JP4137704B2 - Control device for internal combustion engine - Google Patents

Description

[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a control device for an internal combustion engine that controls a boost pressure of intake air via a supercharger and variably controls a valve closing timing with respect to a valve opening timing of the intake valve via a valve timing variable device. .
[0002]
[Prior art]
2. Description of the Related Art Conventionally, as a control device for an internal combustion engine that controls the supercharging pressure of intake air via a supercharger, for example, one described in Patent Document 1 is known. This internal combustion engine is provided with a turbocharger and a knock sensor as a supercharger. In this turbocharger, the flow rate of the exhaust gas to the turbine blade changes with the change in the opening degree of the wastegate valve, thereby changing the supercharging pressure. The knock sensor is for detecting knocking in the internal combustion engine, and outputs a pulse signal indicating the knocking to the control device when knocking occurs.
[0003]
The control device counts the number of occurrences of knocking during one combustion cycle based on the pulse signal from the knock sensor, sets the ignition timing retard value according to the number of occurrences, and knocks during four combustion cycles. The duty ratio of the drive signal for the waste gate valve is set according to the number of occurrences of the above. More specifically, the retard value of the ignition timing is set to a larger value as the number of occurrences of knocking in one combustion cycle is larger. That is, the ignition timing is controlled to be later. Further, the duty ratio of the drive signal for the wastegate valve is set to a smaller value as the number of occurrences of knocking in the four combustion cycles is larger. That is, the supercharging pressure is controlled to be a lower value. As described above, when knocking occurs, the ignition timing is retarded and the supercharging pressure is reduced, thereby suppressing the occurrence of knocking.
[0004]
[Patent Document 1]
Japanese Examined Patent Publication No. 3-37034 (pages 3-5, Fig. 3)
[0005]
[Problems to be solved by the invention]
According to the control device of Patent Document 1 described above, when knocking occurs, the ignition timing is retarded and the supercharging pressure is reduced, so that, for example, the occurrence of knocking in a high load range can be suppressed. By executing such control, both the combustion efficiency and the engine output are lowered, and as a result, the drivability and the merchantability are lowered.
[0006]
The present invention has been made to solve the above-described problems, and can improve both combustion efficiency and engine output while suppressing the occurrence of knocking in a high load range, thereby improving drivability and products. An object of the present invention is to provide a control device for an internal combustion engine that can improve the performance.
[0007]
[Means for Solving the Problems]
In order to achieve this object, the invention according to claim 1 controls the supercharging pressure Pc of the intake air via the supercharger (turbocharger device 10) provided in the intake passage (intake pipe 8), and A control device 1 for an internal combustion engine 3 that variably controls a valve closing timing with respect to a valve opening timing of an intake valve 6 via a valve timing variable device (variable intake valve driving device 40). Load detection means (ECU 2, crank angle sensor 20, accelerator opening sensor 35, step 11) for detecting (required drive torque TRQ_eng), and a target for boost pressure control according to the detected load of the internal combustion engine 3 Target supercharging pressure setting means (ECU2, step 46) for setting the target supercharging pressure Pc_cmd, and supercharging control means (EC for controlling the supercharger in accordance with the set target supercharging pressure) 2, step 47), and when the detected load of the internal combustion engine 3 is in a predetermined high load range (TRQ2 <TRQ_eng <TRQ4) higher than a predetermined load (predetermined value TRQ2), the valve closing timing control of the intake valve 6 The target valve closing timing (basic value θmsi_base of the target sub-intake cam phase) becomes larger as the expansion ratio in the combustion cycle exceeds the compression ratio and the detected load of the internal combustion engine 3 is higher. Target valve closing timing setting means (ECU2, first SPAS controller 221, step for setting timing close to the ratio) 78 ) And valve timing control means (ECU2, step for controlling the valve timing variable device (variable intake valve driving device 40) according to the set target valve closing timing) 74, 79 ).
[0008]
According to this control device for an internal combustion engine, the supercharger is controlled in accordance with the target boost pressure that is the target of the boost pressure control, and the variable valve timing device is the target for the closing timing control of the intake valve. It is controlled according to the target valve closing timing. At that time, when the detected load of the internal combustion engine is in a predetermined high load range higher than the predetermined load, the target valve closing timing is such that the expansion ratio in the combustion cycle exceeds the compression ratio and the detected internal combustion engine The timing is set such that the higher the load, the closer the expansion ratio approaches the compression ratio. That is, in a predetermined high load region, the effective compression volume is controlled to increase as the load on the internal combustion engine increases. Therefore, even when the load on the internal combustion engine is high, the required engine output is ensured. An increase in the supercharging pressure can be suppressed, whereby an increase in intake air temperature can be suppressed. Therefore, the limit at which knocking begins to occur can be expanded without performing retard control of the ignition timing in a predetermined high load region, and both the combustion efficiency and the engine output can be improved. As a result, drivability and merchantability can be improved.
[0009]
According to a second aspect of the present invention, in the control device 1 for the internal combustion engine 3 according to the first aspect, the target valve closing timing setting means is configured such that the load of the internal combustion engine 3 falls within a predetermined high load range (TRQ2 <TRQ_eng <TRQ4). In some cases, the closing timing of the intake valve 6 is set to a timing earlier than a predetermined timing at which the expansion ratio in the combustion cycle of the internal combustion engine 3 becomes equal to the compression ratio.
[0010]
Generally, as a technique for realizing a so-called high expansion ratio cycle operation in which the expansion ratio in the combustion cycle exceeds the compression ratio, the closing timing of the intake valve is set to the closing timing of the so-called Otto cycle in which the expansion ratio is equal to the compression ratio Hereinafter, there is a delayed closing method that is set later than “Otto valve closing timing” and an early closing method that is set earlier than this Otto valve closing timing. In this case, in the former method, the fuel is blown back into the intake manifold, the fuel staying in the intake manifold is increased, and the fuel adhesion to the inner wall of the intake manifold is increased. As a result, control accuracy such as air-fuel ratio control and torque control, particularly control accuracy in a transient operation state, is lowered. More specifically, the air-fuel ratio of the air-fuel mixture shifts to the rich side, which may cause an unnecessary increase in torque and an increase in unburned HC in the exhaust gas. In addition, the blown back fuel adheres to a device such as an intake valve in a carbonized state, which may shorten the life of the intake system device. The above problems become more prominent in the high load range. On the other hand, according to the control device for an internal combustion engine, when the load of the internal combustion engine is in a predetermined high load region, the closing timing of the intake valve is such that the expansion ratio during the combustion cycle of the internal combustion engine is the compression ratio. It is set at a timing earlier than the predetermined timing (that is, the Otto valve closing timing). In other words, since the high-expansion ratio cycle operation is realized by the latter early closing method, the above-described problem due to the former late closing method does not occur. As a result, it is possible to improve control accuracy such as air-fuel ratio control and torque control in a high load range as compared with the former slow closing method, and in particular, control accuracy in a transient operation state can be remarkably improved. . As a result, exhaust gas characteristics and operability can be improved, and the life of the intake system device can be extended.
