JP2004084531A - Compressive ignition type internal combustion engine and hybrid automobile - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To realize lower fuel consumption by improving efficiency in a low rotation and high load area, where compressive ignition combustion is comparatively easily performed. <P>SOLUTION: In this compressive ignition type internal combustion engine, a compression ratio CR is set to 13.5-16.5. A supercharger is installed, and a supercharged pressure Pc is set to a value, where an illustrated average effective pressure obtained in the setting compression ratio becomes maximum, in a range of 25-75 kPa (CR=15, Pc=50 kPa). This hybrid automobile is constituted with such a compressive ignition type internal combustion engine and an electric motor as drive sources, and a motor operating mode is set as a low output area operating mode. <P>COPYRIGHT: (C)2004,JPO

Description

[0001]
TECHNICAL FIELD OF THE INVENTION
The present invention relates to a compression ignition type internal combustion engine and a hybrid vehicle equipped with the same as a power source.
[0002]
[Prior art]
In order to achieve low fuel consumption, a compression ignition type internal combustion engine is employed. In this internal combustion engine, a minus overlap period in which both the intake valve and the exhaust valve are closed before and after the exhaust top dead center is provided, and the air-fuel mixture is controlled by the internal EGR. Is known to expand the compression ignition combustion operation region to the high rotation and low load side (Japanese Patent Application Laid-Open No. 10-266878).
[0003]
[Problems to be solved by the invention]
However, this technique has the following problems. That is, when the air-fuel mixture is heated by the internal EGR to a high temperature, the wall heat loss increases accordingly. For this reason, even if the compression ignition combustion operation region is expanded, a significant improvement in efficiency cannot be expected in this expanded region.
[0004]
Therefore, an object of the present invention is to improve the efficiency in a low-rotation, high-load region where compression ignition combustion is relatively likely to occur in a compression ignition type internal combustion engine, thereby realizing further lower fuel consumption.
[0005]
[Means for Solving the Problems]
For this reason, in the present invention, in the compression ignition type internal combustion engine, a supercharger is installed, and the maximum torque point in the combustion stable region of the compression ignition combustion determined for each compression ratio is set to the supercharging pressure obtained under the atmospheric suction. At the same time, under the supercharging pressure, the compression ratio is set to a value at which the indicated average effective pressure of the compression ignition combustion calculated for the maximum torque point becomes substantially maximum.
This makes it possible to generate compression ignition combustion without using the air-fuel mixture heating means such as the internal EGR in the low-speed high-load region, so that the efficiency is improved and further lower fuel consumption can be realized.
[0006]
In addition, a hybrid vehicle is configured by using such a compression ignition type internal combustion engine and an electric motor as drive sources, and a mode in which the internal combustion engine is stopped and the vehicle runs with the output from the electric motor is set as a low output range driving mode.
[0007]
BEST MODE FOR CARRYING OUT THE INVENTION
Embodiments of the present invention will be described below with reference to the drawings.
FIG. 1 is a sectional view of an automobile engine (hereinafter, referred to as “engine”) 1 according to an embodiment of the present invention.
A cylinder head 3 is fixed on the cylinder block 2. A piston 4 is inserted into the cylinder block 2, and a combustion chamber 5 is formed between the piston 4 and the cylinder head 3. A port portion 31 of an intake passage is formed on one side of the cylinder head 3, and the port portion 31 and the surge tank 6 are connected by an intake pipe 7a. An injector 8 is provided in the intake pipe 7a, and gasoline as fuel is injected from the injector 8. The engine 1 is provided with a turbocharger 9 as a supercharger, and a compressor section thereof and the surge tank 6 are connected by an intake pipe 7b. On the other hand, a port portion 32 of an exhaust passage is formed on the other side of the cylinder head 3, and the port portion 32 and a turbine portion 91 of the turbocharger 9 are connected by an exhaust pipe 10. The exhaust pipe 10 is provided with an exhaust gas purification device 11 containing an oxidation catalyst or a three-way catalyst as a catalyst.
