtained for each cylinder. The internal combustion engine control parameters are corrected for each cylinder in accordance with the cooling loss index.

  As a result, even if the cooling loss due to the cooling water and the lubricating oil is different for each cylinder, an index of the cooling loss for each cylinder is obtained, and the internal combustion engine control parameters are corrected for each cylinder. It will be. As a result, it is possible to equalize the combustion state of the fuel in each cylinder, not only improve the fuel consumption rate, reduce the combustion noise, improve the exhaust emission, but also eliminate the variation between the cylinders, and the internal combustion engine Can also be suppressed.

  Specifically, the amount of fuel injected into the cylinder within the predetermined time is obtained as follows. That is, the fuel injection amount is obtained as an integrated injection amount obtained by multiplying the fuel amount at each fuel injection timing actually injected into the cylinder within a predetermined time by the weighting factor. The weight coefficient is set to a larger value as it is multiplied by the fuel injection amount injected at a timing close to the estimated timing of the cylinder inner surface temperature and the piston surface temperature.

  For this reason, the degree of influence of the actual fuel injection amount on the fuel injection amount (fuel injection amount toward the cylinder within a predetermined time; integrated injection amount) becomes smaller as it goes back in the past. For this reason, an integrated injection amount in which the degree of reflection of the current internal combustion engine operating state and the immediately preceding internal combustion engine operating state is increased is required. As a result, it is possible to estimate the cylinder inner surface temperature and the piston surface temperature while recognizing the residual heat amount in the cylinder with high accuracy, and it is possible to correct the control parameters suitable for the current operating state of the internal combustion engine. .

  Specifically, the internal combustion engine control parameters corrected in accordance with the cooling loss index include fuel injection timing, fuel injection amount, fuel injection pressure, exhaust gas recirculation amount for returning exhaust gas to the intake system, and intake air supercharging pressure. And at least one of the energization periods of the glow plug.

  For example, when the cooling loss due to cooling water or lubricating oil is large and the above cooling loss index is large, the fuel injection timing is advanced, the fuel injection amount is increased, and the fuel injection pressure Are corrected to the low pressure side, the exhaust gas recirculation amount reducing side, the intake supercharging pressure to the high pressure side, and the glow plug energizing period to the long side.

  Further, the following two methods can be cited as methods for correcting the internal combustion engine control parameters. First, based on the cooling loss index, the compression end temperature when the piston reaches the compression top dead center or the in-cylinder gas temperature when the piston reaches a predetermined position is estimated. The internal combustion engine control parameters are corrected according to the above.

  Further, based on the cooling loss index, the compression end pressure when the piston reaches the compression top dead center or the in-cylinder pressure when the piston reaches a predetermined position is estimated, and this estimated value is Accordingly, the internal combustion engine control parameters are corrected.

  In correcting the internal combustion engine control parameter, it is preferable to consider the state quantity of the intake gas introduced into the cylinder. Specifically, the internal combustion engine control parameter is corrected in accordance with the cooling loss index obtained from the cooling loss index, the intake gas amount, and the intake gas temperature.

  As a result, it becomes possible to correct the internal combustion engine control parameters according to the state quantity (intake gas quantity, intake gas temperature) of the intake gas introduced into the cylinder as well as the cooling loss due to cooling water or lubricating oil. The correction amount can be obtained more appropriately.

  In the present invention, an index of cooling loss in one cylinder is obtained based on the cylinder inner surface temperature and piston surface temperature, and the internal combustion engine control parameter is corrected in accordance with the index of cooling loss. For this reason, it is possible to obtain a correction amount that takes into consideration the cooling loss due to the cooling water and the cooling loss due to the lubricating oil, and it becomes possible to improve the fuel consumption rate, reduce the combustion noise, and improve the exhaust emission.

It is a figure which shows schematic structure of the engine which concerns on embodiment, and its control system. It is sectional drawing which shows the combustion chamber of a diesel engine, and its peripheral part. It is a block diagram which shows the structure of control systems, such as ECU. Change in heat generation rate (heat generation amount per unit rotation angle of crankshaft) and change in fuel injection rate (fuel injection amount per unit rotation angle of crankshaft) when ideal combustion is performed in the combustion stroke It is a wave form diagram which shows an example. It is a figure which shows an example of each change of the exhaust gas temperature in a combustion period of a fuel, a bore | bore inner wall surface temperature, a head lower surface temperature, and piston top surface temperature. It is a flowchart figure which shows the correction | amendment procedure of the control parameter by a cold loss index | exponent. 7A is an injection timing correction map, FIG. 7B is an injection amount correction map, FIG. 7C is an injection pressure correction map, FIG. 7D is an EGR rate correction map, and FIG. (E) is a supercharging pressure correction map, and FIG. 7 (f) is a diagram showing a glow energization period correction map. It is a figure which shows an example of the change of in-cylinder gas temperature accompanying the increase in the amount of fuel injection. It is a figure which shows transition of the fuel ignition timing of each cycle. FIG. 10A shows an example of the change in heat generation rate obtained in the embodiment, and FIG. 10B shows the heat generation rate in the prior art. It is a figure which shows an example of the change of incidence.

  Hereinafter, embodiments of the present invention will be described with reference to the drawings. In the present embodiment, a case where the present invention is applied to a common rail in-cylinder direct injection multi-cylinder (for example, in-line 4-cylinder) diesel engine (compression self-ignition internal combustion engine) mounted on an automobile will be described.

-Engine configuration-
First, a schematic configuration of a diesel engine (hereinafter simply referred to as an engine) according to the present embodiment will be described. FIG. 1 is a schematic configuration diagram of an engine 1 and its control system according to the present embodiment. FIG. 2 is a cross-sectional view showing the combustion chamber 3 of the diesel engine 1 and its periphery.

  As shown in FIG. 1, the engine 1 according to the present embodiment is configured as a diesel engine system having a fuel supply system 2, a combustion chamber 3, an intake system 6, an exhaust system 7 and the like as main parts.

  The fuel supply system 2 includes a supply pump 21, a common rail 22, an injector (fuel injection valve) 23, an engine fuel passage 27, and the like.

  The supply pump 21 pumps fuel from the fuel tank, makes the pumped fuel high pressure, and supplies it to the common rail 22 via the engine fuel passage 27. The common rail 22 has a function as a pressure accumulation chamber that holds (accumulates) high pressure fuel at a predetermined pressure, and distributes the accumulated fuel to the injectors 23, 23,. The injector 23 includes a piezoelectric element (piezo element) therein, and is configured by a piezo injector that is appropriately opened to supply fuel into the combustion chamber 3.