[0011]
According to a third aspect of the present invention, in the control device 1 for the internal combustion engine 3 according to the first or second aspect, the target valve closing timing setting means has a predetermined low level in which the load of the internal combustion engine 3 is lower than a predetermined high load range. When in the load range (TRQ_idle ≦ TRQ_eng <TRQott), the closing timing of the intake valve 6 is set to a timing at which the expansion ratio in the combustion cycle of the internal combustion engine 3 exceeds the compression ratio.
[0012]
According to the control device for an internal combustion engine, when the load of the internal combustion engine is in a predetermined low load region lower than a predetermined high load region, the closing timing of the intake valve is determined by the expansion ratio during the combustion cycle of the internal combustion engine. It is set at a timing exceeding the compression ratio and is operated in a high expansion ratio cycle. This eliminates the need to throttle the amount of intake air with a throttle valve, for example, in a low load range, so that the amount of intake air can be set to an appropriate value according to the low load while avoiding pumping loss. Can be improved.
[0013]
According to a fourth aspect of the present invention, in the control device 1 for the internal combustion engine 3 according to the third aspect, the target valve closing timing setting means is configured such that the load of the internal combustion engine 3 falls within a predetermined low load range (TRQ_idle ≦ TRQ_eng <TRQott) In some cases, the closing timing of the intake valve 6 is set to a timing later than a predetermined timing at which the expansion ratio in the combustion cycle of the internal combustion engine 3 becomes equal to the compression ratio.
[0014]
As described above, as a method for realizing the high expansion ratio cycle operation, there are a method of setting the intake valve closing timing later than the Otto valve closing timing, and a method of setting earlier than the Otto valve closing timing, the latter In the case of this method, when the intake air temperature and the engine temperature are low, the liquefaction of the fuel occurs due to a decrease in the temperature in the cylinder due to the adiabatic expansion of the air-fuel mixture in the cylinder at a low load range, and the combustion state is not good. May become stable. On the other hand, according to the control apparatus for an internal combustion engine, the high expansion ratio cycle operation is realized by setting the closing timing of the intake valve to a timing later than the Otto valve closing timing. In the load range, the liquefaction of the fuel as described above can be avoided, thereby stabilizing the combustion state.
[0015]
According to a fifth aspect of the present invention, in the control device for the internal combustion engine 3 according to any one of the first to fourth aspects, the target boost pressure setting means is configured such that the load of the internal combustion engine 3 is a predetermined high load range (TRQ3 <TRQ_eng). When <TRQ4), the target boost pressure Pc_cmd is set to a smaller value as the load on the internal combustion engine 3 is larger.
[0016]
According to the control device for an internal combustion engine, when the load on the internal combustion engine is in a predetermined high load range, the target boost pressure is set to a smaller value as the load on the internal combustion engine is larger. The degree of increase in intake air temperature due to supercharging can be further reduced. Therefore, by setting the predetermined high load range to a load range where knocking occurs, the limit at which knocking starts can be further expanded without performing retard control of the ignition timing.
[0017]
According to a sixth aspect of the present invention, in the control device 1 for the internal combustion engine 3 according to any one of the first to fifth aspects, the variable valve timing device is a hydraulically driven variable valve timing device driven by the supply of hydraulic pressure. The valve timing control means is configured to control the hydraulic pressure Psd supplied to the hydraulically driven variable valve timing device.
[0018]
According to this control device for an internal combustion engine, the variable valve timing device is constituted by a hydraulic drive type driven by the supply of hydraulic pressure, so for example, the valve body of the intake valve is driven by the electromagnetic force of a solenoid. Compared with the case where a variable valve timing device is used, the intake valve can be reliably opened and closed even in a higher load range, power consumption can be reduced, and operating noise of the intake valve can be reduced.
[0019]
According to a seventh aspect of the present invention, in the control device 1 for the internal combustion engine 3 according to any one of the first to sixth aspects, the variable valve timing device is configured such that the valve lift amount of the intake valve 6 can be changed. It is characterized by.
[0020]
According to the control device for the internal combustion engine, the variable valve timing device is configured to be able to change the valve lift amount of the intake valve, so that, for example, the lift amount flows into the combustion chamber by controlling the lift amount to a smaller value side. The in-cylinder flow can be further increased by increasing the flow rate of the intake air. Thereby, combustion efficiency can be improved. In particular, since the combustion can be accelerated by increasing the in-cylinder flow, in the case of the invention according to claim 3, the high expansion ratio cycle operation is performed when the intake valve is closed later than the Otto valve closing timing. Even when it is realized at the valve timing, the liquefaction of the fuel as described above can be avoided, so that the combustion state can be stabilized.
[0021]
According to an eighth aspect of the present invention, in the control device 1 for the internal combustion engine 3 according to the seventh aspect, the valve timing variable device is rotatably supported by a rotation fulcrum (pin 51c), and the intake valve 6 is rotated by the rotation. Between the intake rocker arm 51 that opens and closes, the first and second intake camshafts (main and auxiliary intake camshafts 41 and 42) that rotate at the same rotational speed, and the first and second intake camshafts. An intake cam phase variable mechanism (sub intake cam phase variable mechanism 70) for changing the phase (sub intake cam phase θmsi) and a first intake cam shaft (main intake cam shaft 41) are provided to rotate the first intake cam shaft. The first intake cam (main intake cam 43) that rotates the intake rocker arm 51 around the rotation fulcrum by rotating with the second intake camshaft (sub intake camshaft 4). ) To provided, by rotating with the rotation of the second intake camshaft, a second intake cam for moving the pivot point of the intake rocker arm 51 (auxiliary intake cam 44), characterized in that it comprises a.
[0022]
According to the control device for an internal combustion engine, in the variable valve timing device, the intake cam rocker arm is rotated around the rotation fulcrum by rotating the first cam with the rotation of the first cam shaft. The valve is driven to open and close. At this time, since the second cam rotates with the rotation of the second cam shaft, the pivot point of the intake rocker arm is moved, so that the valve lift amount of the intake valve can be freely changed. Since the relative phase between the first and second intake camshafts is changed by the intake cam phase variable mechanism, both the valve closing timing and the valve lift amount of the intake valve can be freely changed. That is, by using two cams, two camshafts, and a cam phase variable mechanism, it is possible to realize a valve timing variable device that can freely change the valve closing timing and valve lift amount of the intake valve.
[0023]
According to a ninth aspect of the present invention, in the control device 1 for the internal combustion engine 3 according to any one of the first to eighth aspects, the internal combustion engine 3 injects fuel and supplies the first fuel injection valve (mainly supplied into the cylinder). A fuel injection valve 4) and an intake air cooling device (fuel evaporative cooling device 12) that is provided downstream of the supercharger in the intake passage and cools the intake air from the supercharger. The evaporative cooling device 12) injects fuel toward the lipophilic film plate 14 having a lipophilic film having an affinity for fuel on its surface, and supplies the fuel into the cylinders # 1 to # 4. And a second fuel injection valve (sub fuel injection valve 15).
