JP2013064387A - Spark ignition direct injection engine and method for controlling the same - Google Patents

Abstract

PROBLEM TO BE SOLVED: To avoid the generation of torque shock and the deterioration of NVH performance in a spark ignition direct injection engine quickly changing air excessive ratio λ according to engine load.SOLUTION: A controller 100 executes transition control delaying a starting time of fuel injection more than a specific injection period from a final period of a compressing step to a compressing upper dead center such that a main combustion period when the fuel mass ratio of fuel becomes ≥10% and ≤90% is present in an delay angle side more than a specific crank angle time point when intracylinder pressure rising ratio when motoring becomes the negative maximum value while making air excessive ratio when combusting not more than 1, during transition time when an operating state of an engine body (engine 1) changes between a low load area where the air excessive ratio≥λ and a high load area where the air excessive ratio≤1.

Description

  The technology disclosed herein relates to a spark ignition direct injection engine and a control method thereof.

  For example, Patent Document 1 describes an engine in which the compression ratio is set to a high compression ratio of about 16 and the air-fuel mixture is made lean to increase the theoretical thermal efficiency of a spark ignition gasoline engine.

  Further, for example, Patent Document 2 describes a technique in which a surface that defines a combustion chamber is formed of a heat insulating material including a large number of bubbles from the viewpoint of reducing cooling loss and improving thermal efficiency.

Japanese Patent Laid-Open No. 9-217627 JP 2009-243355 A

  By the way, in the Otto cycle, which is the theoretical cycle of a spark ignition engine, the theoretical thermal efficiency increases as the compression ratio increases and the specific heat ratio of the gas increases. For this reason, the combination of the high compression ratio and the lean air-fuel mixture as described in Patent Document 1 is advantageous in improving the thermal efficiency (illustrated thermal efficiency).

  Therefore, when the lean air-fuel mixture is studied, since the fuel injection amount is relatively small when the engine operating state is in the low load region, the excess air ratio λ is set to, for example, a lean air-fuel mixture that is 2 or more. In addition to the improvement in thermal efficiency, it is possible to suppress the production of RawNOx. On the other hand, when the engine operating state is in the high load region, the fuel injection amount increases, so that it is difficult to maintain the excess air ratio λ at 2 or more. When the excess air ratio λ is set lower than 2, the excess air ratio λ has to be set to 1 or less because of the use of a three-way catalyst. Therefore, it is conceivable to switch the excess air ratio λ according to the engine load, such that the excess air ratio λ ≧ 2 is set in the low load region and the excess air ratio λ ≦ 1 is set in the high load region.

  However, when shifting from a low load area where the excess air ratio λ ≧ 2 to a high load area where the excess air ratio λ ≦ 1, for example, due to an acceleration request accompanying an accelerator operation, Since the intake air amount reduction control is not in time, the excess air ratio λ must be rapidly reduced by the fuel injection amount increase control. This leads to a sudden change in torque, resulting in the occurrence of torque shock and the deterioration of NVH performance. In particular, in a high compression ratio engine in which the geometric compression ratio of the engine is set high, there is a torque difference between a low load area where the excess air ratio λ ≧ 2 and a high load area where the excess air ratio λ ≦ 1. It tends to be large, and there is a risk that torque shock and NVH performance will be significantly deteriorated.

  The technology disclosed herein has been made in view of such a point, and the object of the technology is to generate torque shocks in a spark ignition direct injection engine that suddenly changes the excess air ratio λ according to the engine load and The purpose is to avoid the deterioration of the NVH performance.

  At the time of transition between the low load area where the excess air ratio λ ≧ 2 and the high load area where the excess air ratio λ ≦ 1, the inventors of the present application delay the combustion period while setting the excess air ratio λ ≦ 1. It was decided to interpose a transient control to reduce the combustion efficiency by setting it to the corner side.

  Specifically, the spark ignition direct injection engine disclosed herein is configured to inject fuel into the cylinder of the engine body having a cylinder whose geometric compression ratio is set to 18 or more, and the engine body. A fuel injection valve, ignition means configured to ignite an air-fuel mixture in the cylinder, fuel injection into the cylinder through the fuel injection valve in accordance with an operating state of the engine body, and And a controller configured to control ignition by the ignition means.

  When the operating state of the engine body is in a low load region lower than a predetermined load, the controller sets the excess air ratio λ during combustion to 2 or more, and the operating state of the engine body is the predetermined state. When the engine is in a high load region above the load, the excess air ratio λ during combustion is set to 1 or less, and the engine burns within a main combustion period in which the combustion mass ratio of the fuel is 10% or more and 90% or less. The fuel injection timing by the fuel injection valve is set to a specific injection timing near the compression top dead center so that the specific crank angle point at which the rate of increase in the cylinder pressure during the motoring of the main body becomes a negative maximum value is included. The controller also sets the excess air ratio λ during combustion to 1 or less during a transition in which the operating state of the engine body transitions between the low load region and the high load region, As serial main combustion period becomes more retarded than the specific crank angle point, it executes a transient control to delay the timing of the fuel injection than the specific injection timing.

  According to the above configuration, the engine body is a high compression ratio engine having a geometric compression ratio of 18 or more, and an excess air ratio λ ≧ 2 in a low load region lower than a predetermined load. The thermal efficiency shown in the figure is increased by the combination with the leaning of the ki. This is advantageous for improving fuel efficiency and also for improving emission performance. The upper limit value of the geometric compression ratio may be set to 40, for example.

  Further, in a high load region over a predetermined load, it becomes difficult to maintain the excess air ratio λ ≧ 2 as the fuel injection amount increases, so the excess air ratio λ ≦ 1 is switched. This makes it possible to use a three-way catalyst and avoids deterioration of emission performance.

  Here, since the engine body is a high compression ratio engine, the maximum value and the minimum value (that is, the negative maximum value) of the cylinder pressure increase rate during motoring are relatively large values. The cylinder pressure increase rate becomes zero at the compression top dead center, and the maximum value occurs before the compression top dead center and the minimum value occurs after the compression top dead center. For this reason, if combustion is started before compression top dead center, combustion will occur near the crank angle where the rate of increase in cylinder pressure is large, so the maximum value of cylinder pressure increase during combustion will be very large. As a result, combustion noise increases.