Here, in the engine body, the cylinder head 3 is provided with an intake valve 12 for opening and closing the port 31 and an exhaust valve (not shown) for opening and closing the port 32. Of these, the valve train 13 of the intake valve 12 is of a variable type so that the phase of the intake cam 131 can be advanced or delayed with respect to the crank angle. A rocker cover 14 is mounted on the cylinder head 3 and seals a camshaft storage space of the cylinder head 3.
Further, an ignition plug 15 is provided in the cylinder head 3 so as to face the upper center of the combustion chamber 5.
In such an engine 1, air that has passed through an air cleaner (not shown) is compressed by a compressor section of the turbocharger 9 and sent out to an intake pipe 7b. Then, the surge tank 6 reaches the port section 31 via the corresponding manifold section, and a mixture with the fuel injected from the injector 8 is sucked into the cylinder. The exhaust gas generated as a result of the combustion is sent out to the exhaust pipe 10 as the piston 4 rises, and flows into the exhaust purification device 11. In the exhaust gas purification device 11, unburned fuel and carbon monoxide in the exhaust gas are purified. Then, the exhaust gas flows into the turbine section 91 of the turbocharger 9, rotates the turbine wheel, passes through the turbine section 91, and is discharged into the atmosphere.
Here, the engine controller 101 having a function as combustion control means generates a command signal for operating the injector 8, the variable valve device 13, and the ignition plug 15. The engine controller 101 receives various operation state detection signals such as an accelerator opening detection signal from an accelerator sensor, an engine speed detection signal from a crank angle sensor, and a cooling water temperature detection signal from a water temperature sensor. The engine controller 101 determines the operating state of the engine 1 based on the input signal, and generates a command signal according to the determined operating state.
The engine 1 is operated by switching the combustion mode between compression ignition combustion and spark ignition combustion according to the operation state. When the engine controller 101 determines that the engine 1 is in the low-speed high-load region (region A in FIG. 2), it generates a command signal for performing compression ignition combustion. The variable valve device 13 is operated by this command signal, and the lift curve of the intake cam 131 is set to the advance side profile Ca (FIG. 3). In this profile Ca, the intake valve closing timing IVC is the first timing IVCa. On the other hand, when the engine controller 101 determines that the engine 1 is in an operation region other than the region A (region B in FIG. 2), the engine controller 101 generates a command signal for performing spark ignition combustion. The variable valve device 13 is operated by this command signal, and the lift curve of the intake cam 131 is set to the profile Cb on the retard side (FIG. 3). In the profile Cb, the intake valve closing timing IVC is a second timing IVCb that is later with respect to the crank angle than the first timing IVCa. Further, at the time of spark ignition combustion, the engine controller 101 opens the waste gate 911 formed in the turbine section 91 of the turbocharger 9 and flows the exhaust gas around the turbine section 91.
[0008]
Next, the setting of the supercharging pressure Pc and the compression ratio CR of the engine 1 will be described.
In the engine 1, the supercharging pressure Pc by the turbocharger 9 is set to 50 kPa (0.5 bar), and the compression ratio CR is set to 15.
FIG. 4 shows a combustion stable region A of compression ignition combustion determined for each compression ratio CR with respect to the intake air temperature Ti and the air-fuel ratio AF. In the figure, the entire region for Ti and AF is divided into a combustible region C on the high intake air temperature side and a misfire region D on the low intake air temperature side with reference to the combustion stability limit line. In the combustible region C, compression ignition combustion can be generated, while in the misfire region D, compression ignition combustion does not occur or ignition becomes unstable. The combustible region C is divided into a combustion stable region A on the high air-fuel ratio side and a knocking occurrence region B on the low air-fuel ratio side based on the knocking limit line. In the stable combustion region A, the combustion is relatively slow, so that the compression ignition combustion can be stably generated. On the other hand, in the knocking occurrence region B, the combustion becomes steep, so that knocking occurs. Thus, the combustion stable region A of the compression ignition combustion is defined as a wedge-shaped region sandwiched between the combustion stability limit line and the knocking limit line with respect to the intake air temperature Ti and the air-fuel ratio AF. Here, the amount of fuel supplied to the engine 1 increases on the low intake air temperature side in terms of charging efficiency, and increases on the low air-fuel ratio side under a constant intake air temperature. Therefore, assuming that the thermal efficiency is constant, the maximum torque at the compression ratio is obtained at the vertex P of the wedge in the area A.