[0024]
According to this control device for an internal combustion engine, fuel is injected from the first fuel injection valve and is also injected from the second fuel injection valve of the intake air cooling device toward the lipophilic film plate and supplied into the cylinder of the internal combustion engine. Is done. At that time, since the lipophilic film having affinity for the fuel is formed on the surface of the lipophilic film plate, the fuel is thinned on the lipophilic film and is heated by the supercharging operation by the supercharger. Vaporizes due to the rising heat of intake air. As a result, an air-fuel mixture is generated and the intake air is cooled by the heat of vaporization at that time. In this way, it is possible to obtain an intake air cooling action with the generation of the air-fuel mixture, thereby further expanding the limit at which knocking starts without performing ignition timing retard control. .
[0025]
According to a tenth aspect of the present invention, in the control device 1 for the internal combustion engine 3 according to the ninth aspect, the first fuel injection out of the fuel amount (total fuel injection amount TOUT) to be supplied to the cylinders # 1 to # 4. The first fuel injection ratio (main fuel injection ratio Rt_Pre) to be supplied by the valve and the second fuel injection ratio (100−Rt_Pre) to be supplied by the second fuel injection valve are respectively set according to the load of the internal combustion engine 3. The fuel injection ratio setting means (ECU2, step 17) is further provided for setting the second fuel injection ratio (100-Rt_Pre) to a larger value as the load on the internal combustion engine 3 is larger.
[0026]
In general, as the load on the internal combustion engine increases, the boost pressure is increased to a higher value, and the degree of increase in intake air temperature increases. On the other hand, according to the control device for an internal combustion engine, the second fuel injection ratio to be supplied from the second fuel injection valve is set to a larger value as the load of the internal combustion engine is larger. The above-described intake air cooling action by the intake air cooling device can be obtained more effectively and appropriately in accordance with the degree of increase of.
[0027]
DETAILED DESCRIPTION OF THE INVENTION
Hereinafter, an internal combustion engine control apparatus according to an embodiment of the present invention will be described with reference to the drawings. 1 and 2 show a schematic configuration of an internal combustion engine (hereinafter referred to as “engine”) 3 to which the control device 1 of the present embodiment is applied, and FIG. 3 shows a schematic configuration of the control device 1. As shown in FIG. 3, the control device 1 includes an ECU 2, which performs valve timing control and supercharging pressure control of the intake valve 6, as will be described later, according to the operating state of the engine 3. Various control processes are executed.
[0028]
The engine 3 is an in-line four-cylinder gasoline engine mounted on a vehicle (not shown) and includes first to fourth four cylinders # 1 to # 4 (see FIG. 5). In addition, the engine 3 is provided with a main fuel injection valve 4 (first fuel injection valve) and a spark plug 5 for each cylinder (only one of them is shown). 5 is attached to the cylinder head 3a. Each main fuel injection valve 4 is connected to the ECU 2, and its fuel injection amount and fuel injection timing are controlled by a control input from the ECU 2, whereby the fuel is directly injected into the combustion chamber of the corresponding cylinder.
[0029]
Each spark plug 5 is also connected to the ECU 2 and is discharged when a high voltage is applied from the ECU 2 at a timing corresponding to the ignition timing, thereby burning the air-fuel mixture in the combustion chamber.
[0030]
Further, the engine 3 is provided for each cylinder, and an intake valve 6 and an exhaust valve 7 that open and close the intake port and the exhaust port, respectively, and a variable type that opens and closes the intake valve 6 and simultaneously changes its valve timing and valve lift amount. An intake valve driving device 40 and a variable exhaust valve driving device 90 that changes the valve timing and valve lift amount at the same time as opening and closing the exhaust valve 7 are provided. Details of these variable intake drive unit 40 and variable exhaust valve drive unit 90 will be described later. The intake valve 6 and the exhaust valve 7 are urged in the valve closing direction by valve springs 6a and 7a, respectively.
[0031]
On the other hand, a magnet rotor 20 a is attached to the crankshaft 3 b of the engine 3. The magnet rotor 20a, together with the MRE pickup 20b, constitutes a crank angle sensor 20 (load detection means). The crank angle sensor 20 outputs a CRK signal and a TDC signal, both of which are pulse signals, to the ECU 2 as the crankshaft 3b rotates.
[0032]
One pulse of the CRK signal is output every predetermined crank angle (for example, 30 deg). The ECU 2 calculates the engine speed (hereinafter referred to as “engine speed”) NE of the engine 3 in accordance with the CRK signal. The TDC signal is a signal indicating that the piston 3c of each cylinder is at a predetermined crank angle position slightly before the TDC position of the intake stroke, and is every predetermined crank angle (180 deg in the example of the present embodiment). One pulse is output at the same time.
[0033]
An intake pipe 8 (intake passage) of the engine 3 is provided with a turbocharger device 10, an intercooler 11, a fuel evaporative cooling device 12, a throttle valve mechanism 16, and the like in order from the upstream side.
[0034]
The turbocharger device 10 (supercharger) includes a compressor blade 10a housed in a compressor housing in the middle of an intake pipe 8, a turbine blade 10b housed in a turbine housing in the middle of an exhaust pipe 9, and two blades A shaft 10c that integrally connects 10a and 10b, a wastegate valve 10d, and the like are provided.
[0035]
In the turbocharger 10, when the turbine blade 10 b is rotationally driven by the exhaust gas in the exhaust pipe 9, the compressor blade 10 a integrated therewith is also rotated at the same time, so that the intake air in the intake pipe 8 is pressurized. The That is, the supercharging operation is executed.
[0036]
The waste gate valve 10d opens and closes a bypass exhaust passage 9a that bypasses the turbine blade 10b of the exhaust pipe 9, and is composed of an electromagnetic control valve connected to the ECU 2 (see FIG. 3). The waste gate valve 10d changes the flow rate of the exhaust gas flowing through the bypass exhaust passage 9a, in other words, the flow rate of the exhaust gas that drives the turbine blade 10b, by changing the opening according to the control input Dut_wg from the ECU 2. . Thereby, the supercharging pressure Pc by the turbocharger device 10 is controlled.
[0037]
On the other hand, an air flow sensor 21 is provided on the upstream side of the compressor blade 10 a of the intake pipe 8. The air flow sensor 21 is constituted by a hot-wire air flow meter, and outputs a detection signal representing an intake air amount (hereinafter referred to as “TH passage intake air amount”) Gth passing through a throttle valve 17 described later to the ECU 2.
[0038]
The intercooler 11 is a water-cooled type, and cools the intake air whose temperature has risen due to the supercharging operation (pressurizing operation) in the turbocharger device 10 when the intake air passes through the intercooler 11.
[0039]
Further, a supercharging pressure sensor 22 is provided between the intercooler 11 of the intake pipe 8 and the fuel evaporative cooling device 12. The supercharging pressure sensor 22 is composed of a semiconductor pressure sensor or the like, and outputs a detection signal representing the intake pressure in the intake pipe 8 pressurized by the turbocharger device 10, that is, a supercharging pressure Pc (absolute pressure), to the ECU 2. To do.
[0040]
On the other hand, the fuel evaporative cooling device 12 (intake air cooling device) evaporates the fuel and generates an air-fuel mixture, and at the same time, lowers the temperature of the intake air. As shown in FIG. A housing 13 provided in the middle, a number of lipophilic film plates 14 (only six shown) accommodated in the housing 13 in parallel with each other at a predetermined interval, and a sub fuel injection valve 15 (second fuel) Injection valve).