  Therefore, in this engine, in the high load region where the excess air ratio λ ≦ 1, the fuel injection timing by the fuel injection valve is set to a specific injection timing near the compression top dead center, and thereby the fuel combustion mass The main combustion period in which the ratio is 10% or more and 90% or less includes a specific crank angle point at which the cylinder pressure increase rate during motoring becomes a negative maximum value. By doing this, in the high compression ratio engine, the maximum value of the in-cylinder pressure increase rate at the time of combustion is reduced, and the combustion noise is reduced.

  Further, in this engine, more accurately, during the transitional transition between the low load region where the excess air ratio λ ≧ 2 and the high load region where the excess air ratio λ ≦ 1, When transitioning to a region, immediately after transitioning to a high load region with an excess air ratio λ ≦ 1, and when transitioning from a high load region to a low load region, to a low load region with an excess air ratio λ ≧ 2. Immediately before the transition, the main combustion period is set to a specific crank by delaying the fuel injection timing from the specific injection timing while avoiding deterioration of the emission performance by setting the excess air ratio λ during combustion to 1 or less. Transient control is performed so that the angle is retarded from the angle point. Since the combustion efficiency is lowered by setting the main combustion period to the substantially retarded angle side, the generated torque is reduced while the excess air ratio λ ≦ 1.

  Therefore, when shifting from the low load region to the high load region, for example, when acceleration is required due to the accelerator operation, immediately after entering the high load region, the timing of fuel injection is temporarily set to the retarded side. Set the main combustion period to be greatly retarded to reduce the generated torque, reduce the torque difference from the torque in the low load region where the excess air ratio λ ≧ 2, or reduce the torque difference lose. After that, if the fuel injection timing, which has been greatly retarded, is gradually advanced toward the specific injection timing, the main combustion period will advance toward the specific crank angle described above. Since the combustion efficiency increases, the torque gradually increases. Thus, finally, the specific crank angle point is included in the main combustion period, and the transient control is finished. As a result, the transition from the low load region to the high load region is completed while avoiding torque shock and NVH performance deterioration.

  In the transient control during the transition from the high load region to the low load region, on the contrary, immediately before the transition to the low load region, the fuel injection timing is specified while maintaining the excess air ratio λ ≦ 1. What is necessary is just to delay gradually so that it may leave | separate from injection timing. As a result, the torque gradually decreases, and the excess air ratio λ is reduced in a state where the torque difference with the torque in the low load region where the excess air ratio λ ≧ 2 is reduced or eliminated. It becomes possible to switch to ≧ 2.

  During the transient control, the controller causes the fuel injection valve to pre-injection that starts within a period from the end of the compression stroke to the compression top dead center, and after the pre-injection and later than the specific injection timing. Main injection may be executed.

  In the transient control, as described above, the main combustion period is set to a significantly retarded side in the expansion stroke after the compression top dead center, so that the combustion stability is lowered. Further, since this engine is a high compression ratio engine and the compression end temperature and the compression end pressure become relatively high, it is possible to perform compression ignition without forcibly igniting the air-fuel mixture in the cylinder. When performing compression ignition, there is a problem that when the main combustion period is set to the retard side, the ignitability is lowered.

  Therefore, at the time of transient control, two times, a pre-injection that starts fuel injection within a relatively early period, specifically within a period from the end of the compression stroke to the compression top dead center, and the main injection after that pre-injection. Perform fuel injection. Since the pre-injection contributes to the temperature rise in the cylinder, it improves the combustion stability of the fuel injected by the main injection and improves the ignitability when performing compression ignition of the fuel injected by the main injection. To do.

  The controller may cause the ignition means to perform ignition between the pre-injection and the main injection.

  That is, after injecting fuel into the cylinder by pre-injection, ignition is performed by the ignition means to impart energy to the fuel, so that the fuel is heated and ignited, and the temperature and pressure in the cylinder increase. As a result, the temperature and pressure in the cylinder remain high even after the compression top dead center, and the fuel injected by the main injection that is significantly retarded in the expansion stroke is stably self-ignited. In this way, the ignitability and combustion stability of the fuel are improved during transient control in which the main combustion period is greatly retarded in the expansion stroke.

  The controller performs the transient control at the time of transition in which the operation state of the engine body shifts from the low load region to the fully open load region in the high load region, whereas the operation state of the engine body is the During transition from the low load region to the high load region where the load is lower than the fully open load, the intake air amount reduction control may be further performed during the execution of the transient control.

  That is, in the low load region where the excess air ratio λ ≧ 2, the throttle valve may be set to fully open, and the intake air amount does not need to be reduced. On the other hand, in the high load region where the excess air ratio λ ≦ 1, except for the fully open load region, it is necessary to reduce the intake air amount by adjusting the opening of the throttle valve in accordance with the fuel injection amount corresponding to the load. In the fully open load range, the throttle valve is set to fully open.

  Therefore, when the operating state of the engine body shifts from the low load region to the fully open load region in the high load region, the throttle valve is set to fully open in the fully open load region as described above. It is not necessary, and as described above, at least transient control for controlling the fuel injection timing may be performed. In contrast, when the operating state of the engine body shifts from the low load region to the high load region where the load is lower than the fully open load, as described above, the intake air amount must be reduced after the transition is completed. It is preferable to execute the intake air amount reduction control during the execution of the above. This makes it possible to reduce the fuel injection amount as much as the intake air amount is reduced during the transient control in which the excess air ratio λ ≦ 1, and this is advantageous in improving fuel consumption.

  The technology disclosed herein also provides a control method for a spark ignition direct injection engine having a cylinder having a geometric compression ratio set to 18 or more and configured to inject fuel directly into the cylinder. Related.

  In this control method, when the engine operating state is in a low load region lower than a predetermined load, the engine is operated such that the excess air ratio λ during combustion is 2 or more, and the engine is operated. When the state is in a high load region that is equal to or higher than the predetermined load, the excess air ratio λ at the time of combustion is set to 1 or less, and the timing of fuel injection into the cylinder is set to a specific injection near the compression top dead center. By setting the timing, the specific crank angle at which the rate of increase in the cylinder pressure during motoring of the engine becomes a negative maximum value within the main combustion period in which the combustion mass ratio of the fuel becomes 10% or more and 90% or less When the engine is operated in such a way that the time is included, and the engine operating state transitions between the low load region and the high load region, the excess air ratio λ during combustion is 1 or less. When And executing the transient control for delaying the fuel injection timing from the specific injection timing, so that the main combustion period is retarded from the specific crank angle time point. To drive.