  The intake system 6 includes an intake manifold 63 connected to an intake port 15a formed in the cylinder head 15 (see FIG. 2), and an intake pipe 64 that constitutes an intake passage is connected to the intake manifold 63. In addition, an air cleaner 65, an air flow meter 43, and an intake throttle valve (diesel throttle) 62 are arranged in this intake passage in order from the upstream side. The air flow meter 43 outputs an electrical signal corresponding to the amount of air flowing into the intake passage via the air cleaner 65.

  The exhaust system 7 includes an exhaust manifold 72 connected to an exhaust port 71 formed in the cylinder head 15, and an exhaust pipe 73 constituting an exhaust passage is connected to the exhaust manifold 72. An exhaust purification device 77 is disposed in the exhaust passage. The exhaust gas purification device 77 includes a catalyst (NOx storage catalyst or oxidation catalyst) and a DPF (Diesel Particle Filter). Further, as the exhaust purification device 77, a DPNR catalyst (Diesel Particle-NOx Reduction catalyst) may be employed.

  Here, the structure of the combustion chamber 3 of the engine 1 and its peripheral part will be described with reference to FIG. As shown in FIG. 2, a cylinder block 11 constituting a part of the engine body is formed with a cylindrical cylinder bore 12 for each cylinder (four cylinders), and a piston 13 is formed inside each cylinder bore 12. Is accommodated so as to be slidable in the vertical direction.

  The combustion chamber 3 is formed above the top surface 13 a of the piston 13. That is, the combustion chamber 3 is defined by the lower surface 15 b of the cylinder head 15 attached to the upper part of the cylinder block 11, the inner wall surface 12 a of the cylinder bore 12, and the top surface 13 a of the piston 13. A cavity (concave portion) 13 b is formed in a substantially central portion of the top surface 13 a of the piston 13, and this cavity 13 b also constitutes a part of the combustion chamber 3.

  The piston 13 is connected to a crankshaft, which is an engine output shaft, by a connecting rod 18. As a result, the reciprocating movement of the piston 13 in the cylinder bore 12 is transmitted to the crankshaft via the connecting rod 18, and the engine output is obtained by rotating the crankshaft.

  The piston 13 is cooled by an oil jet of lubricating oil. Specifically, the piston 13 is cooled by supplying cooling oil to the inside of a cooling channel (oil gallery) 13d formed inside the piston 13 and the lower surface of the piston 13 from an oil jet nozzle (not shown). (See, for example, Japanese Patent Laid-Open No. 2000-303905). The entire piston 13 including the top surface 13a of the piston 13 is cooled by the lubricating oil supplied from the oil jet nozzle.

  Further, a glow plug 19 is disposed toward the combustion chamber 3. The glow plug 19 is red hot when an electric current is applied immediately before the engine 1 is started (execution of pre-glow), and a part of the fuel spray is blown onto the glow plug 19 as a start assist device that promotes ignition and combustion. Function. Further, the glow plug 19 is used as needed during the start of the engine 1 (during cranking) and after the engine is started (execution of afterglow). For example, power is supplied to ensure combustion stability when the in-cylinder temperature is low (for example, 750 K or less) or the cooling water temperature is low. Further, this energization period can be arbitrarily adjusted.

  The cylinder head 15 is formed with the intake port 15a and the exhaust port 71, respectively, and an intake valve 16 for opening and closing the intake port 15a and an exhaust valve 17 for opening and closing the exhaust port 71 are disposed. The cylinder head 15 is provided with the injector 23 that directly injects fuel into the combustion chamber 3. The injector 23 is disposed at a substantially upper center of the combustion chamber 3 in a standing posture along the cylinder center line P, and injects fuel introduced from the common rail 22 toward the combustion chamber 3 at a predetermined timing.

  Furthermore, as shown in FIG. 1, the engine 1 is provided with a supercharger (turbocharger) 5. The turbocharger 5 includes a turbine wheel 52 and a compressor wheel 53 that are connected via a turbine shaft 51. The compressor wheel 53 is disposed facing the intake pipe 64, and the turbine wheel 52 is disposed facing the exhaust pipe 73. For this reason, the turbocharger 5 performs a so-called supercharging operation in which the compressor wheel 53 is rotated using the exhaust flow (exhaust pressure) received by the turbine wheel 52 to increase the intake pressure. The turbocharger 5 in the present embodiment is a variable nozzle type turbocharger, and a variable nozzle vane mechanism (not shown) is provided on the turbine wheel 52 side. By adjusting the opening of the variable nozzle vane mechanism, the engine 1 supercharging pressure can be adjusted.

  An intake pipe 64 of the intake system 6 is provided with an intercooler 61 for forcibly cooling the intake air whose temperature has been raised by supercharging in the turbocharger 5.

  Further, the engine 1 is provided with an exhaust gas recirculation passage (EGR passage) 8 that connects the intake system 6 and the exhaust system 7. The EGR passage 8 is configured to reduce the combustion temperature by recirculating a part of the exhaust gas to the intake system 6 and supplying it again to the combustion chamber 3, thereby reducing the amount of NOx generated. In addition, the EGR passage 8 is opened and closed steplessly by electronic control, and the exhaust gas passing through the EGR passage 8 (recirculating) is cooled by an EGR valve 81 that can freely adjust the exhaust flow rate flowing through the passage. An EGR cooler 82 is provided. The EGR passage 8, the EGR valve 81, the EGR cooler 82, and the like constitute an EGR device (exhaust gas recirculation device).

-Sensors-
Various sensors are attached to each part of the engine 1, and signals related to environmental conditions of each part and the operating state of the engine 1 are output.

  For example, the air flow meter 43 outputs a detection signal corresponding to the flow rate (intake air amount) of intake air upstream of the intake throttle valve 62 in the intake system 6. The rail pressure sensor 41 outputs a detection signal corresponding to the fuel pressure stored in the common rail 22. The throttle opening sensor 42 detects the opening of the intake throttle valve 62. The intake pressure sensor 48 is disposed in the intake manifold 63 and outputs a detection signal corresponding to the intake air pressure. The intake air temperature sensor 49 is disposed in the intake manifold 63 and outputs a detection signal corresponding to the temperature of the intake air. The A / F (air-fuel ratio) sensor 44 outputs a detection signal that continuously changes in accordance with the oxygen concentration in the exhaust downstream of the exhaust purification device 77 of the exhaust system 7. Similarly, the exhaust temperature sensor 45 outputs a detection signal corresponding to the temperature of exhaust gas (exhaust temperature) downstream of the exhaust purification device 77 of the exhaust system 7.