[0041]
The sub fuel injection valve 15 is connected to the ECU 2, and its fuel injection amount and fuel injection timing are controlled by a control input from the ECU 2, thereby injecting fuel toward a large number of lipophilic film plates 14. . As will be described later, the total fuel injection amount TOUT to be injected from both the auxiliary fuel injection valve 15 and the main fuel injection valve 4 is determined by the ECU 2 according to the operating state of the engine 3, and the total fuel injection amount TOUT is determined. The ratio of the fuel injection amount from the main fuel injection valve 4 (main fuel injection rate Rt_Pre, which will be described later) and the ratio of the fuel injection amount from the sub fuel injection valve 15 to the fuel injection amount TOUT in the operating state of the engine 3 Will be decided accordingly. A lipophilic film having affinity for fuel is formed on the surface of the lipophilic film plate 14.
[0042]
With the above configuration, in the fuel vaporization cooling device 12, the fuel injected from the auxiliary fuel injection valve 15 is thinned by the lipophilicity on the surface of each lipophilic film plate 14, and then vaporized by the heat of the intake air. Thereby, the air-fuel mixture is generated and the intake air is cooled by the heat of vaporization at that time. Due to the cooling effect of the fuel vaporization cooling device 12, the charging efficiency can be increased, and the limit of occurrence of knocking of the engine 3 can be expanded. For example, when the engine 3 is operating at a high load, the limit ignition timing at which knocking starts to occur can be extended to the advance side by a predetermined crank angle (for example, 2 deg), thereby improving combustion efficiency. it can.
[0043]
The throttle valve mechanism 16 described above includes a throttle valve 17 and a TH actuator 18 that opens and closes the throttle valve 17. The throttle valve 17 is rotatably provided in the middle of the intake pipe 8, and changes the TH passing intake air amount Gth by the change of the opening degree accompanying the rotation. The TH actuator 18 is a combination of a motor connected to the ECU 2 and a gear mechanism (both not shown). The TH actuator 18 is controlled by a control input DUTY_th (described later) from the ECU 2, thereby opening the throttle valve 17. Change.
[0044]
The throttle valve 17 is provided with two springs (both not shown) for biasing the throttle valve 17 in the valve opening direction and the valve closing direction, respectively, and the throttle valve 17 is biased by the biasing force of these two springs. Is maintained at a predetermined initial opening TH_def when the control input DUTY_th is not input to the TH actuator 18. The initial opening TH_def is set to a value (for example, 7 °) that is close to the fully closed state and that can secure the intake air amount necessary for starting the engine 3.
[0045]
Further, a throttle valve opening sensor 23 composed of, for example, a potentiometer is provided in the vicinity of the throttle valve 17 of the intake pipe 8. The throttle valve opening sensor 23 outputs a detection signal representing the actual opening (hereinafter referred to as “throttle valve opening”) TH of the throttle valve 17 to the ECU 2.
[0046]
A portion of the intake pipe 8 downstream of the throttle valve 17 is a surge tank 8a, and an intake pipe absolute pressure sensor 24 is provided in the surge tank 8a. The intake pipe absolute pressure sensor 24 is composed of, for example, a semiconductor pressure sensor, and outputs a detection signal representing the absolute pressure (hereinafter referred to as “intake pipe absolute pressure”) PBA in the intake pipe 8 to the ECU 2.
[0047]
On the other hand, on the downstream side of the turbine blade 10b of the exhaust pipe 9, first and second catalytic devices 19a and 19b are provided in order from the upstream side. By these catalytic devices 19a and 19b, NOx, HC, CO and the like are purified.
[0048]
An oxygen concentration sensor (hereinafter referred to as “O2 sensor”) 26 is provided between the first and second catalytic devices 19a and 19b. The O2 sensor 26 is composed of zirconia, a platinum electrode, and the like, and outputs a detection signal based on the oxygen concentration in the exhaust gas downstream of the first catalyst device 19a to the ECU 2.
[0049]
Further, a LAF sensor 25 is provided between the turbine blade 10 b of the exhaust pipe 9 and the first catalyst device 19. The LAF sensor 25 is configured by combining a sensor similar to the O2 sensor 26 and a detection circuit such as a linearizer. The LAF sensor 25 controls the oxygen concentration in the exhaust gas in a wide range of air-fuel ratios from the rich region to the lean region. A detection signal that is linearly detected and proportional to the oxygen concentration is output to the ECU 2. The ECU 2 executes air-fuel ratio control based on the detection signals of the LAF sensor 25 and the O2 sensor 26.
[0050]
Next, the above-described variable intake valve drive device 40 (valve timing variable device) will be described. As shown in FIGS. 2, 5 and 6, this variable intake valve drive device 40 is provided with a main intake camshaft 41 and an auxiliary intake camshaft 42 for driving the intake valves, and is provided for each cylinder. An intake valve drive mechanism 50 (only one is shown) that drives the intake valve 6 to open and close as the intake camshafts 41 and 42 rotate, a main intake cam phase variable mechanism 60, a sub intake cam phase variable mechanism 70, and 3 Two intake cam phase varying mechanisms 80 are provided.
[0051]
The main intake camshaft 41 (first intake camshaft) is rotatably attached to the cylinder head 3a and extends along the cylinder arrangement direction. The main intake camshaft 41 includes a main intake cam 43 (first intake cam) provided for each cylinder, a sprocket 47 provided at one end, and a main intake cam 43 and a sprocket 47 for the first cylinder # 1. And a main gear 45 provided therebetween. The main intake cam 43, the main gear 45, and the sprocket 47 are all attached to the main intake cam shaft 41 so as to rotate coaxially and integrally. The sprocket 47 is connected to the crankshaft 3b via the timing chain 48, whereby the main intake camshaft 41 rotates clockwise (indicated by an arrow Y1) every time the crankshaft 3b rotates twice. Direction).
[0052]
The main intake cam phase varying mechanism 60 is provided at the end of the main intake camshaft 41 on the sprocket 47 side. The main intake cam phase variable mechanism 60 has a relative phase of the main intake camshaft 41 to the sprocket 47, that is, a relative phase of the main intake camshaft 41 to the crankshaft 3b (hereinafter referred to as “main intake cam phase”) θmi. Is steplessly changed to the advance side or the retard side, details of which will be described later.
[0053]
Further, a main intake cam angle sensor 27 is provided at the end of the main intake camshaft 41 opposite to the sprocket 47. Like the crank angle sensor 20, the main intake cam angle sensor 27 is composed of a magnet rotor and an MRE pickup, and a main intake cam signal, which is a pulse signal, is transmitted to a predetermined cam as the main intake cam shaft 41 rotates. It outputs to ECU2 for every angle | corner (for example, 1 deg). The ECU 2 calculates (detects) the main intake cam phase θmi based on the main intake cam signal and the CRK signal.
[0054]
On the other hand, the auxiliary intake camshaft 42 (second intake camshaft) is also rotatably supported by the cylinder head 3 a similarly to the main intake camshaft 41, and extends parallel to the main intake camshaft 41. The auxiliary intake camshaft 42 includes an auxiliary intake cam 44 (second intake cam) provided for each cylinder, and an auxiliary gear 46 having the same number of teeth and the same diameter as the main gear 45. 46 rotates integrally with the auxiliary intake camshaft 42 coaxially.