  At the time of the transient control, as fuel injection into the cylinder, pre-injection that starts within a period from the end of the compression stroke to compression top dead center, and main injection that is delayed after the pre-injection and before the specific injection timing May be performed, and the air-fuel mixture in the cylinder may be ignited between the pre-injection and the main injection.

  Further, the transient control is executed during a transition in which the operating state of the engine shifts from the low load region to the full open load region in the high load region, and the engine operating state is changed from the low load region to the full open load. At the time of transition to the high load region where the load is low, intake air amount reduction control may be further performed during the execution of the transient control.

  As described above, the spark ignition direct injection engine and the control method thereof have a transition transient between a low load region where the excess air ratio λ ≧ 2 and a high load region where the excess air ratio λ ≦ 1. Sometimes the main combustion period has a negative cylinder pressure increase rate by significantly retarding the fuel injection timing while making the excess air ratio λ during combustion less than 1 to avoid deterioration of the emission performance. Transient control is executed on the retard side from the specific crank angle point at which the maximum value is reached. By this transient control, it becomes possible to complete the transition between the load regions accompanied by a sudden change in the excess air ratio λ while avoiding torque shock and NVH performance deterioration.

It is a figure showing roughly composition of a spark ignition type direct injection engine. It is a figure which illustrates the driving | operation map of an engine. It is a figure which illustrates the change of the cylinder pressure which concerns on the combustion in the low load area | region of excess air ratio (lambda)> = 2. It is a figure which shows the change of the cylinder pressure increase rate at the time of motoring when the geometric compression ratios of an engine are 20, 30 and 40. It is a figure which illustrates the change of the cylinder pressure which concerns on the combustion in the high load area | region of excess air ratio (lambda) <= 1. It is a figure which illustrates the change of the cylinder pressure which concerns on the combustion at the time of transition transition between a low load area | region and a high load area | region. It is a flowchart which concerns on the transient control in the case of transfer from a low load area | region to a high load area | region.

  Hereinafter, an embodiment of a spark ignition direct injection engine (hereinafter also simply referred to as an engine) will be described with reference to the drawings. The following description of the preferred embodiment is merely exemplary in nature. As shown in FIG. 1, the engine system includes an engine 1, various actuators associated with the engine 1, various sensors, and an engine controller 100 that controls the actuators based on signals from the sensors.

  The engine 1 is a spark ignition internal combustion engine, and has only a plurality of cylinders 11 although only one is shown in the figure. The engine 1 is mounted on a vehicle such as an automobile, and its output shaft is connected to drive wheels via a transmission, although not shown. The vehicle is propelled by the output of the engine 1 being transmitted to the drive wheels. The engine 1 includes a cylinder block 12 and a cylinder head 13 mounted thereon, and a cylinder 11 is formed inside the cylinder block 12.

  The piston 15 is slidably inserted into each cylinder 11, and defines a combustion chamber 17 together with the cylinder 11 and the cylinder head 13. In this embodiment, a recess 15 a is formed on the crown surface of the piston 15.

  Although only one is shown in FIG. 1, two intake ports 18 are formed in the cylinder head 13 for each cylinder 11, and each opens to the lower surface of the cylinder head 13 (the ceiling surface that defines the upper surface of the combustion chamber 17). And communicates with the combustion chamber 17. Similarly, two exhaust ports 19 are formed in the cylinder head 13 for each cylinder 11, and each communicates with the combustion chamber 17 by opening in the ceiling surface of the cylinder head 13. The intake port 18 is connected to an intake passage (not shown) through which fresh air introduced into the cylinder 11 flows. A throttle valve 20 that adjusts the intake air flow rate is provided upstream of the intake passage, and the throttle valve 20 receives a control signal from the engine controller 100 and its opening degree is adjusted. On the other hand, the exhaust port 19 is connected to an exhaust passage (not shown) through which burned gas (exhaust gas) from each cylinder 11 flows. An exhaust gas purification system having one or more catalytic converters is disposed in the exhaust passage. The catalytic converter includes, for example, a three-way catalyst.

  The intake valve 21 and the exhaust valve 22 are arranged so that the intake port 18 and the exhaust port 19 can be shut off (closed) from the combustion chamber 17, respectively. The intake valve 21 is driven by an intake valve drive mechanism, and the exhaust valve 22 is driven by an exhaust valve drive mechanism. The intake valve 21 and the exhaust valve 22 reciprocate at a predetermined timing to open and close the intake port 18 and the exhaust port 19 to exchange gas in the cylinder 11. Although not shown, the intake valve drive mechanism and the exhaust valve drive mechanism each have an intake camshaft and an exhaust camshaft that are drivingly connected to the crankshaft. These camshafts are synchronized with the rotation of the crankshaft. Rotate. Further, at least the intake valve drive mechanism includes a hydraulic or mechanical phase variable mechanism (Variable Valve Timing: VVT) 23 that can continuously change the phase of the intake camshaft within a predetermined angle range. ing. You may make it provide the lift variable mechanism (CVVL (Continuous Variable Valve Lift)) which can change a valve lift amount continuously with VVT23.

  The spark plug 31 is attached to the cylinder head 13 by a known structure such as a screw. In this embodiment, the spark plug 31 is attached in a state inclined to the exhaust side with respect to the central axis of the cylinder 11, and the tip thereof faces the ceiling of the combustion chamber 17. The arrangement of the spark plug 31 is not limited to this. The ignition system 32 receives a control signal from the engine controller 100 and energizes the spark plug 31 to generate a spark at a desired ignition timing. As an example, the ignition system 32 may include a plasma generation circuit, and the ignition plug may be a plasma ignition type plug.

  In this embodiment, the fuel injection valve 33 is disposed along the central axis of the cylinder 11 and is attached to the cylinder head 13 with a known structure such as using a bracket. The tip of the fuel injection valve 33 faces the center of the ceiling of the combustion chamber 17. In this example, the fuel injection valve 33 is a piezo-type injector of an outer valve type. This injector has a relatively low fuel spray penetration, but is excellent in fuel atomization. Further, the piezo injector has characteristics that the fuel spray penetration increases and the particle size of the fuel spray increases as the lift amount of the outer opening valve that opens and closes the nozzle opening increases. The piezo injector is also characterized by high responsiveness. However, the fuel injection valve 33 is not limited to the piezo-type injector of the outer opening type.