-ECU-
The ECU 100 includes a microcomputer including a CPU, a ROM, a RAM, and the like (not shown) and an input / output circuit. As shown in FIG. 3, the input circuit of the ECU 100 includes the rail pressure sensor 41, the throttle opening sensor 42, the air flow meter 43, the A / F sensor 44, the exhaust temperature sensor 45, the intake pressure sensor 48, and the intake temperature sensor 49. Is connected. Further, a detection signal corresponding to the cooling water temperature of the engine 1 (the temperature of the cooling water flowing through the water jackets 11a and 15c (see FIG. 2) formed in the cylinder block 11 and the cylinder head 15) is input to the input circuit. A water temperature sensor 46 that outputs a signal, an accelerator opening sensor 47 that outputs a detection signal corresponding to the amount of depression of the accelerator pedal, and a detection signal (pulse) each time the output shaft (crankshaft) of the engine 1 rotates by a certain angle. A crank position sensor 40 and the like are connected.

  On the other hand, the supply circuit 21, the injector 23, the intake throttle valve 62, the EGR valve 81, and the variable nozzle vane mechanism (actuator for adjusting the opening degree of the nozzle vane) 54 of the turbocharger 5 are connected to the output circuit of the ECU 100. ing.

  Then, the ECU 100 executes various controls of the engine 1 based on outputs from the various sensors described above, calculated values obtained by arithmetic expressions using the output values, or various maps stored in the ROM. .

  For example, the ECU 100 executes pilot injection (sub-injection) and main injection (main injection) as fuel injection control of the injector 23.

  The pilot injection is an operation for injecting a small amount of fuel in advance prior to the main injection from the injector 23. The pilot injection is an injection operation for suppressing the ignition delay of fuel due to the main injection and leading to stable diffusion combustion, and is also referred to as sub-injection. Further, the pilot injection in the present embodiment has not only a function of suppressing the initial combustion speed by the main injection described above but also a preheating function of increasing the in-cylinder temperature. That is, after the pilot injection is performed, the fuel injection is temporarily interrupted, and the compressed gas temperature (in-cylinder temperature) is sufficiently increased until the main injection is started to reach the fuel self-ignition temperature (for example, 900K). In this way, the ignitability of the fuel injected by the main injection is ensured satisfactorily. Further, the basic injection amount and the basic injection timing of this pilot injection are set in advance as values according to the engine operating state through experiments or simulations. The final injection amount and the final injection timing are determined by correcting the basic injection amount and the basic injection timing of the pilot injection by a correction amount corresponding to a cooling loss index, which will be described later. Details will be described later.

  The main injection is an injection operation (torque generation fuel supply operation) for generating torque of the engine 1. The injection amount in the main injection is basically determined so as to obtain the required torque according to the operation state such as the engine speed, the accelerator operation amount, the coolant temperature, the intake air temperature, and the like. For example, the higher the engine speed (the engine speed calculated based on the detection value of the crank position sensor 40), the larger the accelerator operation amount (the accelerator pedal depression amount detected by the accelerator opening sensor 47). As the accelerator opening becomes larger, the required torque value of the engine 1 is higher, and accordingly, the fuel injection amount in the main injection is also set higher. The basic injection timing of the main injection is set in advance as a value corresponding to the engine operating state through experiments or simulations. It should be noted that the final injection timing is determined by correcting the basic injection timing of the main injection by a correction amount corresponding to a later-described cooling loss index. This will also be described in detail later.

  The fuel injected by the main injection starts combustion after a physical ignition delay period and a chemical ignition delay period. At this time, when the preheating in the cylinder is appropriately performed by the pilot injection, the premixed combustion of the fuel injected by the main injection is hardly performed, and most of the fuel is diffusion combustion. As a result, controlling the injection timing of the main injection is substantially equivalent to controlling the start timing of diffusion combustion as it is, and the controllability of combustion can be greatly improved.

  The physical ignition delay is the time required for evaporation and mixing of the fuel droplets, and depends on the gas temperature of the combustion field. On the other hand, chemical ignition delay is the time required for chemical bonding, decomposition, and oxidation heat generation of fuel vapor.

  As an example of a specific fuel injection mode, the pilot injection (fuel injection from a plurality of injection holes formed in the injector 23) is performed before the piston 13 reaches the compression top dead center, and the fuel injection is temporarily stopped. After that, the main injection is executed when the piston 13 reaches the vicinity of the compression top dead center after a predetermined interval. As a result, the fuel is combusted by self-ignition, and energy generated by this combustion is kinetic energy for pushing the piston 13 toward the bottom dead center (energy serving as engine output), and heat energy for raising the temperature in the combustion chamber 3. The heat energy is radiated to the outside (for example, cooling water) through the cylinder block 11 and the cylinder head 15.

  In the diesel engine 1, it is important to meet various requirements such as improvement of exhaust emission by reducing NOx generation amount and smoke generation amount, reduction of combustion noise during combustion stroke, and sufficient securing of engine torque. As a method for simultaneously satisfying these requirements, it is effective to appropriately control the change state of the heat generation rate in the cylinder during the combustion stroke (change state represented by the heat generation rate waveform).

  The waveform shown in the upper part of FIG. 4 is an example of an ideal heat generation rate waveform related to combustion of fuel injected by pilot injection and main injection, with the horizontal axis representing the crank angle and the vertical axis representing the heat generation rate. . TDC in the figure indicates the crank angle position corresponding to the compression top dead center of the piston 13. The waveform shown in the lower part of FIG. 4 shows the waveform of the injection rate of fuel injected from the injector 23 (fuel injection amount per unit rotation angle of the crankshaft).

  As the heat release rate waveform, for example, combustion of fuel injected by main injection is started from the vicinity of the compression top dead center (TDC) of the piston 13, and a predetermined piston position after the compression top dead center of the piston 13 (for example, The heat generation rate reaches a maximum value (peak value) at 10 degrees after compression top dead center (ATDC 10 °), and a predetermined piston position after compression top dead center (for example, 25 degrees after compression top dead center ( At the time of ATDC 25 °), the combustion of the fuel injected in the main injection ends. If combustion of the air-fuel mixture is performed in such a state where the heat generation rate changes, for example, 50% of the air-fuel mixture in the cylinder burns at 10 degrees after compression top dead center (ATDC 10 °). Completed status. That is, the combustion center of gravity is 10 degrees after compression top dead center (ATDC 10 °), and about 50% of the total heat generation amount in the combustion stroke is generated by ATDC 10 °, and the engine 1 is operated with high thermal efficiency. Is possible. Each of the above values is not limited to this, and since the ideal heat release rate waveform differs for each model of the engine 1, these values also differ for each model.

