JP2010159683A - Internal combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To improve a specific power of an internal combustion engine, eliminating restrictions in knocking. <P>SOLUTION: The internal combustion engine includes a turbocharger and a variable compression ratio mechanism for varying a mechanical compression ratio. In a first operating range where a torque is increased by supercharging but a load is relatively small, a spark ignited combustion mode (SI) is set, where combustion approximated to a constant volume combustion cycle by a spark ignition of a premixed air fuel mixture is achieved. In a second operating range where the load is relatively large, a constant pressure combustion mode is set, where combustion approximated to a constant pressure combustion cycle by a control of a heat generating ratio is achieved. In the spark ignited combustion mode, a target compression ratio is set lower as the load increases. On the other hand, in the constant pressure combustion mode, the target compression ratio is set higher in order to compensate a time loss due to lengthening of the combustion. Thereby, a greater amount of fuel and fresh air can be supplied into a cylinder to improve the specific power. <P>COPYRIGHT: (C)2010,JPO&INPIT

Description

  The present invention relates to an internal combustion engine capable of obtaining a high maximum output, that is, a specific output with respect to the weight and external dimensions of the internal combustion engine.

In order to improve the maximum output while suppressing an increase in the size and weight of an internal combustion engine, a supercharging technique and a variable compression ratio mechanism are conventionally known. For example, Patent Document 1 discloses a supercharger and a variable compression. A spark ignition internal combustion engine that combines a ratio mechanism and a variable valve timing mechanism is disclosed. In this device, in a high load range, the compression ratio by the variable compression ratio mechanism is lowered while supercharging, and the effective compression ratio is lowered by the variable valve timing mechanism to prevent knocking.
JP 63-120820 A

  As is well known, the torque of the internal combustion engine correlates with the amount of fuel supplied into the cylinder, that is, the maximum output depends on how much fresh air can be introduced into the cylinder. Therefore, in order to obtain a high maximum output in a limited size and weight, it is necessary to apply a higher boost pressure. However, as in Patent Document 1 described above, the compression ratio is reduced in a high load range. Even if knocking is suppressed, if the supercharging pressure is set high, the combustion pressure becomes excessive, and knocking occurs in any case, so there is a limit in improving the output. Further, as the compression ratio is lowered in order to avoid knocking, the efficiency deteriorates and the fuel consumption deteriorates.

  Therefore, the present invention utilizes a constant pressure combustion cycle and combines a high supercharging pressure and a variable compression ratio mechanism to supply more fuel and fresh air into the cylinder, thereby improving the specific output. It is a thing. That is, an internal combustion engine according to the present invention includes a supercharger, a variable compression ratio mechanism that changes a mechanical compression ratio of the internal combustion engine, a fuel injection device that injects fuel in an amount corresponding to a load into a cylinder, A spark ignition combustion mode that realizes combustion approximated to a constant volume combustion cycle by spark ignition of a premixed gas, and a constant pressure combustion mode that realizes combustion approximated to a constant pressure combustion cycle by controlling the heat generation rate And can be switched.

  In the first operation region where the load is relatively small, the spark ignition combustion mode is set, and the compression ratio by the variable compression ratio mechanism is set to be lower as the load is higher. In the second operation region, the constant pressure combustion mode is set, and the compression ratio by the variable compression ratio mechanism is set high.

  In the constant pressure combustion mode, for example, by starting the generation of heat near top dead center and controlling the fuel supply so that the calorific value increases in proportion to the volume change rate, combustion approximating the constant pressure combustion cycle can be realized. As a result, even in a state where a large amount of fuel is supplied with high supercharging, the maximum combustion pressure can be suppressed lower than in the constant volume combustion cycle. In the constant pressure combustion cycle, the time loss increases because the combustion time becomes long, but at the same time, the efficiency reduction due to the prolonged combustion is compensated by setting the compression ratio high. Further, in the first operating region where the load is relatively small, the spark ignition combustion mode is the same as that of a general premixed spark ignition internal combustion engine, so that high efficiency can be obtained by ignition near the so-called MBT point.

  The boundary between the first operation region and the second operation region is, for example, the limit at which ignition can be performed at the MBT point in the spark ignition combustion mode, that is, if the load is higher than this, the ignition is performed more than the MBT point to suppress knock This corresponds to the load limit that needs to be retarded.

  In the present invention, it does not mean that the theoretically complete constant pressure combustion cycle and constant volume combustion cycle are realized, but as a comparison between the two combustion modes, one of them is a relatively constant pressure combustion cycle. It is sufficient if the other is close and the other is relatively close to the constant volume combustion cycle.