  The fuel supply system 34 includes a fuel supply system that supplies fuel to the fuel injection valve 33, and an electric circuit that drives the fuel injection valve 33. The electric circuit receives a control signal from the engine controller 100 and operates the fuel injection valve 33 to inject a desired amount of fuel into the combustion chamber 17 at a predetermined timing. Here, the fuel of the engine 1 is gasoline in this embodiment, but is not limited thereto, and may be various liquefied fuels containing gasoline, for example.

  The engine controller 100 is a controller based on a well-known microcomputer, and includes a central processing unit (CPU) that executes a program, a memory that is configured by, for example, RAM and ROM, and stores a program and data, And an input / output (I / O) bus for inputting and outputting signals.

  The engine controller 100 includes at least a signal related to the intake air flow from the air flow sensor 71, a crank angle pulse signal from the crank angle sensor 72, an accelerator opening signal from the accelerator opening sensor 73 that detects the amount of depression of the accelerator pedal, A vehicle speed signal from the vehicle speed sensor 74 is received. The engine controller 100 calculates the following control parameters of the engine 1 based on these input signals. For example, a desired throttle opening signal, fuel injection pulse, ignition signal, valve phase angle signal, etc. The engine controller 100 outputs these signals to the throttle valve 20 (throttle actuator that moves the throttle valve 20), the fuel supply system 34, the ignition system 32, the VVT 23, and the like.

  The characteristic feature of the engine 1 is that the geometrical compression ratio ε of the engine 1 is 18 or more and 40 or less from the viewpoint of improving the illustrated thermal efficiency of the engine and greatly improving the fuel consumption performance compared to the conventional one. In addition to setting the excess air ratio λ to 2 or more and 8 or less in at least a partial load operation region to make the air-fuel mixture lean, the heat insulating structure of the combustion chamber 17 is further combined is there.

  Here, the engine 1 is also an engine 1 having a relatively high expansion ratio at the same time as the high compression ratio because the compression ratio = expansion ratio. In addition, you may employ | adopt the structure (for example, Atkinson cycle and a mirror cycle) used as compression ratio <expansion ratio.

  As shown in FIG. 1, the combustion chamber 17 includes a wall surface of the cylinder 11, a crown surface of the piston 15, a lower surface (ceiling surface) of the cylinder head 13, and valve heads of the intake valve 21 and the exhaust valve 22. The combustion chamber 17 is thermally insulated by providing heat insulation layers 61, 62, 63, 64, and 65 having a configuration described later on each of these surfaces. In addition, below, when these heat insulation layers 61-65 are named generically, a code | symbol "6" may be attached | subjected to a heat insulation layer. The heat insulation layer 6 may be provided on all of these section screens, or may be provided on a part of these section screens. Further, in the illustrated example, the heat insulating layer 61 on the cylinder wall surface is provided at a position above the piston ring 14 in a state where the piston 15 is located at the top dead center. 14 is configured not to slide. However, the heat insulating layer 61 on the cylinder wall surface is not limited to this configuration, and the heat insulating layer 61 may be provided over the entire stroke or a part of the stroke of the piston 15 by extending the heat insulating layer 61 downward. In addition, the thickness of each heat insulation layer 61-65 illustrated in FIG. 1 does not show actual thickness, but is only an illustration, and does not show the magnitude relationship of the thickness of the heat insulation layer in each surface.

In the lean burn engine 1, the geometric compression ratio ε is set to 18 ≦ ε ≦ 40 as described above. The theoretical thermal efficiency η _th in the Otto cycle, which is the theoretical cycle, is η _th = 1−1 / (ε ^κ−1 ), and the theoretical thermal efficiency η _th increases as the compression ratio ε increases. Further, the higher the specific heat ratio κ of gas, in other words, the higher the excess air ratio λ, the higher the theoretical thermal efficiency η _th .

  However, the illustrated thermal efficiency of the engine (more precisely, the engine having no combustion chamber insulation structure) peaks at a predetermined geometric compression ratio ε (for example, about 15), and the geometric compression ratio ε is more than that. However, the illustrated thermal efficiency does not increase, and conversely, the illustrated thermal efficiency decreases. This is because, when the geometric compression ratio is increased while the fuel amount and the intake air amount are kept constant, the higher the compression ratio, the higher the combustion pressure and the combustion temperature. That is, the cooling loss due to heat released through the surface defining the combustion chamber 17 is determined by cooling loss = heat transfer coefficient × heat transfer area × (gas temperature−zone screen temperature), and the combustion gas pressure and The higher the temperature, the higher the heat transfer rate. Therefore, the higher the combustion pressure and the combustion temperature will increase the cooling loss accordingly. As a result, in the lean burn engine, the higher the geometric compression ratio, the lower the illustrated thermal efficiency. In this way, even if it is attempted to increase the indicated thermal efficiency of the engine by increasing the geometric compression ratio while making the air-fuel mixture lean, the increase in cooling loss results in a peak in the indicated thermal efficiency that is significantly lower than the theoretical thermal efficiency. End up.

  On the other hand, in the lean burn engine 1, the heat insulating structure of the combustion chamber 17 is combined so that the illustrated thermal efficiency is increased at a high geometric compression ratio ε. That is, the heat loss of the combustion chamber 17 is reduced to reduce the cooling loss, thereby increasing the indicated thermal efficiency.

  On the other hand, merely reducing the cooling loss by insulating the combustion chamber 17 converts the reduced cooling loss into the exhaust loss and does not contribute much to the improvement in the illustrated thermal efficiency. As described above, the increase in the expansion ratio accompanying the increase in the compression ratio efficiently converts the combustion gas energy corresponding to the reduced cooling loss into mechanical work. That is, it can be said that the lean burn engine 1 greatly improves the illustrated thermal efficiency by adopting a configuration that reduces both the cooling loss and the exhaust loss.