[0055]
Both the main gear 45 and the sub gear 46 are pressed so as to always mesh with each other by a pressing spring (not shown), and backlash is not generated by a backlash compensation mechanism (not shown). Due to the meshing of both gears 45 and 46, the auxiliary intake camshaft 42 rotates counterclockwise in FIG. 6 (the direction indicated by the arrow Y2) at the same rotational speed as the main intake camshaft 41 rotates clockwise. .
[0056]
The auxiliary intake cam phase variable mechanism 70 (intake cam phase variable mechanism) is provided at the end of the auxiliary intake camshaft 42 on the timing chain 48 side, and the auxiliary intake camshaft 42 is relative to the main intake camshaft 41. In other words, the relative phase of the auxiliary intake cam 44 for the first cylinder # 1 with respect to the main intake cam 43 (hereinafter referred to as “sub intake cam phase”) θmsi is changed steplessly. Details of the auxiliary intake cam phase varying mechanism 70 will be described later.
[0057]
Further, a sub intake cam angle sensor 28 is provided at the end of the sub intake cam shaft 42 opposite to the sub intake cam phase varying mechanism 70. Similarly to the main intake cam angle sensor 27, the auxiliary intake cam angle sensor 28 is also composed of a magnet rotor and an MRE pickup. Along with the rotation of the auxiliary intake cam shaft 42, an auxiliary intake cam signal that is a pulse signal is predetermined. Is output to the ECU 2 at every cam angle (for example, 1 deg). The ECU 2 calculates the auxiliary intake cam phase θmsi based on the auxiliary intake cam signal, the main intake cam signal, and the CRK signal.
[0058]
Further, in the four auxiliary intake cams 44, the auxiliary intake cam 44 for the first cylinder # 1 is attached so as to rotate integrally with the auxiliary intake camshaft 42, and the other second to fourth cylinders # are provided. Each of the auxiliary intake cams 44 for # 2 to # 4 is connected to the auxiliary intake camshaft 42 via the intake cam phase varying mechanism 80. These intake cam inter-phase variable mechanisms 80 have a relative phase of the auxiliary intake cam 44 for the second to fourth cylinders # 2 to # 4 relative to the auxiliary intake cam 44 for the first cylinder # 1 (hereinafter referred to as “intake air”). Θssi # i (referred to as “inter-cam phase”) is changed steplessly independently of each other, and details thereof will be described later. Note that the symbol #i in the intake cam phase θssi # i represents a cylinder number and is set to # i = # 2 to # 4. This also applies to the following description.
[0059]
Furthermore, three # 2 to # 4 auxiliary intake cam angle sensors 29 to 31 are electrically connected to the ECU 2 (see FIG. 3). These # 2 to # 4 auxiliary intake cam angle sensors 29 to 31 are pulse signals # 2 to # 4 auxiliary as the auxiliary intake cams 44 for the second to fourth cylinders # 2 to # 4 rotate. An intake cam signal is output to the ECU 2 at every predetermined cam angle (for example, 1 deg). The ECU 2 calculates the intake cam phase θssi # i based on the # 2 to # 4 auxiliary intake cam signals, the auxiliary intake cam signal, the main intake cam signal, and the CRK signal.
[0060]
On the other hand, the intake valve drive mechanism 50 includes main and auxiliary intake cams 43 and 44, a rocker arm 51 that opens and closes the intake valve 6, a link mechanism 52 that supports the rocker arm 51, and the like. The cam profiles of these main and auxiliary intake cams 43 and 44 will be described later.
[0061]
The link mechanism 52 is of a four-joint link type, and includes a first link 53 that extends substantially parallel to the intake valve 6, two second links 54 and 54 that are provided vertically in parallel with each other, and a bias spring 55, A return spring 56 and the like. In the first link 53, the central portion of the rocker arm 51 is rotatably attached to the lower end portion via a pin 51c, and a rotatable roller 53a is provided on the upper end portion.
[0062]
The rocker arm 51 is provided with a rotatable roller 51a at the end on the main intake cam 43 side, and an adjustment bolt 51b is attached to the end on the intake valve 6 side. The valve clearance between the lower edge of the adjustment bolt 51b and the upper edge of the intake valve 6 is set to a predetermined value as described later. The bias spring 55 has one end attached to the rocker arm 51 and the other end attached to the first link 53. Due to the biasing force of the bias spring 55, the rocker arm 51 is biased in the clockwise direction of FIG. 6, and thereby always abuts against the main intake cam 43 via the roller 51a.
[0063]
With the above configuration, when the main intake cam 43 rotates in the clockwise direction in FIG. 6, the rocker arm 51 causes the roller 51 a to roll on the cam surface of the main intake cam 43, thereby responding to the cam profile of the main intake cam 43. Then, the pin 51c is rotated clockwise and counterclockwise using the rotation fulcrum. By the rotation of the rocker arm 51, the adjustment bolt bolt 51b reciprocates in the vertical direction to open and close the intake valve 6.
[0064]
Each second link 54 has one end rotatably connected to the cylinder head 3a via a pin 54a and the other end rotatable to a predetermined portion of the first link 53 via a pin 54b. It is connected. Further, one end of the return spring 56 is attached to the upper second link 54, and the other end is attached to the cylinder head 3a. Due to the biasing force of the return spring 56, the upper second link 54 is biased counterclockwise in FIG. 6, so that the first link 53 is always applied to the auxiliary intake cam 44 via the roller 53a. It is in contact.
[0065]
With the above configuration, when the auxiliary intake cam 44 rotates counterclockwise in FIG. 6, the first link 53 causes the roller 53 a to roll on the cam surface of the auxiliary intake cam 44, thereby Move up and down depending on the profile. As a result, the pin 51c, which is the pivot point of the rocker arm 51, moves in the vertical direction between the lowest position (position shown in FIG. 6) and the uppermost position (position shown in FIG. 15). Along with this, when the rocker arm 51 rotates as described above, the position of the reciprocating movement of the adjusting bolt bolt 51b changes.
[0066]
The cam crest height of the main intake cam 43 is higher than that of the sub intake cam 44, and the cam crest height ratio between the main intake cam 43 and the sub intake cam 44 is from the adjustment bolt 51b to the roller. It is set to a value equal to the ratio between the distance to the center of 51a and the distance from the adjustment bolt 51b to the center of the pin 51c. That is, when the intake rocker arm 51 is driven by the main and auxiliary intake cams 43, 44, the vertical fluctuation amount of the adjustment bolt 51 b due to the cam peak of the main intake cam 43 and the adjustment bolt bolt due to the cam peak of the auxiliary intake cam 44. The amount of fluctuation in the vertical direction of 51b is set to be the same.
[0067]
Next, the main intake cam phase varying mechanism 60 described above will be described. As shown in FIG. 7, the main intake cam phase varying mechanism 60 includes a housing 61, a three-blade vane 62, a hydraulic pump 63, an electromagnetic valve mechanism 64, and the like.