  According to the present invention, more fuel can be burned in the cylinder while avoiding knocking in a region where the load near the full load is high, and the specific output of the internal combustion engine can be improved. In addition, there is no reduction in efficiency due to the retard of the ignition timing or a reduction in the compression ratio, and high efficiency can be maintained.

  Hereinafter, an embodiment of the present invention will be described in detail with reference to the drawings.

  As schematically shown in FIG. 1, an internal combustion engine 1 to which the present invention is applied includes a turbocharger 2 as a supercharger between an exhaust passage 4 and an intake passage 3, and a supercharging pressure sensor. The engine control unit 6 adjusts the opening degree of the exhaust bypass valve 7 based on the supercharging pressure detected by the engine 5, so that the supercharging pressure is controlled to a desired target supercharging pressure.

  FIG. 2 shows a system configuration of the internal combustion engine 1. A piston 22 having a recess 22 a on the crown surface is slidably fitted to a cylinder 21, and a combustion chamber 23 defined by the piston 22. An intake port 24 and an exhaust port 25 are connected to each other, and a fuel injection valve 26 and a spark plug 27 are disposed at the top of the combustion chamber 23. The intake port 24 and the exhaust port 25 are opened and closed by an exhaust valve 28 and an intake valve 29, respectively. High-pressure fuel is supplied to the fuel injection valve 26 via the fuel pressure control device 30, and fuel injection is performed based on a control signal from the control unit 6 as described later. The pressure of the fuel supplied to the fuel injection valve 26, that is, the fuel pressure, is detected by the fuel pressure sensor 31, and feedback control is performed based on the detected pressure.

  The control unit 6 that controls the entire internal combustion engine 1 is connected to the above-described sensors, an air flow meter 32 that detects the intake air amount, and an accelerator that detects the amount of depression of the accelerator pedal operated by the driver. Signals of known sensors such as an opening sensor 33, a crank angle sensor 34, and a coolant temperature sensor 35 are input.

  The internal combustion engine 1 includes a known multi-link variable compression ratio mechanism 36, and the compression ratio can be variably controlled by vertically displacing the position of the piston 22 with respect to the crankshaft. . This variable compression ratio mechanism 36 is of the type described in Japanese Patent Application Laid-Open No. 2005-147068, Japanese Patent Application Laid-Open No. 2002-61501, and the like, and the piston 22 is disclosed in the former document. In addition, it is desirable to shorten the pin boss portion so that the counterweight passes through both sides of the pin boss portion so that the engine has a longer stroke.

  In the present invention, the variable compression ratio mechanism 36 is not necessarily limited to the above-mentioned type, and various types can be used.

  In order to realize the constant pressure combustion cycle, the heat generation characteristics as shown in FIG. FIG. 3B shows the in-cylinder pressure change corresponding to FIG. Note that these are based on numerical calculation of simulation. In the constant pressure combustion cycle, heat generation is started generally from the top dead center, and the heat generation amount is increased toward the later stage of combustion. Then, the maximum amount of heat generation is obtained in the later stage of combustion, and combustion is completed simultaneously throughout the cylinder. Although the amount of heat generation is gradually increased toward the latter stage of the combustion process, the resulting heat generation pattern is generally proportional to the volume change of the combustion chamber 23. In order to cope with the difference in piston motion due to so-called piston offset or the like, the heat generation pattern may be determined in accordance with the actual volume change rate of the combustion chamber.

  In the constant pressure combustion cycle, it is necessary to appropriately control the heat generation pattern and the heat generation period in order to keep the pressure at the compression end in the combustion chamber substantially constant. In the constant pressure combustion cycle, ignition combustion starts near the top dead center. In order to suppress or control self-ignition, so-called “spray combustion” (Japanese Patent Laid-Open No. 2007-285244, Japanese Patent Laid-Open No. 2008-121429). Etc.) is performed. Therefore, the injection rate of the fuel injection valve 26 is controlled so as to gradually increase toward the latter half of the heat generation, as shown in FIG.

  FIGS. 5A to 5C show an example of the fuel injection valve 26 suitable for such a constant pressure combustion cycle. The needle hole 42 is opened and closed by the needle 41, and the tip of the needle 41 is also opened. A guide portion 43 is provided. When the lift is low, the spray is guided by the guide portion 43 with a wide cone angle as shown in FIG. 5B, and when the lift is high, the spray is sprayed as shown in FIG. Is straight and jetted at a relatively narrow cone angle. That is, it is possible to perform injection in a form that changes from FIG. 6A to FIG. By changing the injection direction and the injection rate in this manner, the air utilization rate in the cylinder can be improved, and combustion at the stoichiometric air-fuel ratio becomes possible.