  Here, the excess air ratio λ will be examined. When the excess air ratio λ is lower than 2, the maximum combustion temperature in the combustion chamber 17 becomes high, and RawNOx can be discharged from the combustion chamber 17. As described above, since the lean burn engine 1 aims to reduce exhaust loss as well as cooling loss, the exhaust temperature is relatively low, which is disadvantageous for catalyst activation. Therefore, it is desirable to avoid or suppress the discharge of RawNOx from the combustion chamber 17, and for that purpose, the excess air ratio λ is preferably set to 2 or more. More preferably, it is 2.5 or more. In other words, it is desirable to set the excess air ratio λ in a range where the maximum combustion temperature in the combustion chamber 17 is a predetermined temperature (for example, 1800 K (Kelvin) as a temperature at which RawNOx can be generated) or less. The engine controller 100, for example, within the operation region of the partial load of the engine 1 is accompanied by an increase in the load (in other words, as the excess air ratio λ decreases due to an increase in the fuel injection amount), and the maximum combustion temperature becomes a predetermined temperature. When exceeding the above, it is desirable to operate the engine 1 by increasing the excess air ratio λ.

  On the other hand, according to the study by the inventors of the present application, the illustrated thermal efficiency peaks when the excess air ratio λ = 8. Therefore, the range of the excess air ratio λ is preferably 2 ≦ λ ≦ 8. In the high load operation region including the full load of the engine 1, the excess air ratio λ may be further reduced so that, for example, λ = 1 or λ ≦ 1 by torque priority. The numerical range of the excess air ratio λ is a preferable range of the engine 1 in a medium load and low load operation region. Note that the lean air-fuel mixture sets the throttle valve 20 on the open side, which can contribute to the improvement of the indicated thermal efficiency by reducing the gas exchange loss (pumping loss).

  Here, FIG. 2 shows an example of an operation map when the engine 1 is warm. As described above, the lean operation in which the excess air ratio λ is 2 ≦ λ ≦ 8 (or G / F is 30 to 120) in the low load region where the load is lower than the predetermined load indicated by the solid line in FIG. On the other hand, in a high load region (including a fully open load region) of a predetermined load or more, an air excess ratio λ ≦ 1 is set to λ ≦ 1 operation region. This makes it possible to use a three-way catalyst in a high load region.

  Next, the heat insulation structure of the combustion chamber 17 will be described in more detail. As described above, the heat insulating structure of the combustion chamber 17 is configured by the heat insulating layers 61 to 65 provided on the respective screens that define the combustion chamber 17, and these heat insulating layers 61 to 65 are the combustion in the combustion chamber 17. In order to suppress the release of the heat of the gas through the section screen, the thermal conductivity is set lower than that of the metal base material constituting the combustion chamber 17. Here, for the heat insulating layer 61 provided on the wall surface of the cylinder 11, the cylinder block 12 is the base material, and for the heat insulating layer 62 provided on the crown surface of the piston 15, the piston 15 is the base material. For the heat insulating layer 63 provided on the ceiling surface, the cylinder head 13 is a base material, and for the heat insulating layers 64 and 65 provided on the valve head surfaces of the intake valve 21 and the exhaust valve 22, respectively, the intake valve 21 and the exhaust valve 22 are provided. Are the base materials. Therefore, the material of the base material is aluminum alloy or cast iron for the cylinder block 12, the cylinder head 13 and the piston 15, and the heat-resisting steel or cast iron for the intake valve 21 and the exhaust valve 22. However, as described above, since the lean burn engine 1 has reduced exhaust loss, the exhaust gas temperature has been greatly reduced. It is also possible to use a material that could not be used or was difficult to use (for example, an aluminum alloy).

  In addition, the heat insulating layer 6 preferably has a volumetric specific heat smaller than that of the base material in order to reduce cooling loss. That is, the gas temperature in the combustion chamber 17 varies with the progress of the combustion cycle, but in a conventional engine that does not have a heat insulation structure of the combustion chamber, cooling water flows in a water jacket formed in the cylinder head or cylinder block. Thus, the temperature of the surface defining the combustion chamber 17 is maintained substantially constant regardless of the progress of the combustion cycle.

  On the other hand, as described above, the cooling loss is determined by cooling loss = heat transfer coefficient × heat transfer area × (gas temperature−temperature of the section screen), so that the difference between the gas temperature and the wall surface temperature is large. The higher the loss, the greater the cooling loss. In order to suppress the cooling loss, it is desirable to reduce the difference temperature between the gas temperature and the temperature of the section screen. However, as described above, when the temperature of the section screen of the combustion chamber 17 is maintained substantially constant, It is inevitable that the temperature difference will increase as the temperature changes.

  Therefore, it is preferable that the heat insulating layer 6 has a small heat capacity, and the temperature of the section screen of the combustion chamber 17 changes following the fluctuation of the gas temperature in the combustion chamber 17.

  Further, reducing the heat capacity of the heat insulating layer 6 is advantageous for reducing exhaust loss. That is, if the heat capacity of the heat insulating layer is large, the temperature of the section screen does not decrease even when the temperature in the combustion chamber 17 decreases, but the combustion chamber 17 has a heat insulating structure. The temperature inside is kept high. This results in an increase in exhaust loss and hinders improvement in the thermal efficiency of the engine 1.

  On the other hand, reducing the heat capacity of the heat insulating layer 6 reduces the temperature of the section screen following that when the temperature in the combustion chamber 17 decreases. Therefore, it is possible to avoid maintaining the temperature in the combustion chamber 17 at a high temperature, which can be advantageous in suppressing exhaust loss in addition to the above-described suppression of cooling loss due to temperature followability.

As an example of the heat insulating layer 6, the heat insulating layer 6 includes the wall surface of the cylinder 11, the crown surface of the piston 15, the ceiling surface of the cylinder head 13, and the valve head surfaces of the intake valve 21 and the exhaust valve 22, that is, the combustion chamber. 17 to the partition face defining the, for example, formed by plasma spraying, zirconia (ZrO _2), or may be formed by coating of partially stabilized zirconia (PSZ). Zirconia or partially stabilized zirconia has a relatively low thermal conductivity and a relatively low volumetric specific heat, so that the thermal conductivity is lower than that of the base material and the specific heat of volume is the same as or lower than that of the base material. Layer 6 is constructed.