[0068]
The housing 61 is formed integrally with the sprocket 47 described above, and includes three partition walls 61a formed at equal intervals. The vane 62 is coaxially attached to the end of the main intake camshaft 41 on the sprocket 47 side, extends radially outward from the main intake camshaft 41, and is rotatably accommodated in the housing 61. In the housing 61, three advance chambers 65 and three retard chambers 66 are formed between the partition wall 61 a and the vane 62.
[0069]
The hydraulic pump 63 is of a mechanical type connected to the crankshaft 3b. When the crankshaft 3b rotates, along with this, the lubricating oil stored in the oil pan 3d of the engine 3 passes through the oil passage 67c. And is supplied to the electromagnetic valve mechanism 64 via the oil passage 67c in a state where the pressure is increased.
[0070]
The electromagnetic valve mechanism 64 is a combination of a spool valve mechanism 64a and a solenoid 64b, and is connected to the advance chamber 65 and the retard chamber 66 through an advance oil passage 67a and a retard oil passage 67b, respectively. At the same time, the hydraulic pressure supplied from the hydraulic pump 63 is output to the advance chamber 65 and the retard chamber 66 as the advance hydraulic pressure Pad and the retard hydraulic pressure Prt, respectively. The solenoid 64b of the electromagnetic valve mechanism 64 is electrically connected to the ECU 2, and when the control input DUTY_mi from the ECU 2 is input, the spool valve body of the spool valve mechanism 64a is moved according to the control input DUTY_mi. By moving within the range, both the advance hydraulic pressure Pad and the retard hydraulic pressure Prt are changed.
[0071]
In the main intake cam phase variable mechanism 60 described above, during operation of the hydraulic pump 63, the electromagnetic valve mechanism 64 operates in accordance with the control input DUTY_mi, whereby the advance hydraulic pressure Pad is transferred to the advance chamber 65 and the retard hydraulic pressure Prt is transferred. Each is supplied to the retard chamber 66, whereby the relative phase between the vane 62 and the housing 64 is changed to the advance side or the retard side. As a result, the main intake cam phase θmi described above is steplessly changed to the advance side or the retard side within a predetermined range (for example, a range of cam angle 45 deg to 60 deg). The main intake cam phase varying mechanism 60 is provided with a lock mechanism (not shown), and the lock mechanism locks the operation of the main intake cam phase varying mechanism 60 when the hydraulic pressure supplied from the hydraulic pump 63 is low. Is done. That is, the main intake cam phase θmi is prohibited from being changed by the main intake cam phase varying mechanism 60, and the main intake cam phase θmi is locked to a value suitable for idle operation or engine start.
[0072]
Next, the auxiliary intake cam phase varying mechanism 70 described above will be described. As shown in FIG. 8, the auxiliary intake cam phase varying mechanism 70 includes a housing 71, a single vane 72, a hydraulic piston mechanism 73, a motor 74, and the like.
[0073]
The housing 71 is constructed integrally with the gear 46 of the auxiliary intake camshaft 42, and a vane chamber 75 having a sectoral cross section is formed therein. The vane 72 is coaxially attached to the end of the auxiliary intake camshaft 42 on the timing chain 48 side, extends outward from the auxiliary intake camshaft 42, and is rotatably accommodated in the vane chamber 75. By this vane 72, the vane chamber 75 is partitioned into first and second vane chambers 75a and 75b.
[0074]
One end of a return spring 72 a is attached to the vane 72, and the other end of the return spring 72 is attached to the housing 71. By the return spring 72a, the vane 72 is urged in the counterclockwise direction of FIG. 8, that is, in the direction of reducing the volume of the first vane chamber 75a.
[0075]
On the other hand, the hydraulic piston mechanism 73 includes a cylinder 73a and a piston 73b. The internal space of the cylinder 73a communicates with the first vane chamber 75a via the oil passage 76, and hydraulic oil is contained in the internal space of the cylinder 73a, the oil passage 76, and the first vane chamber 75a. Filled. The second vane chamber 75b communicates with the atmosphere side.
[0076]
A rack 77 is attached to the piston 73 b, and a pinion 78 that meshes with the rack 77 is coaxially attached to the rotating shaft of the motor 74. The motor 74 is electrically connected to the ECU 2. When a control input DUTY_msi from the ECU 2 is input, the motor 74 rotates the pinion 78, thereby sliding the piston 73 b in the cylinder 73 a via the rack 77. Let As a result, the hydraulic pressure Psd in the first vane chamber 75a changes, and the vane 72 rotates clockwise or counterclockwise depending on the balance between the changing hydraulic pressure Psd and the biasing force of the return spring 72a. As a result, the auxiliary intake cam phase θmsi is steplessly changed to the advance side or the retard side within a predetermined range (a range corresponding to a cam angle of 180 deg described later).
[0077]
As described above, in the auxiliary intake cam phase variable mechanism 70, the hydraulic piston mechanism 73 and the motor 74 are used in place of the hydraulic pump 63 and the electromagnetic valve mechanism 64 of the main intake cam phase variable mechanism 60 described above, whereby the auxiliary intake cam phase variable mechanism 70. The intake cam phase θmsi is changed. This is because the auxiliary intake cam phase varying mechanism 70 is used for adjusting the intake air amount to each cylinder, and therefore, higher responsiveness than the main intake cam phase varying mechanism 60 is required. Therefore, when the auxiliary intake cam phase variable mechanism 70 does not require high responsiveness (for example, in the valve timing control of the intake valve 6 described later, only one of the late closing control and the early closing control needs to be executed). Instead of the hydraulic piston mechanism 73 and the motor 74, the hydraulic pump 63 and the electromagnetic valve mechanism 64 may be used in the same manner as the main intake cam phase variable mechanism 60.
[0078]
As shown in FIG. 9, the auxiliary intake cam phase varying mechanism 70 is provided with a return spring 72b for urging the vane 72 in the clockwise direction in the figure, and the urging force of the return spring 72b is combined with the return spring 72a. While setting to the same value, the neutral position of the vane 72 shown in the figure may be set to a position corresponding to a value at which the auxiliary intake cam phase θmsi is controlled most frequently. In this way, during the operation of the auxiliary intake cam phase variable mechanism 70, the time during which the vane 72 is held in the neutral position becomes longer, so that the operation stop time of the motor 74 can be secured longer, thereby reducing the power consumption. Can be reduced.
[0079]
Next, the intake cam phase varying mechanism 80 will be described. Since the three intake cam phase varying mechanisms 80 are configured in the same manner, hereinafter, the intake cam phase for changing the intake cam phase θssi # 2 of the auxiliary intake cam 44 for the second cylinder # 2. The variable mechanism 80 will be described as an example. The intake cam phase varying mechanism 80 is for adjusting a steady variation between the cylinders of the intake air amount and does not require high responsiveness. Therefore, the main intake cam phase varying mechanism described above is used. The configuration is the same except for 60 and a part. That is, the intake cam phase varying mechanism 80 includes a housing 81, a vane 82, a hydraulic pump 83, an electromagnetic valve mechanism 84, and the like, as shown in FIG.
[0080]
The housing 81 is configured integrally with the auxiliary intake cam 44 for the second cylinder # 2, and includes one partition wall 81a. The vane 82 is coaxially attached in the middle of the auxiliary intake camshaft 42 and is rotatably accommodated in the housing 81. In the housing 81, an advance chamber 85 and a retard chamber 86 are formed between the partition wall 81 a and the vane 82.