  FIG. 7 is a characteristic diagram showing the relationship between the combustion mode and the compression ratio. As shown in the figure, in the first operation region where the load is relatively small, the spark ignition combustion mode (abbreviated as “SI” in the drawing) is performed, and the second operation where the load is relatively large near the full load. Within the region, the constant pressure combustion mode is set. Note that supercharging by the turbocharger 2 is performed in almost the entire operation region.

  The spark ignition combustion mode performs combustion similar to a constant volume combustion cycle similar to that of a general premixed spark ignition internal combustion engine.For example, fuel is injected into a cylinder during an intake stroke to form a premixed gas. The premixed gas is ignited at the MBT point. In this constant volume combustion cycle, the higher the load, the more likely knocking occurs. Therefore, the higher the load, the lower the target compression ratio by the variable compression ratio mechanism 36 is set. Note that the target compression ratio is high in the low-speed and low-load region, and the target compression ratio is shown in two steps of high and low in the figure.

  Similarly, in order to avoid knocking, the ignition timing is controlled to be retarded from the MBT point on the high load side of the first operating region. The boundary L between the spark ignition combustion mode and the constant pressure combustion mode shown in the figure is the limit at which ignition at the MBT point is possible in the spark ignition combustion mode, that is, if the load is higher than this, the ignition timing is set higher than the MBT point to suppress knocking. It substantially corresponds to the load limit that needs to be retarded.

  On the other hand, the constant pressure combustion mode performs combustion similar to a constant pressure combustion cycle as described above, and as described above, as so-called “spray combustion”, fuel injection is started near top dead center and ignition is performed. The fuel supply is controlled so that heat generation is started and the calorific value increases in proportion to the volume change rate. In this constant pressure combustion cycle, even when the torque is the same as that in the constant volume combustion cycle, the maximum combustion pressure is lowered, and therefore, more fuel is supplied without being restricted by knocking so that the torque can be improved. Can do. That is, compared with the case of only the spark ignition combustion mode, the fully open torque can be set to a higher torque side. In this constant pressure combustion cycle, the target compression ratio of the variable compression ratio mechanism 36 is set high. It should be noted that the target compression ratio in the second operating region may be fixedly set to a constant high compression ratio, or the compression ratio may be lowered as the load increases. In the constant pressure combustion cycle, the time loss increases because the combustion time becomes long. However, by giving a high compression ratio in this way, the efficiency reduction due to the prolonged combustion is compensated. Therefore, the target compression ratio basically has a discontinuous characteristic at the boundary between the first operation region and the second operation region.

  The determination of the first operation region and the second operation region, that is, the switching of the combustion mode, is based on a predetermined map using the engine speed Ne and the load (for example, fuel injection amount) Te as parameters as shown in FIG. Although it may be simply performed, it is desirable to calculate the efficiency in each combustion mode according to the flowchart of FIG. 8 and to select a combustion mode with higher efficiency by comparison.

  Specifically, first, in step 1, the required air amount is obtained from the engine speed and the required torque. In step 2, the target compression ratio in the case of the spark ignition combustion mode is determined based on the map, and in step 3, the ignition timing in the case of the spark ignition combustion mode is determined. The ignition timing is basically the MBT point as described above, but correction based on the intake air temperature may be added. On the other hand, in parallel with this, at step 4, the target compression ratio when the constant pressure combustion mode is set is determined based on the map, and at step 5, the combustion period when the constant pressure combustion mode is set is calculated. In order to simplify the process, the fuel injection period can be regarded as a combustion period. In this embodiment, the ignition timing in the constant pressure combustion mode is a constant ignition timing near top dead center.

  Next, in step 6, the efficiency in the spark ignition combustion mode is compared with the efficiency in the constant pressure combustion mode. If the former is high efficiency, the spark ignition combustion mode is selected, and the latter is high efficiency. If so, the constant pressure combustion mode is selected (steps 7 and 8). The efficiency in the spark ignition combustion mode is, for example, the amount of intake air, engine speed, ignition timing, and EGR gas amount (the sum of so-called internal exhaust gas recirculation and external exhaust gas recirculation) under the target compression ratio determined in step 2. It can be obtained as a function. On the other hand, the efficiency in the constant pressure combustion mode is obtained as a function of the intake air amount, the engine speed, the effective compression ratio, and the expansion ratio (which varies depending on the combustion period) under the target compression ratio determined in Step 4. Can do. Since the internal combustion engine 1 is actually operated in one of the combustion modes, the efficiency of one combustion mode is obtained sequentially using a part of actual parameters corresponding to the operation conditions at that time, and the other The efficiency of the combustion mode is sequentially estimated according to the operating conditions at that time.