[0009]
FIG. 5 shows, for each compression ratio CR, a combustion stable region A of compression ignition combustion in a case where the engine 1 is operated at an engine speed of 1200 rpm under the natural suction performed while the supercharging by the turbocharger 9 is stopped. I have. CR = 18 (A11), 16.5 (A21), 15 (A31), 13.5 (A41), and 12 (A51) are shown. As shown in the figure, the combustion stable region A under the natural intake is formed on the lower intake temperature side as the compression ratio CR is larger, and the combustion stable region A under the atmospheric intake of Ti = 25 to 50 ° C. at CR = 18 (A11). Enables compression ignition combustion.
[0010]
FIG. 6 shows the indicated mean effective pressure IMEP calculated for each of the maximum torque points P11 to P51 in the combustion stable regions A11 to A51 in FIG. Under natural aspiration, IMEP increases as the compression ratio CR increases, and is maximized at CR = 18 (P11). In the figure, IMEP is 540 kPa for P11, 520 kPa for P21, 450 kPa for P31, 430 kPa for P41, and 410 kPa for P51.
As described above, under the natural intake, the indicated average effective pressure IMEP becomes the maximum at the compression ratio CR = 18 at which the compression ignition combustion can be performed under the atmospheric suction, and the efficiency is also the best. This means that when the CR is 12 to 16.5, the air-fuel mixture needs to be heated in some way to cause compression ignition combustion, such as forming a negative overlap period between the intake valve 12 (FIG. 1) and the exhaust valve. Due to increased wall heat loss.
[0011]
FIG. 7 shows a combustion stable region A of compression ignition combustion when the engine 1 is operated at an engine speed of 1200 rpm under supercharging performed by forming a supercharging pressure Pc of 50 kPa (0.5 bar) by the turbocharger 9. Is shown for each compression ratio CR. As in FIG. 5, CR = 18 (A12), 16.5 (A22), 15 (A32), 13.5 (A42), and 12 (A52) are shown. Here, too, the higher the compression ratio CR, the more the combustion stable region A tends to be formed on the low intake air temperature side. However, under supercharging, compression ignition combustion can be performed with A12 to A32 of CR = 15 or more under atmospheric suction. It has become.
[0012]
FIG. 8 shows the indicated mean effective pressure IMEP calculated for each of the maximum torque points P12 to P52 in the combustion stable regions A12 to A52 in FIG. Even under supercharging, IMEP tends to increase as the compression ratio CR increases, but IMEP is maximum at CR = 15 (P32). This is because when CR = 18 (P12) and CR = 16.5 (P22), the compression ignition combustion under the supercharging of Pc = 50 kPa under the atmospheric suction becomes steep, so that the air-fuel ratio is reduced to avoid knocking. This is because it is necessary to increase the AF. In FIG. 8, IMEP is 500 kPa for P12, 610 kPa for P22, 744 kPa for P32, 663 kPa for P42, and 675 kPa for P52.
In this way, under the supercharging of Pc = 50 kPa, the indicated average effective pressure IMEP becomes maximum at the compression ratio CR = 15, and the efficiency also becomes the best. This is because when the CR is 16.5 and CR, the relative friction increases due to the decrease in the torque, and when the CR is 12 and 13.5, the relative friction increases due to the decrease in the torque, and the wall heat due to the intake air heating increases. This is because the efficiency decreases due to an increase in the loss. (If the CR is 13.5 or less, it is necessary to heat the air-fuel mixture in order to cause compression ignition combustion under atmospheric suction, and the wall heat loss due to this also lowers the efficiency. Will contribute to this.)