[0081]
Similar to the hydraulic pump 63 described above, the hydraulic pump 83 is a mechanical type coupled to the crankshaft 3b. When the crankshaft 3b rotates, the lubrication stored in the oil pan 3d of the engine 3 is accordingly accompanied. Oil is sucked in through the oil passage 87c and supplied to the electromagnetic valve mechanism 84 through the oil passage 87c in a state where the pressure is increased.
[0082]
The solenoid valve mechanism 84 is a combination of the spool valve mechanism 84a and the solenoid 84b, similar to the solenoid valve mechanism 64 described above, and the advance chamber 85 and the retard oil passage 87b are connected to each other via the advance oil passage 87a and the retard oil passage 87b. The hydraulic pressure supplied from the hydraulic pump 83 is output to the advanced angle chamber 85 and the retarded angle chamber 86 as the advanced angle hydraulic pressure Pad and the retarded hydraulic pressure Prt. The solenoid 84b of the electromagnetic valve mechanism 84 is electrically connected to the ECU 2, and when the control input DUTY_ssi # 2 from the ECU 2 is input, the spool valve body of the spool valve mechanism 84a is controlled according to the control input DUTY_ssi # 2. Thus, both the advance hydraulic pressure Pad and the retard hydraulic pressure Prt are changed by moving within a predetermined movement range.
[0083]
In the above-described intake cam phase variable mechanism 80, during operation of the hydraulic pump 83, the electromagnetic valve mechanism 84 operates in accordance with the control input DUTY_ssi # 2, whereby the advance hydraulic pressure Pad is transferred to the advance chamber 85 and the retard hydraulic pressure is increased. Prt is supplied to each of the retard chambers 86, whereby the relative phase between the vane 82 and the housing 84 is changed to the advance side or the retard side. As a result, the above-described intake cam phase θssi # 2 is steplessly changed to the advance side or the retard side within a predetermined range (for example, a range of 30 deg cam angle). The intake cam phase variable mechanism 80 is provided with a lock mechanism (not shown), which locks the operation of the intake cam phase variable mechanism 80 when the hydraulic pressure supplied from the hydraulic pump 83 is low. Is done. That is, the change of the intake cam phase θssi # 2 by the intake cam phase variable mechanism 80 is prohibited, and the intake cam phase θssi # 2 is locked to the control target value (value 0 described later) at that time.
[0084]
When the internal EGR amount and the intake air amount of each cylinder need to be controlled with high responsiveness and high accuracy as in a compression ignition type internal combustion engine, the intake cam phase variable mechanism 80 is connected to You may comprise similarly to the intake cam phase variable mechanism 70. FIG.
[0085]
Next, the operation of the variable intake valve driving device 40 configured as described above will be described. In the following description, the main and auxiliary intake cams 43 and 44 will be described by taking the one for the first cylinder # 1 as an example. FIG. 11 is a diagram for explaining the cam profiles of the main and auxiliary intake cams 43 and 44, and the operation state when the auxiliary intake cam phase θmsi = 0 deg is set by the auxiliary intake cam phase variable mechanism 70, that is, The operation state when there is no phase difference between the auxiliary intake cam 44 and the main intake cam 43 is shown.
[0086]
A curve indicated by a one-dot chain line in the figure represents a contact point between the main intake cam 43 and the intake rocker arm 51, that is, a fluctuation amount of the roller 51a and its fluctuation timing, and a broken line in the figure. A curve indicated by represents a variation amount of the first link 53, that is, the pin 51c and a variation timing thereof when the auxiliary intake cam 44 rotates. This also applies to FIGS. 12 to 16 below.
[0087]
Further, a curve indicated by a two-dot chain line in FIG. 11 shows, for comparison, an intake cam (hereinafter referred to as an intake cam) of a general engine operated in an Otto cycle, that is, an engine operated such that the expansion ratio and the compression ratio are the same. This represents the fluctuation amount and fluctuation timing of the adjustment bolt 51b due to “Otto intake cam”. Since this curve with the valve clearance added corresponds to the valve lift curve of the intake valve by the Otto intake cam, this curve will be referred to as a valve lift curve as appropriate in the following description.
[0088]
As shown in the figure, the main intake cam 43 has the same lift start timing, that is, valve opening timing, and lift end timing, that is, valve closing timing, which is later than the compression stroke, compared to the Otto intake cam. The cam profile is configured as a closed cam, and the state where the maximum valve lift amount is maintained within a predetermined range (for example, cam angle of 150 deg). In the following description, the state in which the intake valve 6 is closed at a later timing and earlier timing than the Otto intake cam is referred to as “slow closing” or “early closing” of the intake valve 6, respectively.
[0089]
Further, the auxiliary intake cam 44 has a cam profile in which the valve opening timing is earlier than that of the main intake cam 43 and the state where the maximum valve lift amount is maintained within the predetermined range (for example, cam angle of 150 deg). is doing.
[0090]
The operation when the intake valve 6 is actually driven by the main and auxiliary intake cams 43 and 44 having the above cam profile will be described with reference to FIGS. FIG. 12 shows an operation example when the auxiliary intake cam phase θmsi = 0 deg. Note that the curve indicated by the solid line in FIG. 5B shows the actual fluctuation amount and timing of the adjustment bolt 51b, and as described above, the valve valve clearance is added to the intake valve 6b. This corresponds to a valve lift curve indicating the actual valve lift amount and valve timing. Therefore, in the following description, this curve is appropriately referred to as a valve lift curve of the intake valve 6, and the variation amount and timing of the adjustment bolt 51 b are referred to as the valve lift amount and valve timing of the intake valve 6. This point is shown in FIGS. 6 The same applies to the above.
[0091]
As shown in FIG. 12A, when the auxiliary intake cam phase θmsi = 0 deg, the auxiliary intake cam 44 is in the cam intake period while the main intake cam 43 is in contact with the intake rocker arm 51 at a portion where the cam peak is high. It will be in the state contact | abutted to the 1st link 53 in a site | part with a high mountain. That is, during the valve opening operation by the main intake cam 43, the rotation fulcrum of the intake rocker arm 51 is held at the lowest position. As a result, as shown in FIG. 12 (b), the valve lift amount and valve timing of the intake valve 6 are the same as the Otto intake cam, and the valve opening timing is later and the valve closing timing is later. The intake valve 6 is being driven.
[0092]
FIGS. 13 to 15 show operation examples when the auxiliary intake cam phase θmsi is set to 90 deg, 120 deg and 180 deg by the auxiliary intake cam phase variable mechanism 70, respectively. In other words, an operation example is shown in which the phase of the auxiliary intake camshaft 42 is shifted toward the advance side by the cam angle 90 deg, 120 deg and 180 deg with respect to the main intake camshaft 41. FIG. 16 shows an operation example when the auxiliary intake cam phase θmsi is changed from 120 deg to 180 deg.
[0093]
As shown in FIG. 13 (a), when θmsi = 90 deg, the auxiliary intake cam 44 has a cam peak at the second half of the period in which the main intake cam 43 is in contact with the rocker arm 51 at a portion where the cam peak is high. It will be in the state which contacts the 1st link 53 in a low part instead of a high part. As a result, as shown in FIG. 5B, the closing timing of the intake valve 6, that is, the end timing of the valve opening operation by the main intake cam 43 is earlier than when θmsi = 0 deg. Same valve timing.