  If the ignition timing is retarded from the MBT point in the spark ignition combustion mode, the efficiency deteriorates rapidly. In an extreme case, so-called “fuel cooling” is required for the rise in the exhaust gas temperature accompanying the retardation of the ignition timing, and the efficiency is further deteriorated because the air-fuel ratio is set to be rich. On the other hand, the constant pressure combustion cycle has a large time loss and is generally less efficient than the constant volume combustion cycle with the ignition timing as the MBT point. Therefore, when the efficiency in each combustion mode is compared and the higher efficiency is selected, the boundary L along the limit of the MBT point as shown in FIG. 6 is obtained as a result.

  Next, the embodiment shown in FIG. 9 is provided with a knocking sensor 51 that detects that knocking has occurred. When knocking occurs in the spark ignition combustion mode, the ignition timing is set to MBT by a known method. The delay angle is corrected from the point.

  In the present invention, the same knocking sensor 51 may be used to switch to the constant pressure combustion mode when knocking is detected during operation in the spark ignition combustion mode. Note that switching from the constant pressure combustion mode to the spark ignition combustion mode can be performed by, for example, returning to the spark ignition combustion mode when knocking is not detected for a certain period of time.

  Next, an embodiment in which a variable valve mechanism is provided on the intake valve side in order to variably control the effective compression ratio will be described. The variable valve mechanism needs to be configured to at least delay the closing timing of the intake valve. For example, an electromagnetically driven intake valve or a mechanical variable valve mechanism (Japanese Patent Laid-Open No. 2002-89303) , JP-A-2002-285898, etc.) can be used. In this embodiment, in the constant pressure combustion mode (especially in the full load region), by reducing the effective compression ratio using a variable valve mechanism, more fuel is burned while suppressing the maximum combustion pressure. Is possible. Here, as the variable valve mechanism, a mechanical variable valve mechanism of a type capable of expanding and reducing the intake valve operating angle while keeping the opening timing of the intake valve substantially constant is used. Accordingly, when the operating angle is reduced, the intake valve closing timing is advanced.

  FIG. 10 shows an example of the valve timing of the intake / exhaust valve in a state where the effective compression ratio is lowered. As shown in the figure, the so-called early closing in which the intake valve closing timing IVC is advanced from the bottom dead center. It has become. As a result, as indicated by the line L1 in the PV diagram of FIG. 11, the expansion ratio is relatively large with respect to the effective compression ratio, resulting in a so-called mirror cycle. The characteristic indicated by the line L2 in FIG. 11 shows a case where the supercharging pressure is low and the IVC is near the bottom dead center. As is well known, in order to reduce the effective compression ratio, it is also possible to employ so-called delayed closing in which the intake valve closing timing IVC is retarded from the bottom dead center.

  FIG. 12 is a characteristic diagram showing the relationship among the combustion mode, the compression ratio, and the intake valve operating angle in this embodiment. As shown in the figure, the spark ignition combustion mode (SI) is set in the first operation region having a relatively small load, and the constant pressure combustion mode is set in the second operation region having a relatively large load including the entire load region. It becomes. Note that supercharging by the turbocharger 2 is performed in almost the entire operation region.

  In the spark ignition combustion mode, as described above, combustion close to a constant volume combustion cycle is performed by ignition near the MBT point with respect to the premixed gas, but the compression ratio control by the variable compression ratio mechanism 36 is the same as in the above-described embodiment. In the region on the low speed and low load side, the target compression ratio is high, and the higher the load, the lower the target compression ratio is set. The operating angle correlated with the effective compression ratio becomes a small operating angle in the low-speed and low-load region, and the operating angle expands to a medium operating angle as the load increases. As with the compression ratio, the operating angle also changes continuously.

  As in the above-described embodiment, the boundary L3 between the spark ignition combustion mode and the constant pressure combustion mode is the limit at which ignition at the MBT point is possible in the spark ignition combustion mode, that is, the MBT point for suppressing knocking when the load is higher than this. This substantially corresponds to the limit of the load that needs to retard the ignition timing.