[0013]
FIG. 9 shows the supercharging pressure Pc required to cause compression ignition combustion under atmospheric suction when the compression ratio CR is set to 12, 13.5, 15, 16.5 or 18. In the present embodiment, CR is set to 15 and Pc is set to 50 kPa. However, compression ignition combustion can be generated even when CR is 13.5 or 16.5 by adjusting Pc. That is, when CR = 13.5, compression ignition combustion can be caused without heating the air-fuel mixture by increasing Pc to 75 kPa, and when CR = 16.5, the air-fuel ratio can be reduced by reducing Pc to 25 kPa. It is possible to cause compression ignition combustion without lowering the AF. Therefore, the maximum indicated average effective pressure IMEPmax when CR is 13.5 or 16.5 under Pc = 75 kPa and 25 kPa is obtained. As can be seen from FIG. 5, if CR = 18, it is possible to cause compression ignition combustion under natural intake (Pc = 0 kPa), but knocking can be avoided in the case of spark ignition combustion. Instead, the combustion stable region of spark ignition combustion becomes extremely narrow.
[0014]
According to the engine 1 according to the present embodiment, the following effects can be obtained.
First, the compression ratio CR is set to 15, and the turbocharger 9 generates a supercharging pressure Pc of 50 kPa. For this reason, compression ignition combustion can be generated under air suction without heating the air-fuel mixture by means such as forming a minus overlap period between the intake valve 12 and the exhaust valve (FIG. 7). It is possible to improve the efficiency in the low rotation and high load region A of FIG. In the other operating region B in which compression ignition combustion is unlikely to occur, the combustion method is switched to spark ignition combustion, and the variable valve train 13 is used to greatly delay the intake valve closing timing IVC from the bottom dead center. The second time was IVCb. For this reason, spark ignition combustion can be performed without knocking by the Atkinson cycle in which the compression ratio CR is substantially reduced, and a decrease in efficiency due to switching to spark ignition combustion is suppressed as much as possible. As described above, the efficiency of the entire operation region can be improved through the efficiency improvement in the low-rotation high-load region A in which the compression ignition combustion is relatively likely to occur, so that further lower fuel consumption can be realized.
Here, in the spark ignition combustion operation region B, the intake valve closing timing IVC may be advanced as the engine speed increases (FIG. 2). This is because knocking in the case of spark ignition combustion is reduced as the engine speed increases. By doing so, it is possible to achieve both the suppression of knocking and the output at the highest output point.
[0015]
Secondly, by locating the exhaust gas purification device 11 upstream of the turbocharger 9, it is possible to generate a supercharging pressure Pc of 50 kPa even under lean combustion. For example, it is known that when the exhaust gas of the engine 1 contains about 4000 ppm of HC, the exhaust gas generates heat when passing through the exhaust gas purification device 11 and the temperature rises by about 100 ° C. In the case of compression ignition combustion, the air-fuel ratio AF is set to 30 to 40. Therefore, simply flowing exhaust gas from the engine 1 directly into the turbocharger 9 results in insufficient energy and a sufficient supercharging pressure Pc. Cannot be caused. Thus, by causing the exhaust gas to generate heat in the exhaust gas purification device 11 before flowing into the turbocharger 9, the exhaust gas can have sufficient energy and a supercharging pressure Pc of 50 kPa can be generated even under lean combustion. In addition, since the exhaust gas purification device 11 is located upstream of the turbocharger 9, the pressure in the exhaust pipe 10 downstream of the exhaust gas purification catalyst 11 becomes high, so that the temperature of the catalyst is kept high and the oxidation reaction is promoted. The effect is also obtained.
[0016]
Next, a hybrid vehicle to which the engine 1 is applied will be described.