  In the situation where combustion with such an ideal heat generation rate waveform is performed, the pre-injection in the cylinder is sufficiently performed by pilot injection, and the fuel injected in the main injection by this pre-heating is By immediately being exposed to a temperature environment equal to or higher than the self-ignition temperature, combustion is started through the physical ignition delay period and the chemical ignition delay period.

  In addition to the pilot injection and main injection described above, after injection and post injection are performed as necessary. Since these injection functions are well known, description thereof is omitted here.

  Further, the ECU 100 controls the opening degree of the EGR valve 81 according to the operating state of the engine 1 to adjust the exhaust gas recirculation amount (EGR amount) toward the intake manifold 63. The basic operation of adjusting the EGR amount (adjusting the EGR rate) is performed according to an EGR map that is created in advance by experiments, simulations, etc. and stored in the ROM. This EGR map is a map for determining the EGR amount (EGR rate) using the engine speed (engine speed) and the engine load (engine load) as parameters. The EGR rate determined according to the EGR map in this way is also corrected by a correction amount according to a later-described cold loss index, and thereby the final EGR rate (target EGR rate) is determined.

  Further, the fuel injection pressure at the time of executing the fuel injection is determined by the internal pressure of the common rail 22. As the common rail internal pressure, in general, the target value of the fuel pressure supplied from the common rail 22 to the injector 23 becomes higher as the engine load increases and the engine speed increases. This rail pressure is set according to a fuel pressure setting map stored in the ROM, for example. In the present embodiment, the fuel pressure is adjusted between 30 MPa and 200 MPa according to the engine load and the like. The fuel injection pressure determined in this way is also corrected by a correction amount corresponding to a later-described cooling loss index, whereby the final fuel injection pressure (target fuel injection pressure) is determined.

  As a feature of the present embodiment, the ECU 100 individually considers a cooling loss due to cooling water and a cooling loss due to lubricating oil as cooling loss (cooling loss) for each cylinder, and based on the cooling loss for each cylinder. Various control parameters are corrected. Hereinafter, a calculation operation of a value serving as an index of cooling loss for each cylinder and a correction operation of various control parameters based on the index will be described.

-Concept of values that serve as indicators of cooling loss-
First, the concept of a value serving as an index of cooling loss for each cylinder will be described. This value is obtained from the history of the fuel injection amount from the injector 23, the temperature of the cylinder bore inner wall surface 12a, the temperature of the cylinder head lower surface 15b, the temperature of the piston top surface 13a, and the like. Hereinafter, the inner wall surface temperature of the cylinder bore 12 and the temperature of the cylinder lower surface are collectively referred to as “cylinder inner surface temperature”.

  Specifically, the cooling loss, which is a factor for lowering the in-cylinder temperature, includes cooling loss due to the cooling effect of the cooling water (cooling water flowing through the water jackets 11a and 15c) and lubricating oil (from the oil jet nozzle). Cooling loss due to the cooling effect of the supplied lubricating oil) can be mentioned.

  The cooling water to be cooled is the cylinder block 11 and the cylinder head 15. In the cylinder block 11, the cylinder bore inner wall surface 12 a faces the cylinder, and in the cylinder head 15, the cylinder head lower surface 15 b is the cylinder. It faces inside. That is, in the cylinder block 11 and the cylinder head 15 cooled by the cooling water, the temperature of the cylinder bore inner wall surface 12a and the cylinder head lower surface 15b changes due to the cooling loss due to the cooling water, and the combustion state in the cylinder is brought about. Will have an impact.

  On the other hand, what the lubricating oil is to be cooled is the piston 13, and the piston top surface 13 a faces the inside of the cylinder. That is, in the piston 13 cooled by the lubricating oil, the temperature of the piston top surface 13a changes due to the cooling loss due to the lubricating oil, and affects the combustion state in the cylinder.

  The cylinder inner surface temperature and the temperature of the piston top surface 13a (hereinafter referred to as “piston top surface temperature”) are determined by the amount of residual heat in the cylinder corresponding to the history of the engine operating state and the cooling loss. Further, the amount of residual heat in the cylinder corresponding to the history of the engine operating state is determined by the total amount of fuel injected within a predetermined period (the amount of fuel contributing to the temperature increase in the cylinder). Therefore, the cylinder inner surface temperature can be estimated based on the total fuel injection amount and the coolant temperature within a predetermined period. Further, the piston top surface temperature can be estimated based on the total fuel injection amount within a predetermined period and the lubricating oil temperature.

  FIG. 5 shows the exhaust gas temperature (solid line), bore inner wall surface temperature (dashed line), head lower surface temperature (one-dot chain line), piston top surface temperature (two-dot chain line) during the fuel combustion period (period Ta in the figure). An example of the change is shown. As described above, the bore inner wall surface temperature, the head lower surface temperature, and the piston top surface temperature rise with a slight delay with respect to the rise in the exhaust gas temperature accompanying the combustion of fuel, and the respective temperatures are the cooling water cooling. The values are different from each other depending on the cooling loss due to the effect and the cooling loss due to the cooling effect of the lubricating oil, and the influence of each on the combustion in the cylinder is different. That is, the cooling loss caused by the cooling water and the cooling loss caused by the lubricating oil have different effects on combustion. Therefore, in this embodiment, the cylinder inner surface temperature and the piston top surface temperature that change in accordance with these cooling losses are individually obtained (estimated).

  In this embodiment, from the thus determined (estimated) cylinder inner surface temperature and piston top surface temperature, a cold loss index reference value serving as an index corresponding to the cold loss in each cylinder is calculated. The control amounts (control parameters) of various control devices are adjusted based on the reference value of the cooling loss index and the state quantity of the intake gas.

-Adjustment of control amount-
Next, a flowchart of FIG. 6 shows a calculation procedure of a cold loss index reference value serving as an index corresponding to the cold loss and a correction operation of control amounts (control parameters) of various control devices using the cold loss index reference value. This will be described in detail. The flowchart shown in FIG. 6 is executed every predetermined time after an unillustrated ignition switch (start switch) is turned on and the engine 1 is started.