  Further, in the present embodiment, as shown in FIG. 1, a port liner 181 made of aluminum titanate having an extremely low thermal conductivity, excellent heat insulation, and excellent heat resistance is integrated with the cylinder head 13. A heat insulating layer is provided in the intake port 18 by casting. This configuration suppresses or avoids an increase in temperature due to heat received from the cylinder head 13 when fresh air passes through the intake port 18. As a result, the temperature of the fresh air introduced into the cylinder 11 (initial gas temperature) is lowered, so that the gas temperature at the time of combustion is lowered and the temperature difference between the gas temperature and the section screen of the combustion chamber 17 is reduced. Become advantageous. Lowering the gas temperature at the time of combustion lowers the heat transfer coefficient, which is advantageous for reducing cooling loss. In addition, the structure of the heat insulation layer provided in the intake port 18 is not limited to the casting of the port liner 181.

  As shown in the operation map shown in FIG. 2, in the low load region where the excess air ratio λ ≧ 2, the fuel injected into the cylinder 11 is compressed and ignited by the fuel injection valve 33. FIG. 3 shows an example of a change in the cylinder pressure in the low load region. The engine controller 100 sets the fuel injection timing (main injection) by the fuel injection valve 33, for example, in the first half of the compression stroke or the intake stroke. The fuel injection timing may be changed as appropriate according to the engine load. As a result, a relatively homogeneous air-fuel mixture is formed and compression ignition is performed near the compression top dead center. In FIG. 3, the alternate long and short dash line indicates the change in the cylinder pressure during motoring.

  On the other hand, in the high load region where the excess air ratio λ ≦ 1, as shown in FIG. 5, the engine controller 100 sets the main injection timing by the fuel injection valve 33 to the specific injection timing near the compression top dead center. Before the main injection, the pre-injection that starts within the period from the end of the compression stroke to the compression top dead center is executed.

  Further, in the high load region, the engine controller 100 applies energy to the fuel injected into the cylinder by the plasma ignition by the spark plug 31 after the pre-injection, and self-ignition combustion of the fuel. Assist. As a result, the fuel to which the energy is applied is heated and ignited, and the fuel injected in the subsequent main injection is ignited in a chain manner with the ignited fuel as a base point. In the main combustion period in which the combustion mass ratio of the fuel is 10% or more and 90% or less, the energy is applied within the cylinder, which is the pressure change (ΔP) in the cylinder with respect to the crank angle change (Δθ) during motoring. It is performed within the period from the start of fuel injection to the beginning of the expansion stroke (in the present embodiment, near the compression top dead center) so that the specific crank angle time point at which the rate (ΔP / Δθ) is the negative maximum value is included. .

  Further, it is preferable to apply this energy so that the time when the combustion mass ratio of the fuel becomes 10% comes after the compression top dead center (that is, the main combustion period starts after the compression top dead center).

  Here, as shown in FIG. 4, the cylinder pressure increase rate during motoring is maximum before compression top dead center, is 0 at compression top dead center, and is negative after compression top dead center. Eventually, it becomes a negative maximum value (minimum value). The crank angle point at which the cylinder pressure increase rate during motoring becomes a negative maximum value (that is, the specific crank angle point) varies depending on the geometric compression ratio ε, but the geometric compression ratio ε is 18 or more. If it is 40 or less, it becomes 4 ° to 15 ° CA after compression top dead center (the hatched range in FIG. 4).

  It is assumed that the main combustion starts before the compression top dead center with respect to the change in the cylinder pressure increase rate during the motoring. In this case, since the combustion is performed at the crank angle when the rate of increase in the cylinder pressure during motoring is large, the maximum value of the rate of increase in cylinder pressure during combustion increases. This increases combustion noise and is disadvantageous in terms of NVH performance.

  On the other hand, if combustion starts after compression top dead center or after compression top dead center, combustion occurs at the crank angle when the cylinder pressure increase rate during motoring is small. In comparison, the maximum value of the cylinder pressure increase during combustion is considerably smaller. This reduces combustion noise and improves NVH performance. Therefore, as described above, the period from the start of fuel injection to the initial stage of the expansion stroke is included in the main combustion period so that the specific crank angle point at which the cylinder pressure increase rate during motoring becomes a negative maximum value is included. The energy is given to the fuel injected into the cylinder. In other words, it is possible to paraphrase that the timing of the main injection is set so that the specific crank angle point at which the cylinder pressure increase rate during motoring becomes a negative maximum value is included in the main combustion period. .

  Specifically, the fuel injection in the high load region is particularly preferably performed as follows. That is, pre-injection is performed in which a predetermined amount of fuel is injected from the start of fuel injection to the vicinity of compression top dead center (TDC). The predetermined amount is preferably set to a relatively small amount, for example, 1% or less in terms of mass percentage with respect to all the injected fuel. Therefore, the lift amount of the fuel injection valve 33 at the time of pre-injection is considerably smaller than the lift amount of the fuel injection valve 33 at the time of main injection, which will be described later (excluding the initial stage of main injection). Therefore, the penetration of the fuel spray is small, and the fuel by pre-injection is located in the vicinity of the tip of the spark plug 31 as a fine air-fuel mixture.

  After the pre-injection, the main injection is performed to inject the remaining fuel continuously with respect to the pre-injection. The initial stage of the main injection is in the vicinity of the compression top dead center and is a time when assist of self-ignition combustion is performed as described later. At this time, the lift amount of the fuel injection valve 33 is smaller than that at the time of pre-injection. Thereafter, the lift amount of the fuel injection valve 33 suddenly increases, and all the remaining fuel is injected.

  The engine controller 100 performs fuel injection by pre-injection by plasma ignition by the spark plug 31 within a period from the end of pre-injection to the initial stage of main injection (in this embodiment, near the compression top dead center (initial stage of main injection)). That is, energy is imparted to the fine air-fuel mixture) to assist the self-ignition combustion of the fuel by the pre-injection and the fuel by the main injection. As described above, by performing plasma ignition at the initial stage of main injection (before a large amount of injection), energy can be concentrated on a minute air-fuel mixture. As a result, the air-fuel mixture suddenly rises in temperature and ignites immediately after compression top dead center. . As a result, the cylinder pressure during combustion changes as shown by the solid line in FIG. The two-dot chain line represents the cylinder pressure during motoring, as in FIG.

  By doing so, in the high load region, the maximum value of the cylinder pressure increase rate during combustion is suppressed, combustion noise is reduced, and NVH performance is improved.