  In the constant pressure combustion mode, combustion similar to a constant pressure combustion cycle is performed by “spray combustion”. In this constant pressure combustion mode, the target compression ratio of the variable compression ratio mechanism 36 is high, as in the above-described embodiment. Is set. The effective compression ratio, i.e., the operating angle of the intake valve, is a large operating angle for securing torque except for the entire load region above the line L4. More specifically, the operating angle increases with load, and the intake valve closing timing IVC is near the bottom dead center near the line L4.

  In the constant pressure combustion mode region, in the full load region, that is, the region sandwiched between the lines L4 and L5, the operating angle is set small, and the effective compression ratio is suppressed. As a result, knocking or excessive increase in the maximum combustion pressure can be avoided, and the full opening torque can be obtained higher. In other words, the characteristic of the line L4 substantially corresponds to the fully open torque of the embodiment of FIG. 7, and when combined with a decrease in the effective compression ratio by the variable valve mechanism, the fully open torque can be increased to the line L5. It becomes possible.

  The determination of the first operation region and the second operation region, that is, the switching of the combustion mode, is based on a predetermined map using the engine speed Ne and the load (for example, fuel injection amount) Te as parameters as shown in FIG. In this embodiment, the efficiency in each combustion mode is calculated in the same procedure as the flowchart of FIG. 8 described above, and a combustion mode with high efficiency is selected by comparison. It is possible to switch.

  Of course, in this case, the target intake valve opening / closing timing (that is, the effective compression ratio) is determined according to the operating conditions at that time, and the efficiency in each combustion mode is calculated in consideration of this effective compression ratio. Become.

Explanatory drawing which shows the outline of the internal combustion engine which concerns on this invention. Explanatory drawing which shows the system configuration | structure of this invention. The characteristic view which shows the characteristic of (a) heat release rate and (b) in-cylinder pressure for a constant pressure combustion cycle. The characteristic view which shows the injection rate change in the case of a constant pressure combustion cycle. Explanatory drawing which shows an example of the fuel injection valve suitable for a constant pressure combustion cycle. Explanatory drawing which shows the change of the spray by this fuel injection valve. The characteristic view which shows the combustion mode with respect to an operating condition, and a target compression ratio. The flowchart of combustion mode switching. Explanatory drawing which shows the Example provided with the knocking sensor. The valve timing chart in the Example provided with the variable valve mechanism. The PV diagram of this Example. The characteristic view which shows the combustion mode etc. with respect to driving | running conditions.

DESCRIPTION OF SYMBOLS 1 ... Internal combustion engine 2 ... Turbocharger 6 ... Engine control unit 36 ... Variable compression ratio mechanism

Claims (6)

A turbocharger,
A variable compression ratio mechanism for changing the mechanical compression ratio of the internal combustion engine;
A fuel injection device that injects fuel in an amount corresponding to the load into the cylinder;
An ignition device;
Switching between a spark ignition combustion mode that realizes combustion approximating a constant volume combustion cycle by spark ignition of a premixed gas and a constant pressure combustion mode that realizes combustion approximated to a constant pressure combustion cycle by controlling the heat generation rate In the first operating region where the load is relatively small, the spark ignition combustion mode is set, and the higher the load, the lower the compression ratio by the variable compression ratio mechanism is set.
In the second operating region where the load is relatively large, the constant pressure combustion mode is set, and the compression ratio by the variable compression ratio mechanism is set high.   The efficiency in the spark ignition combustion mode and the efficiency in the constant pressure combustion mode are sequentially calculated with respect to the operating conditions at that time, and the combustion mode is switched to a combustion mode that has a relatively high efficiency. Internal combustion engine.   A variable valve mechanism for changing the effective compression ratio by changing at least the closing timing of the intake valve; and reducing the effective compression ratio by the variable valve mechanism in the vicinity of the maximum load in the second operating region. The internal combustion engine according to claim 1 or 2, characterized by the above.   In the constant pressure combustion mode, fuel injection is started near top dead center, ignition is started to start heat generation, and fuel supply is controlled so that the heat generation amount increases in proportion to the volume change rate. The internal combustion engine according to any one of claims 1 to 3.   The internal combustion engine according to any one of claims 1 to 4, wherein a fuel injection valve having a variable injection rate is used, and combustion close to a constant pressure combustion cycle is realized by controlling the injection rate.   There is further provided ignition timing correction means for correcting the ignition timing in the spark ignition combustion mode from the MBT point based on the detection of knocking, and the spark ignition combustion mode and the constant pressure combustion mode are arranged in the vicinity of a load where the retardation correction occurs. 6. The internal combustion engine according to claim 1, wherein the internal combustion engine is switched.

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