  First, in step ST1, the integrated injection amount within a predetermined time for each cylinder is calculated. Here, each fuel injection amount from the injector 23 in each cylinder is represented by Qv (i, j), and the time when the fuel injection is performed is represented by Zit (i, j). “I” is a cylinder number. In the case of a four-cylinder engine, “i” takes a value of “1 to 4”. Further, “j” is an injection cycle, and the injection cycle in the fuel injection executed immediately before is “1”, and the fuel injection executed once before the fuel injection (executed in the same cylinder) The injection cycle is set to “2”, and the injection cycle in the fuel injection that is executed a predetermined time before (most preceded in the same cylinder, for example, within a period of 5 min) is set to “j”. The target fuel injection is fuel injection (for example, pilot injection and main injection) that affects the amount of residual heat in the cylinder (becomes a heat source in the cylinder).

  Therefore, when fuel injection is performed in a certain cylinder, “j” of the fuel injection amount Qv (i, j) is incremented (incremented) one by one in that cylinder. That is, when fuel injection is executed in the i-th cylinder, the fuel injection amount in the fuel injection executed this time is Qv (i, 1), and the fuel injection amount executed last time in the same cylinder is Qv (i , 1) is changed from Qv (i, 2) to Qv (i, 2), and the fuel injection amount previously executed in the same cylinder is changed from Qv (i, 2) to Qv (i, 3). That is, the fuel injection amount is changed from Qv (i, j) to Qv (i, j + 1).

  Similarly, when fuel injection is performed in a certain cylinder, “j” of the fuel injection time Zit (i, j) in that cylinder is incremented (incremented) one by one. That is, when fuel injection is performed in the i-th cylinder, the fuel injection time is Zit (i, 1), and the fuel injection time previously executed in the same cylinder is Zit (i, 1) to Zit (i). , 2), and the fuel injection time executed two times before in the same cylinder is changed from Zit (i, 2) to Zit (i, 3). That is, the fuel injection time is changed from Zit (i, j) to Zit (i, j + 1).

  Then, the maximum value of “j” that satisfies the following expression (1) is obtained, and the period from Zit (i, 1) to Zit (i, j) is set as the current total fuel injection amount integration period.

Zit (i, 1) −Zit (i, j) ≦ ZitB (1)
“B” in the equation (1) is a value for setting the total fuel injection amount integration period, and is set to a value corresponding to, for example, 5 min. This value is not limited to this and is set as appropriate.

  A predetermined history weight coefficient K (i) is assigned to each of the injection amounts Qv (i, 1) to Qv (i, j) of each fuel injection of the i-th cylinder in the total fuel injection amount integration period set as described above. An integrated injection amount SQ (i) that is the sum obtained by multiplication is calculated by the following equation (2).

SQ (i) = Qv (i, 1) · K (1) + Qv (i, 2) · K (2) +... + Qv (i, j−1) · K (j−1) + Qv (i, j)・ K (j) (2)
The history weight coefficient K (i) is set to a larger value as “i” is smaller. That is, the cumulative injection amount SQ (i) is obtained by reducing the degree of influence of the fuel injection amount on the integrated injection amount SQ (i) as it goes back in the past, and increasing the degree of reflection of the current engine operating state and the immediately preceding engine operating state. It is supposed to be. The history weight coefficient K (i) is set in advance for each model of the engine 1 through experiments and simulations. For example, since this integrated injection amount SQ (i) is obtained as a value correlated with the amount of residual heat in the cylinder, in the case of an engine having a small heat capacity of the cylinder block 11 or the piston 13, Each history weight coefficient K (1) to K (i) is set as a small value compared to the case. As a result, an integrated injection amount SQ (i) corresponding to the amount of heat in a state where the amount of residual heat in the cylinder is small is obtained.

  The above operation is performed every time fuel injection is performed in any cylinder, and the integrated injection amount SQ (i) in that cylinder is calculated.

  In this way, after calculating the integrated injection amount SQ (i) in the cylinder in which the fuel injection has been performed, the routine proceeds to step ST2 where the cylinder inner surface temperature is estimated.

  In this cylinder inner surface temperature estimation operation, the cylinder inner surface temperature Twall () is calculated by the function f (Thw, SQ (i)) having the cooling water temperature Thw detected by the water temperature sensor 46 and the integrated injection amount SQ (i) as variables. i) is calculated. This function f (Thw, SQ (i)) is obtained in advance by experiments and simulations and stored in the ROM of the ECU 100. Further, a map for obtaining the cylinder inner surface temperature Twall (i) from the coolant temperature Thw and the integrated injection amount SQ (i) may be stored in the ROM of the ECU 100.

  Thereafter, in step ST3, the piston top surface temperature is estimated.

  In the operation of estimating the piston top surface temperature, first, the lubricating oil temperature Tho (n) is calculated. The lubricating oil temperature Th0 (n) is calculated from the cooling water temperature Thw detected by the water temperature sensor 46. Specifically, the lubricating oil temperature Th0 (n) rises with a delay (primary response delay) with respect to the rising speed of the cooling water temperature Thw, so the cooling water temperature Thw and the delay coefficient Th0 (n -1) as a variable.

  Then, the piston top surface temperature Tpis (i) is calculated by a function g (Tho (n), SQ (i)) having the calculated lubricating oil temperature Tho (n) and the integrated injection amount SQ (i) as variables. To do. This function g (Tho (n), SQ (i)) is also obtained in advance by experiments and simulations and stored in the ROM of the ECU 100. Further, a map for obtaining the piston top surface temperature Tpis (i) from the lubricating oil temperature Tho (n) and the integrated injection amount SQ (i) may be stored in the ROM of the ECU 100. When an oil temperature sensor capable of directly detecting the lubricating oil temperature Tho (n) is provided, the function g (Tho) with the lubricating oil temperature Tho (n) detected by the oil temperature sensor as a variable. (N), SQ (i)), the piston top surface temperature Tpis (i) is calculated. That is, it is not necessary to consider the rise delay (primary response delay) of the lubricating oil temperature Tho (n).

  After calculating the cylinder inner surface temperature Twall (i) and the piston top surface temperature Tpis (i) as described above, the process proceeds to step ST4 to calculate the cooling loss index reference value Lsc (i) for each cylinder. Specifically, a cooling loss index reference value Lsc (for each cylinder) by a function h (Twall (i), Tpis (i)) having the cylinder inner surface temperature Twall (i) and the piston top surface temperature Tpis (i) as variables. i) is calculated. This function h (Twall (i), Tpis (i)) is obtained in advance by experiments and simulations and stored in the ROM of the ECU 100. Further, a map for determining the cooling loss index reference value Lsc (i) from the cylinder inner surface temperature Twall (i) and the piston top surface temperature Tpis (i) may be stored in the ROM of the ECU 100.