FIG. 10 is a configuration diagram of a drive system of a hybrid vehicle V as one example. The hybrid vehicle V includes an engine 1 as a drive source, and two motor generators 51 and 52 which also function as a motor and a generator. The crankshaft of the engine 1 and the rotating shaft of one motor generator 51 are directly connected, and this motor generator 51 and the other motor generator 52 are connected via an electromagnetic clutch 53. The torque output from these drive sources is transmitted to the axle 55, the differential gear 56, and the wheel drive shafts 57, 57 via the transmission 54, so that the drive wheels 58, 58 are rotated. Motor generators 51 and 52 are connected to battery 59 via inverter 60. The hybrid vehicle V having such a drive system can travel by switching modes in a motor traveling mode, a series hybrid traveling mode, and a parallel hybrid traveling mode, which will be described below, according to the driving region.
1) Motor running mode The clutch 53 is disconnected, the mechanical connection between the engine 1 and the axle 55 is disconnected, and the motor generator 52 is operated by the electric power from the battery 59 to run. When the charge amount of the battery 59 falls below a certain value, the engine 1 operates the motor generator 51 as a generator, and the obtained power can be used for charging the battery 59. In the motor running mode, it is generally unnecessary to operate the inefficient engine 1 in a low output range.
2) Series hybrid traveling mode The vehicle travels only by the motor generator 52 with the clutch 53 disconnected, but the electric power of the motor generator 52 is generated by the engine 1 operating the motor generator 51 as a generator. Here, the load on motor generator 51 is changed according to the power required by motor generator 52. When such real-time power generation is not used, the engine 1 can be operated in the most efficient region.
3) Parallel / hybrid running mode The engine 1 and the axle 55 are connected by coupling the clutch 53, and the vehicle runs by the engine 1 and assists the insufficient torque by the motor generator 52. By adopting the parallel hybrid drive mode, the maximum output of the engine 1 can be reduced, so that fuel efficiency can be improved through downsizing of the engine 1.
[0017]
FIG. 11 shows the frequency distribution of the engine output in an automobile using only the engine as a drive source, and FIG. 12 shows the frequency distribution of the engine output in a hybrid vehicle V having an engine 1 and motor generators 51 and 52 as a drive source. Are shown in comparison with fuel efficiency. In the hybrid vehicle V, the motor driving mode is set in a low output range where the efficiency of the engine 1 is low and the mode is appropriately switched to the series hybrid driving mode in accordance with the driving conditions. The output Lp2 (FIG. 12) is formed on the lowest fuel consumption point side as compared with Lp1 (FIG. 11). In the hybrid vehicle V having a displacement of 2000 cc, the frequency of occurrence of engine output is concentrated in the range of 10 to 20 kW.
[0018]
FIG. 13 is a diagram in which 10 kW and 20 kW equal output lines are added to the operation area of the engine 1 of FIG. 2 on the assumption that the displacement is in the 2000 cc class. As described above, the compression ignition combustion operation region A is included in the range of 10 to 20 kW where the engine output of the hybrid vehicle V is concentrated. Therefore, the operating condition of the engine 1 required for the hybrid vehicle V can be almost satisfied by the operation by the compression ignition combustion.
[0019]
As described above, in the hybrid vehicle V, the motor running mode in which the engine 1 is stopped in the low output range and the vehicle runs by the motor generator 52 is set, so that the engine 1 does not need to be operated in the inefficient area, and Since the entire required output range can be substantially covered by the compression ignition combustion, further lower fuel consumption can be realized.
[0020]
In the above description, an example in which the present invention is applied to the port injection type engine 1 in which fuel is added to the air before being taken into the cylinder is described. However, the present invention is not limited to this. It is also possible to apply the present invention to a direct injection type engine in which fuel is injected directly into a cylinder by installing the fuel injection valve in a cylinder.
Further, as the supercharger, a mechanically driven supercharger or the like operated by means other than the exhaust turbine may be employed. A clutch is installed between the supercharger and the crankshaft, and the clutch is operated in accordance with the combustion method to connect or release the supercharger and the crankshaft. According to the supercharger, the supercharging pressure Pc can be quickly formed regardless of the temperature of the exhaust gas.