  Then, in the driving map shown in FIG. 2, for example, as indicated by an arrow, the driving state in the low load region indicated by the black circle changes to the driving state in the high load region indicated by the white circle in accordance with the acceleration request by the accelerator operation. When shifting, the excess air ratio λ needs to be suddenly changed from λ ≧ 2 to λ ≦ 1. Here, in the low load region where the excess air ratio λ ≧ 2, the throttle valve 20 is set to be fully open, and the intake air amount reduction control is not in time. Therefore, by increasing the fuel injection amount, the excess air ratio λ Must be changed to λ ≦ 1. However, in this case, torque shocks due to sudden changes in torque and NVH performance are deteriorated. Since the engine 1 is a high compression ratio engine, torque shock and NVH performance are likely to be significantly deteriorated.

  Therefore, in the engine 1, transient control is performed when the operation region is shifted between the low load region and the high load region, thereby avoiding torque shock and NVH performance deterioration. Specifically, in the transient control, as shown in FIG. 6, the engine controller 100 sets the excess air ratio λ ≦ 1 and sets the timing of the main injection by the fuel injection valve 33 in the high load region. It is set to the retarded angle side with respect to the specific injection timing near the compression top dead center, so that the main combustion period is greater than the above-mentioned specific crank angle point at which the cylinder pressure increase rate becomes the negative maximum value. Set to the retard side (see the solid line in FIG. 6).

  In addition, pre-injection is also performed in transient control. Similar to the pre-injection in the high load region, the pre-injection is set to start within a period from the end of the compression stroke to the compression top dead center. Further, after pre-injection, the engine controller 100 imparts energy to the fuel injected into the cylinders up to the time of plasma ignition by plasma ignition by the ignition plug 31.

  In this transient control, since the main combustion period is set to be retarded from the specific crank angle point, the main combustion period is greatly retarded in the expansion stroke, so that the combustion efficiency is reduced and the fuel is reduced. Even if the injection amount is increased to set the excess air ratio λ ≦ 1, the generated torque is reduced.

  On the other hand, the main injection period is delayed, thereby delaying the main combustion period, which is disadvantageous in terms of ignitability and combustion stability, but starts within the period from the end of the compression stroke to the compression top dead center The pre-injection and the subsequent plasma ignition maintain the temperature and pressure in the cylinder at a high level even after the compression top dead center, and the fuel injected by the main injection surely reaches the compression ignition, and the main combustion Combustion stability is ensured.

  Here, when the operating state of the engine 1 shifts from the low load region to the high load region in accordance with the accelerator operation as in the case of shifting from the black circle to the white circle in FIG. 2, the operating state of the engine 1 changes to the high load region. Immediately after entering, in other words, at the start of the transient control, the finger pressure waveform of the same BMEP as the net average effective pressure (BMEP) generated in the operating state in the low load region before the start of the transient control is obtained. It is preferable to set the timing of main injection (see the acupressure waveform shown by the solid line in FIG. 6). Then, thereafter, as shown in the acupressure waveform of the broken line and the alternate long and short dash line in FIG. 6, the timing of the main injection is gradually advanced so that the main combustion period is advanced toward the specific crank angle point. By doing so, it becomes possible to gradually increase the torque while avoiding the torque shock, and finally by setting the timing of the main injection so that the specific crank angle time point is included in the main combustion period. The transient control ends. That is, the transition to the high load region is completed (see FIG. 5).

  Next, the transient control executed by the engine controller 100 will be described with reference to the flowchart shown in FIG. This flowchart relates to the transient control when the operating state of the engine 1 shifts from the low load region to the high load region in accordance with the accelerator operation.

  First, in step S71 after the start, it is determined whether or not the excess air ratio λ is λ ≧ 2. That is, it is determined whether the engine operating state is in a low load region. When the determination in step S71 is YES, the process proceeds to step S72.

  In step S72, it is determined whether or not there is an acceleration request due to the accelerator operation. When there is an acceleration request (YES), the process proceeds to step S73.

  In step S73, the operating state as the transition destination is estimated from parameters such as the accelerator depression amount and the engine speed, and a point for switching the excess air ratio to λ ≦ 1 is calculated. Then, in the following step S74, it is determined whether or not there is a switching point to the excess air ratio λ ≦ 1, and there is no switching point, in other words, the engine operating state does not shift to the high load region. When the engine remains in the low load region, the flow is returned, and there is a switching point. In other words, when the engine operating state shifts from the low load region to the high load region, the process proceeds to step S75.

  In step S75, it is determined whether full-open acceleration is requested. When full-open acceleration is requested, that is, when the operating state of the engine 1 finally reaches the full-open load region in the high load region, the process proceeds to step S76, while when full-open acceleration is not requested, that is, When the operating state of the engine 1 finally reaches a load lower than the fully open load in the high load region (for example, when reaching the load indicated by a white circle in FIG. 2), the process proceeds to step S77.

  In step S76, the aforementioned transient control is executed. That is, the pre-injection and the main injection delayed from the specific injection timing are executed, and the plasma ignition is executed between the pre-injection and the main injection. Thus, immediately after shifting to the high load region where the excess air ratio λ ≦ 1, the generated torque is suppressed while setting the excess air ratio λ ≦ 1, thereby avoiding torque shock and NVH performance deterioration. In step S76, the torque is gradually increased by gradually advancing the timing of main injection.

  On the other hand, in step S77 that is not full-open acceleration, in addition to performing the transient control similar to step S76, throttle control, that is, intake air amount reduction control by adjusting the opening of the throttle valve 20 is performed. This is because lean combustion with an excess air ratio λ ≧ 2 is performed in the low load region before the transition of the operating state, so that the throttle valve 20 is fully open and the fully open load after the transition of the operating state is at full open acceleration. Since the throttle valve 20 is fully open even in the range, it is not necessary to adjust the opening of the throttle valve 20, but when the load reaches a load lower than the fully open load after the transition of the operating state, it is set according to the load. This is because the opening of the throttle valve 20 is throttled to meet the fuel injection amount. That is, in step S77, since the opening degree of the throttle valve 20 is reduced after the end of the transient control, the transient control is started by starting to reduce the opening degree of the throttle valve 20 while the transient control is being executed. This is advantageous in improving fuel consumption by reducing the fuel injection amount during the execution of.