  The cooling loss index reference value Lsc (i) is a value serving as an index corresponding to the cooling loss caused by the cooling water and the cooling loss caused by the lubricating oil. That is, the larger the cooling loss, the larger the value calculated as the cooling loss index reference value Lsc (i).

  Thereafter, the process proceeds to step ST5, where a cooling loss index Lscf (i) for each cylinder is calculated. Specifically, the cooling loss index Lscf for each cylinder is determined by a function k (Lsc (i), Gcyl, Thim) having the cooling loss index reference value Lsc (i), the intake gas amount Gcyl, and the intake gas temperature Thim as variables. i) is calculated. The intake gas amount Gcyl is determined according to the intake air amount detected by the air flow meter 43 and the exhaust gas recirculation amount (EGR gas amount) estimated from the opening degree of the EGR valve 81 and the like. Further, the intake gas temperature Thim is detected by the intake temperature sensor 49. This function k (Lsc (i), Gcyl, Thim) is also obtained in advance by experiments and simulations and stored in the ROM of the ECU 100. Further, a map for obtaining the cooling loss index Lscf (i) from the cooling loss index reference value Lsc (i), the intake gas amount Gcyl, and the intake gas temperature Thim may be stored in the ROM of the ECU 100. The cooling loss index Lscf (i) is calculated as a larger value as the cooling loss index reference value Lsc (i) is larger, the intake gas temperature Thim is lower, and the intake gas amount Gcyl is larger. become.

  After the cooling loss index Lscf (i) is calculated in this way, the process proceeds to step ST6, and each control parameter of the engine 1 is corrected according to the cooling loss index Lscf (i). Specific correction will be described below.

(Fuel injection timing)
As the correction of the fuel injection timing, for example, the larger the cooling loss index Lscf (i) (corresponding to the situation where the cooling loss is large and the temperature in the cylinder is likely to decrease), the pilot injection timing and the main injection timing are corrected to the advance side. To do. That is, the basic injection timing is corrected by a correction amount corresponding to the cooling loss index Lscf (i), and this is determined as the final injection timing. The correction value corresponding to the cooling loss index Lscf (i) is obtained from the injection timing correction map shown in FIG. This injection timing correction map is created in advance by experiments and simulations and stored in the ROM of the ECU 100. A correction value for correcting the injection timing to the advance side is obtained as the cooling loss index Lscf (i) increases. It has become.

(Fuel injection amount)
As the correction of the fuel injection amount, for example, the pilot injection amount is corrected to the increase side as the cooling loss index Lscf (i) is larger. That is, the basic injection amount is corrected by a correction amount corresponding to the cooling loss index Lscf (i), and this is determined as the final injection amount. The correction value corresponding to the cooling loss index Lscf (i) is obtained from the injection amount correction map shown in FIG. This injection amount correction map is also created in advance by experiments and simulations and stored in the ROM of the ECU 100, and a correction value that corrects the pilot injection amount to the increase side as the cooling loss index Lscf (i) increases is obtained. It has become.

  FIG. 8 shows an example of a change in the in-cylinder gas temperature (for example, the in-cylinder gas temperature at the start of main injection) when the pilot injection amount increase correction is performed. As is apparent from FIG. 8, even when the engine water temperature (cooling water temperature) is the same, when the injection amount increase correction is performed (the fuel injection amount is increased in the order of the one-dot chain line, the broken line, and the solid line). Increases the in-cylinder gas temperature, improves the ignitability in main injection, and stabilizes combustion.

(Fuel injection pressure)
As the correction of the fuel injection pressure, for example, the fuel injection pressure is corrected to the lower pressure side as the cooling loss index Lscf (i) is larger. That is, the fuel injection pressure set according to the engine load or the like is corrected by a correction amount according to the cooling loss index Lscf (i) and determined as the final fuel injection pressure (target fuel injection pressure). To do. Thus, the larger the cooling loss index Lscf (i) (for example, the lower the temperature of the cylinder bore inner wall surface 12a), the more the fuel injection pressure is corrected to the low pressure side, whereby the fuel injected from the injector 23 becomes the cylinder bore inner wall surface. It is difficult to reach 12a. This is because when the temperature of the cylinder bore inner wall surface 12a is low, the fuel adhering to the cylinder bore inner wall surface 12a is difficult to evaporate, and when the amount increases, the dilution of oil (lubricating oil) by the fuel proceeds. In order to prevent this, the fuel injection pressure is corrected to the low pressure side.

  The fuel injection pressure correction value corresponding to the cooling loss index Lscf (i) is obtained from the injection pressure correction map shown in FIG. This injection pressure correction map is also created in advance by experiments and simulations and stored in the ROM of the ECU 100, and a correction value for correcting the fuel injection pressure to the low pressure side as the cooling loss index Lscf (i) increases is obtained. It has become.

(EGR rate)
As the correction of the EGR rate, for example, the larger the cooling loss index Lscf (i), the lower the target EGR rate (the ratio of the exhaust gas recirculation amount in the intake gas). That is, the EGR rate determined according to the EGR map is corrected by a correction amount corresponding to the cooling loss index Lscf (i), and is determined as the final EGR rate (target EGR rate). Thus, the larger the cooling loss index Lscf (i), the higher the target EGR rate is corrected to the lower side, so that the combustion temperature in the combustion field is maintained higher and the combustion efficiency can be improved.

  The correction value of the target EGR rate corresponding to the cooling loss index Lscf (i) is obtained from the EGR rate correction map shown in FIG. This EGR rate correction map is also created in advance by experiments and simulations and stored in the ROM of the ECU 100, and a correction value that corrects the EGR rate to the lower side as the cooling loss index Lscf (i) increases is obtained. It has become.

  Although the target EGR rate is corrected according to the cooling loss index Lscf (i) here, the target oxygen concentration in the intake system and the EGR control amount (the control amount of the EGR valve 81) are corrected. May be. That is, as the cooling loss index Lscf (i) is larger, the target oxygen concentration may be corrected to a higher side, or the opening degree of the EGR valve 81 may be corrected to a smaller side.

(Supercharging pressure)
As the correction of the supercharging pressure in the turbocharger 5, for example, the larger the cooling loss index Lscf (i), the higher the target supercharging pressure is corrected. That is, the basic boost pressure set according to the engine speed, the accelerator opening, etc. is corrected by a correction amount according to the cooling loss index Lscf (i), and this is corrected to the final boost pressure (target (Supercharging pressure) is determined. As described above, the larger the cooling loss index Lscf (i) is, the higher the target supercharging pressure is corrected, so that the combustion efficiency can be improved.