[Brief description of the drawings]
FIG. 1 is a cross-sectional view of an automobile engine according to an embodiment of the present invention. FIG. 2 is a distribution diagram of a combustion method according to operating conditions of the engine. FIG. 3 is an operation characteristic of an intake valve and an exhaust valve of the engine. Fig. 4 Combustion stable region and maximum torque point of the same engine. Fig. 5 Combustion stable region for each compression ratio under natural intake. Fig. 6 Indicated average effective pressure under natural intake calculated for maximum torque point. FIG. 7: Combustion stable region for each compression ratio under supercharging. FIG. 8: Indicated average effective pressure under supercharging calculated for the maximum torque point. FIG. 9: To generate compression ignition combustion under atmospheric suction. FIG. 10 Drive system of a hybrid vehicle equipped with an engine according to an embodiment of the present invention FIG. 11 Distribution of occurrence frequency of engine output in a vehicle using only an engine as a drive source 12 High For the entire operating range of the occurrence frequency distribution 13 Engine Engine output the lid automobiles, [REFERENCE NUMERALS] concentration range of the engine output frequency in the hybrid vehicle
DESCRIPTION OF SYMBOLS 1 ... Automotive engine as a compression ignition type internal combustion engine, 2 ... Cylinder block, 3 ... Cylinder head, 4 ... Piston, 5 ... Combustion chamber, 7a, 7b ... Intake pipe, 8 ... Injector, 9 ... As supercharger Turbocharger, 91: Turbine unit of turbocharger, 911: Wastegate of turbocharger, 10: Exhaust pipe, 11: Exhaust gas purification device, 12: Intake valve, 13: Variable valve operating device as variable valve operating means, 131 ... Intake cam, 15: ignition plug, 101: engine controller as combustion control means, 51, 52: motor generator as electric motor, 53: electromagnetic clutch as clutch, 54: transmission, 56: differential gear, 58: drive Wheel, 59 ... battery, 60 ... inverter.

Claims (8)

A compression ignition type internal combustion engine in which the combustion method is switched between compression ignition combustion and spark ignition combustion,
Equipped with a supercharger,
While the compression ratio is set to 13.5 to 16.5,
A compression ignition type internal combustion engine in which the supercharging pressure is set to 25 to 75 kPa. Equipped with a supercharger,
While the compression ratio is set to 13.5 to 16.5,
A compression ignition type internal combustion engine set to a supercharging pressure within a range of 25 to 75 kPa, at which the indicated average effective pressure obtained under the set compression ratio becomes substantially maximum. 3. The compression ignition type internal combustion engine according to claim 2, wherein the compression ratio is set to approximately 15, and the supercharging pressure is set to approximately 50 kPa. 4. The compression ignition type internal combustion engine according to claim 2, wherein a turbocharger is provided as a supercharger, and a catalyst having an oxidation function is disposed upstream of the turbocharger in an exhaust passage. The compression ignition type internal combustion engine according to claim 2 or 3, further comprising a supercharger as a supercharger. Combustion control means for switching the combustion mode between compression ignition combustion and spark ignition combustion according to the engine operating conditions;
Variable valve operating means for opening and closing the intake valve with a plurality of valve closing timings different with respect to the crank angle,
The combustion control means stops the supercharging by the supercharger when spark ignition combustion is performed, and at least at the time of low rotation, lowers the closing timing of the intake valve by the variable valve means at a lower dead center than when compression ignition combustion is performed. The compression ignition type internal combustion engine according to any one of claims 2 to 5, which is set at a predetermined time later. 7. The compression ignition type internal combustion engine according to claim 6, wherein in the case of spark ignition combustion, the intake valve closing timing by the variable valve means is changed to the advanced side as the engine speed increases. A compression ignition internal combustion engine according to any one of claims 1 to 7, and an electric motor provided in a drive source,
A hybrid vehicle having a traveling mode set in a low output range, in which the internal combustion engine is stopped and the vehicle travels with an output from an electric motor.

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