  In step S77 as well, as in step S76, the timing of main injection is gradually advanced to gradually increase the torque.

  If the transient control is completed in each of step S76 and step S77, the main injection timing is set in step S78 so that the specific crank angle point is included in the main combustion period, whereby the high load region is set. The transition to will be completed.

  In this way, during the transition in which the operating state of the engine shifts from the low load region to the high load region, by interposing the transient control, while maintaining the emission performance by setting the excess air ratio λ to λ ≦ 1, the torque shock And the deterioration of NVH performance can be avoided.

  On the other hand, when the engine operating state transitions from the high load region to the low load region, the reverse process may be executed. In other words, in the transient control related to the transition from the high load region to the low load region, the BMEP is almost the same as the net average effective pressure (BMEP) generated in the operating state in the high load region before the transient control is started. The timing of the main injection is set so that the acupressure waveform, for example, the acupressure waveform shown by the one-dot chain line in FIG. Then, thereafter, as shown by broken line and solid acupressure waveforms in FIG. 6, the timing of the main injection is gradually retarded so that the main combustion period moves away from the specific crank angle point. By doing so, it is possible to gradually reduce the torque while avoiding the torque shock. Then, when the load is reduced to a load (torque) that can satisfy the excess air ratio λ ≧ 2, the transient control is finished, and the transition to the low load region is completed.

  Note that the transient control at this time is started based on a load set on the high load side with respect to a predetermined load that becomes a boundary between the low load region and the high load region, as shown by a one-dot chain line in FIG. Is preferred. That is, the transient control when shifting from the high load region to the low load region may be performed immediately before entering the low load region. By doing so, it is possible to shift to the low load region without reducing the low load region where the excess air ratio λ ≧ 2, and this is advantageous in improving fuel consumption.

  In this way, even during a transition in which the engine operating state shifts from a high load region to a low load region, by interposing the transient control, the excess air ratio λ is set to λ ≦ 1, while ensuring the emission performance, the torque Shock and deterioration of NVH performance can be avoided.

  The technique disclosed here is not limited to the application to the high compression ratio lean burn engine 1 having the heat insulation structure of the combustion chamber 17 as described above. For example, the heat insulation structure of the combustion chamber 17 is omitted. May be.

1 Engine (Engine body)
11 cylinders
100 Engine controller 20 Throttle valve 31 Spark plug (ignition means)
32 Ignition system (ignition means)
33 Fuel injection valve

Claims (7)

An engine body having a cylinder with a geometric compression ratio set to 18 or more;
A fuel injection valve configured to inject fuel into the cylinder of the engine body;
Ignition means configured to ignite the air-fuel mixture in the cylinder;
A controller configured to control fuel injection into the cylinder through the fuel injection valve and ignition by the ignition means in accordance with an operating state of the engine body,
The controller is
When the operating state of the engine body is in a low load region lower than a predetermined load, the excess air ratio λ during combustion is set to 2 or more,
When the operating state of the engine body is in a high load region above the predetermined load, the excess air ratio λ during combustion is set to 1 or less, and the combustion mass ratio of the fuel is set to 10% or more and 90% or less. In the main combustion period, the fuel injection timing by the fuel injection valve is compressed so that a specific crank angle point at which the rate of increase in the cylinder pressure during the motoring of the engine body becomes a negative maximum value is included. Set to a specific injection time near the dead center,
The controller also controls the main combustion while keeping the excess air ratio λ during combustion at 1 or less during a transition in which the operating state of the engine body transitions between the low load region and the high load region. A spark ignition direct injection engine that performs transient control for delaying the fuel injection timing from the specific injection timing so that the period is retarded from the specific crank angle time point. In the spark ignition direct injection engine according to claim 1,
During the transient control, the controller causes the fuel injection valve to pre-injection that starts within a period from the end of the compression stroke to the compression top dead center, and after the pre-injection and later than the specific injection timing. A spark ignition direct injection engine that performs main injection. The spark ignition direct injection engine according to claim 2,
The controller is a spark ignition direct injection engine that causes the ignition means to perform ignition between the pre-injection and the main injection. The spark ignition direct injection engine according to any one of claims 1 to 3,
The controller performs the transient control at the time of transition in which the operation state of the engine body shifts from the low load region to the fully open load region in the high load region, whereas the operation state of the engine body is the A spark ignition direct injection engine that further performs a reduction control of an intake air amount during execution of the transient control during a transition from a low load region to the high load region where the load is lower than the fully open load. A control method for a spark ignition direct injection engine having a cylinder with a geometric compression ratio set to 18 or more and configured to inject fuel directly into the cylinder,
When the operating state of the engine is in a low load region lower than a predetermined load, the engine is operated such that the excess air ratio λ during combustion is 2 or more,
When the operating state of the engine is in a high load region equal to or higher than the predetermined load, the excess air ratio λ during combustion is set to 1 or less, and the timing of fuel injection into the cylinder is set to the compression top dead center. By setting the specific injection timing in the vicinity, the cylinder pressure increase rate during motoring of the engine becomes a negative maximum value within the main combustion period in which the combustion mass ratio of the fuel is 10% or more and 90% or less. The engine is operated such that a specific crank angle point is included, and
When the engine operating state transitions between the low load region and the high load region, the excess air ratio λ during combustion is set to 1 or less and the timing of the fuel injection is set to the specific injection. A control method for a spark ignition direct injection engine that operates the engine so that the main combustion period is retarded from the specific crank angle point by executing transient control that is delayed from the timing. In the control method of the spark ignition direct injection engine according to claim 5,
At the time of the transient control, as fuel injection into the cylinder, pre-injection that starts within a period from the end of the compression stroke to compression top dead center, and main injection that is delayed after the pre-injection and before the specific injection timing And a spark ignition direct injection engine control method for igniting an air-fuel mixture in the cylinder between the pre-injection and the main injection. In the control method of the spark ignition direct injection engine according to claim 5 or 6,
When the engine operating state transitions from the low load region to the fully open load region in the high load region, the transient control is executed,
A spark ignition type direct injection engine that further performs intake air amount reduction control during the execution of the transient control during a transition in which the operating state of the engine shifts from the low load region to the high load region where the load is lower than the fully open load. Control method.

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