  The correction value of the target boost pressure corresponding to the cooling loss index Lscf (i) is obtained from the boost pressure correction map shown in FIG. This supercharging pressure correction map is also created in advance by experiments and simulations and stored in the ROM of the ECU 100. A correction value for correcting the target supercharging pressure to be higher as the cooling loss index Lscf (i) is larger. It has been obtained.

  Here, the target supercharging pressure is corrected according to the cooling loss index Lscf (i), but the turbo control amount (the vane opening degree of the variable nozzle vane mechanism 54) may be corrected. That is, the larger the cooling loss index Lscf (i), the smaller the vane opening degree of the variable nozzle vane mechanism 54 is corrected.

(Glow energization period)
As the correction of the energization period (afterglow execution period) after the engine start in the glow plug 19, for example, the larger the cooling loss index Lscf (i), the longer the energization period is corrected. That is, when the afterglow is executed when the cooling water temperature is equal to or lower than the predetermined temperature, the energization period is corrected by a correction amount corresponding to the cooling loss index Lscf (i), and this is determined as the final energization period. As described above, the larger the cooling loss index Lscf (i) is, the longer the energization period is corrected, so that the combustion can be stabilized.

  The correction value of the energization period according to the cooling loss index Lscf (i) is obtained from the glow energization period correction map shown in FIG. This glow energization period correction map is also created in advance by experiments and simulations and stored in the ROM of the ECU 100, and a correction value for correcting the energization period to be longer as the cooling loss index Lscf (i) is larger is obtained. It has become a thing.

  As described above, by correcting the fuel injection timing, the fuel injection amount, the fuel injection pressure, the EGR rate, the supercharging pressure, and the glow energization period in accordance with the cooling loss index Lscf (i), It is possible to optimize various control parameters according to the amount of residual heat in the cylinder, the cooling loss due to the cooling water (cooling loss of the cylinder), and the cooling loss due to the lubricating oil (cooling loss of the piston). As a result, an engine operating state suitable for the current demand (for example, demanded torque) can be obtained without overcorrecting the control parameters, improving the fuel consumption rate, reducing the combustion noise, and improving the exhaust emission. Can be achieved.

  Further, according to the present embodiment, even when the cooling loss due to the cooling water and the lubricating oil is different for each cylinder, the cold loss index reference value and the cooling loss index are obtained for each cylinder, Control parameters are corrected for each cylinder. For this reason, the combustion state of the fuel in each cylinder can be equalized, and the vibration of the engine 1 can be suppressed along with the elimination of the variation between the cylinders.

  FIG. 9 is a diagram showing the transition of the fuel ignition timing of each cycle from the start of operation of the engine 1 or after the engine load has changed. Various control parameters are set according to the prior art (only the cooling water temperature is changed). The transition of the fuel ignition timing in this embodiment is indicated by a broken line, and the transition of the fuel ignition timing in the present embodiment is indicated by a solid line. These waveforms are obtained by experiments or simulations.

  As is apparent from FIG. 9, in the prior art, a period until the fuel ignition timing converges (optimum fuel ignition timing (fuel ignition timing that can improve the fuel consumption rate and exhaust emission) is obtained. The period until it is generated) is relatively long (see period Tb in the figure). For this reason, in the period until the optimum fuel ignition timing is obtained, there is a concern that the fuel consumption rate will deteriorate, the combustion noise will increase, and the exhaust emission will deteriorate.

  On the other hand, in this embodiment, it converges rapidly at the optimal fuel ignition timing. For this reason, it is possible to obtain at an early stage an engine operating state capable of improving the fuel consumption rate, reducing the combustion noise, improving the exhaust emission, and the like.

  FIG. 10 shows the change in the heat generation rate in the cylinder during the combustion stroke, and FIG. 10 (a) shows an example of the change in the heat generation rate obtained in this embodiment. In FIG. 10A, the two-dot chain line indicates the change in the heat generation rate at the start of the operation of the engine 1, and the solid line indicates the change in the heat generation rate at the time after several cycles after the start of the operation of the engine 1. . On the other hand, FIG.10 (b) has shown an example of the change of the heat release rate in a prior art (what correct | amends various control parameters only according to a cooling water temperature). In FIG. 10B, the two-dot chain line indicates the change in the heat generation rate at the start of the operation of the engine 1, and the one-dot chain line indicates the change in the heat generation rate at the time after several cycles after the operation of the engine 1. Yes. Further, the broken line indicates a change in the heat generation rate at a point after several cycles, and the solid line indicates a change in the heat generation rate at a point after further several cycles.

  As described above, the conventional technique requires a long period until the heat generation rate converges and an ideal heat generation rate waveform is obtained. For this reason, the period during which the fuel consumption rate deteriorates, the combustion noise increases, and the exhaust emission deteriorates is long.

  On the other hand, in this embodiment, it is possible to obtain an ideal heat generation rate waveform at an early stage. For this reason, it is possible to obtain at an early stage an engine operating state capable of improving the fuel consumption rate, reducing combustion noise, improving exhaust emission, and the like.

  In the present embodiment, all of the fuel injection timing, the fuel injection amount, the fuel injection pressure, the EGR rate, the supercharging pressure, and the glow energization period are corrected according to the cooling loss index Lscf (i). It was. Not limited to this, one or more of fuel injection timing, fuel injection amount, fuel injection pressure, EGR rate, supercharging pressure, glow energization period, or the like is corrected according to the cooling loss index Lscf (i) You may make it do. Further, control parameters other than those described above may be corrected according to the cooling loss index Lscf (i).

(Modification 1)
Next, Modification 1 will be described. In this modification, the control parameter correction operation according to the cooling loss index Lscf (i) is different from that of the above-described embodiment. Therefore, only the control parameter correction operation will be described here.

  Specifically, based on the cooling loss index Lscf (i), the compression end temperature (the in-cylinder temperature when the piston 13 reaches the compression top dead center) or the in-cylinder temperature at a predetermined crank angle (the piston 13 The in-cylinder temperature at the time of reaching a predetermined position (for example, the crank angle at a position 30 ° before compression top dead center) is estimated, and the fuel injection timing, fuel injection amount, fuel injection pressure, EGR rate are estimated from this estimated value. At least one of the supercharging pressure and the glow energization period is corrected. In this case, as the in-cylinder temperature is lower, the same control (for example, advance correction of the fuel injection timing, increase correction of the fuel injection amount, etc.) as in the case where the cooling loss index Lscf (i) is larger in the above embodiment is performed. It will be.

  As a method for estimating the compression end temperature or the in-cylinder temperature at a predetermined crank angle from the cooling loss index Lscf (i), the cooling loss index Lscf (i) and the compression end temperature are calculated in advance by experiments and simulations. The relationship or the relationship between the cooling loss index Lscf (i) and the in-cylinder temperature at a predetermined crank angle is obtained, mapped, and stored in the ROM. Then, the compression end temperature or the in-cylinder temperature at a predetermined crank angle is obtained by fitting the cooling loss index Lscf (i) to this map.

(Modification 2)
Next, Modification 2 will be described. In the present modification, the control parameter correction operation according to the cooling loss index Lscf (i) is different from that in the above-described embodiment and modification 1. Therefore, only the control parameter correction operation will be described here.

  Specifically, based on the cooling loss index Lscf (i), the compression end pressure (cylinder pressure when the piston 13 reaches compression top dead center) or the cylinder pressure (piston 13 The in-cylinder pressure at the time of reaching a predetermined position (for example, the crank angle at a position 30 ° before compression top dead center) is estimated, and the fuel injection timing, fuel injection amount, fuel injection pressure, EGR rate are estimated from this estimated value. At least one of the supercharging pressure and the glow energization period is corrected. In this case, as the in-cylinder pressure is lower, the same control as in the case where the cooling loss index Lscf (i) is larger in the above embodiment is performed.

  As a method for estimating the compression end pressure or the in-cylinder pressure at a predetermined crank angle from the cooling loss index Lscf (i), the cooling loss index Lscf (i) and the compression end pressure can be estimated in advance through experiments and simulations. The relationship or the relationship between the cooling loss index Lscf (i) and the in-cylinder pressure at a predetermined crank angle is obtained, and this is mapped and stored in the ROM. Then, the compression end pressure or the in-cylinder pressure at a predetermined crank angle is obtained by fitting the cooling loss index Lscf (i) to this map.

-Other embodiments-
In the above-described embodiments and modifications, the case where the present invention is applied to an in-line four-cylinder diesel engine has been described. However, the present invention is not limited to the number of cylinders and the engine type (in-line engine, V-type engine, horizontally opposed engine). Etc.) is not particularly limited. Further, the present invention is applicable not only to diesel engines but also to gasoline engines.

  Further, in the above-described embodiment and each modified example, the “cylinder inner surface temperature” is the temperature of both the cylinder bore inner wall surface 12a and the cylinder head lower surface 15b. The present invention is not limited to this, and the function f (Thw, SQ (i)) is set so that only one temperature (for example, the temperature of the cylinder bore inner wall surface 12a) is treated as the cylinder inner surface temperature, and the cooling loss index reference value Lsc. (I) may be calculated.

  Moreover, although the said embodiment and each modification demonstrated the engine 1 which applied the piezo injector 23 which changes a fuel-injection rate by becoming a valve opening state of full open only during an electricity supply period, this invention is variable injection rate. Application to an engine to which an injector is applied is also possible.

  The present invention can be applied to control parameter correction control in a common rail in-cylinder direct injection multi-cylinder diesel engine mounted on an automobile.

1 engine (internal combustion engine)
11a Water jacket 12 Cylinder bore 12a Cylinder bore inner wall surface 13 Piston 13a Piston top surface 15 Cylinder head 15b Cylinder head lower surface 15c Water jacket 23 Injector (fuel injection valve)
3 Combustion chamber 46 Water temperature sensor 48 Intake pressure sensor 49 Intake temperature sensor 100 ECU
Thw Cooling water temperature Th0 (n) Lubricating oil temperature SQ (i) Cumulative injection amount Twall (i) Cylinder inner surface temperature Tpis (i) Piston top surface temperature Lsc (i) Cooling loss index reference value Lscf (i) Cooling loss index

Claims (7)

In a control device for an internal combustion engine that generates output by combustion of fuel injected from a fuel injection valve into a cylinder,
Estimated from the cylinder inner surface temperature estimated from the fuel injection amount toward the cylinder and the coolant temperature within a predetermined time, and the fuel injection amount and lubricant temperature toward the cylinder within the predetermined time A control apparatus for an internal combustion engine, characterized in that an index of cooling loss in one cylinder is obtained based on a piston surface temperature, and an internal combustion engine control parameter is corrected according to the index of cooling loss. The control apparatus for an internal combustion engine according to claim 1,
The internal combustion engine includes a plurality of cylinders, and fuel is individually injected from the fuel injection valve into the cylinders for each cylinder.
A control device for an internal combustion engine, characterized in that the index of cooling loss is obtained for each cylinder, and the internal combustion engine control parameter is corrected for each cylinder in accordance with the index of cooling loss. The control device for an internal combustion engine according to claim 1 or 2,
The amount of fuel injected into the cylinder within the predetermined time is the sum obtained by multiplying the fuel amount at each fuel injection timing actually injected into the cylinder within the predetermined time by a weighting factor. Calculated as the injection amount,
The control apparatus for an internal combustion engine, wherein the weighting coefficient is set to a larger value as it is multiplied by a fuel injection amount injected at a timing closer to the estimated timing of the cylinder inner surface temperature and the piston surface temperature. The control device for an internal combustion engine according to claim 1, 2, or 3,
The internal combustion engine control parameters corrected in accordance with the cooling loss index are: fuel injection timing, fuel injection amount, fuel injection pressure, exhaust gas recirculation amount for returning exhaust gas to the intake system, intake air supercharging pressure, glow plug A control device for an internal combustion engine, characterized by being at least one of energization periods. In the control device for an internal combustion engine according to any one of claims 1 to 4,
Based on the cooling loss index, the compression end temperature when the piston reaches compression top dead center or the in-cylinder gas temperature when the piston reaches a predetermined position is estimated, and the estimated value is An internal combustion engine control apparatus characterized in that the internal combustion engine control parameter is corrected. In the control device for an internal combustion engine according to any one of claims 1 to 4,
Based on the cooling loss index, the compression end pressure when the piston reaches compression top dead center or the in-cylinder pressure when the piston reaches a predetermined position is estimated. A control device for an internal combustion engine, wherein the control device corrects the internal combustion engine control parameter. The control apparatus for an internal combustion engine according to any one of claims 1 to 6,
The control apparatus for an internal combustion engine, wherein the internal combustion engine control parameter is configured to be corrected according to a cooling loss index obtained from the cooling loss index, intake gas amount, and intake gas temperature.

Images