[0094]
Further, when θmsi is larger than 90 deg, for example, θmsi = 120 deg shown in FIG. 14A, the auxiliary intake cam 44 is in a period during which the main intake cam 43 is in contact with the rocker arm 51 at a portion where the cam peak is high. The time of contact with the first link 53 at the high cam crest portion is shorter than when θmsi = 90 deg. As a result, as shown in FIG. 5B, the closing timing of the intake valve 6 is earlier than when the above-described θmsi = 90 deg, and the closing timing is the same as that of the Otto intake cam. The timing becomes earlier and the intake valve 6 is driven by the early closing cam.
[0095]
Further, as shown in FIG. 16, when the auxiliary intake cam phase θmsi is changed from 120 deg to 180 deg, the auxiliary intake cam 43 is in contact with the intake rocker arm 51 at a portion where the cam peak is high. The time at which 44 contacts the first link 53 at a portion with a high cam peak is gradually reduced. As a result, the valve closing timing of the intake valve 6 is gradually advanced, and the valve lift amount of the intake valve 6 is also gradually decreased from the maximum value. . Thus, when the auxiliary intake cam phase θ msi is set by the auxiliary intake cam phase variable mechanism 70 so that the valve lift amount of the intake valve 6 becomes smaller than the maximum value, the flow velocity of the intake air flowing into the combustion chamber is increased. It is possible to increase the in-cylinder flow. Thereby, combustion efficiency can be improved.
[0096]
Finally, when θmsi = 180 deg, as shown in FIG. 15A, during the period in which the main intake cam 43 is in contact with the rocker arm 51 at a portion where the cam peak is high, the auxiliary intake cam 44 Is in a state where it abuts on the first link 53 at a portion where the cam crest is low. As a result, as shown in FIG. 5B, the amount of fluctuation of the adjustment bolt 51b becomes extremely small and its maximum value is The value is slightly smaller than the valve clearance. As a result, when θmsi = 180 deg, the intake valve 6 is not driven by the adjustment bolt 51b, so that the intake valve 6 is held closed.
[0097]
The variable intake valve driving device 40 described above is configured such that the valve lift curve of the intake valve 6 is the same as that of the Otto intake cam when the auxiliary intake cam phase θmsi = 90 deg. The value of the auxiliary intake cam phase θmsi at which the valve lift curve of the valve 6 is the same as that of the Otto intake cam can be changed as appropriate by changing the cam profiles of the main and auxiliary intake cams 43 and 44.
[0098]
Next, the variable exhaust valve driving device 90 will be described. The variable exhaust valve driving device 90 is configured in substantially the same manner as the variable intake valve driving device 40 described above, and includes a main exhaust camshaft 91 and a sub exhaust camshaft 92 for driving the exhaust valve, and each cylinder. An exhaust valve drive mechanism 100 (only one is shown in FIG. 2) that opens and closes the exhaust valve 7 as the main and sub exhaust camshafts 91 and 92 rotate, and a main exhaust cam phase variable mechanism 110; A sub exhaust cam phase varying mechanism 120 and three exhaust cam phase varying mechanisms 130 are provided.
[0099]
The main exhaust camshaft 91 includes a main exhaust cam 93 provided for each cylinder, a main gear 95 attached integrally, and a sprocket 97 provided at one end. Similar to the sprocket 47 of the main exhaust camshaft 41, the sprocket 97 is connected to the crankshaft 3b via the timing chain 48 described above. As a result, the main exhaust camshaft 91 rotates once every time the crankshaft 3b rotates twice.
[0100]
Further, the main exhaust cam phase variable mechanism 110 has a relative phase with respect to the sprocket 97 of the main exhaust camshaft 91, that is, a relative phase with respect to the crankshaft 3b of the main exhaust camshaft 91 (hereinafter referred to as “main exhaust cam phase”). θme is steplessly changed to the advance side or the retard side. Since the main exhaust cam phase variable mechanism 110 is specifically configured in the same manner as the main intake cam phase variable mechanism 60 described above, the description thereof is omitted here.
[0101]
Further, a main exhaust cam angle sensor 32 is provided at the end of the main exhaust camshaft 91 opposite to the sprocket 97. Like the main intake cam angle sensor 27, the main exhaust cam angle sensor 32 is composed of a magnet rotor and an MRE pickup, and a main exhaust cam signal that is a pulse signal is supplied in accordance with the rotation of the main exhaust camshaft 91. Is output to the ECU 2 at every cam angle (eg, 1 deg). The ECU 2 calculates the main exhaust cam phase θme based on the main exhaust cam signal and the CRK signal.
[0102]
On the other hand, the auxiliary exhaust camshaft 92 has an auxiliary exhaust cam 94 provided for each cylinder and an auxiliary gear 96 having the same number of teeth as the main gear 95. Both the main and sub gears 95 and 96 are pressed so as to always mesh with a pressing spring (not shown), as in the case of the main and sub gears 45 and 46 described above. It is configured not to generate rush. Due to the meshing of both gears 95 and 96, the sub exhaust camshaft 92 rotates in the opposite direction at the same rotational speed as the main exhaust camshaft 91 rotates.
[0103]
Further, the auxiliary exhaust cam phase variable mechanism 120 has a relative phase with respect to the gear 96 of the auxiliary exhaust camshaft 92, that is, relative to the main exhaust camshaft 91 of the auxiliary exhaust camshaft 92 (hereinafter referred to as “subexhaust cam phase”). ) Θmse is changed steplessly. Since the auxiliary exhaust cam phase varying mechanism 120 is specifically configured in the same manner as the auxiliary intake cam phase varying mechanism 70 described above, the description thereof is omitted here.
[0104]
On the other hand, a sub exhaust cam angle sensor 33 is provided at an end of the sub exhaust cam shaft 92 opposite to the sub exhaust cam phase varying mechanism 120. Similar to the main exhaust cam angle sensor 32, the sub exhaust cam angle sensor 33 is composed of a magnet rotor and an MRE pickup. As the sub exhaust cam shaft 92 rotates, a sub exhaust cam signal which is a pulse signal is predetermined. Is output to the ECU 2 at every cam angle (eg, 1 deg). The ECU 2 calculates the sub exhaust cam phase θmse based on the sub exhaust cam signal, the main exhaust cam signal, and the CRK signal.
[0105]
Further, the auxiliary exhaust cam 94 for the first cylinder # 1 is attached to the auxiliary exhaust cam shaft 92 so as to rotate coaxially and integrally, and for the other second to fourth cylinders # 2 to # 4. Each of the sub exhaust cams 94 is connected to the sub exhaust cam shaft 92 via the exhaust cam inter-phase variable mechanism 130. These inter-exhaust cam phase varying mechanisms 130 have a relative phase of the auxiliary exhaust cam 94 for the second to fourth cylinders # 2 to # 4 relative to the auxiliary exhaust cam 94 for the first cylinder # 1 (hereinafter referred to as “exhaust gas”). Θmse # 2 to # 4 (referred to as “phase between cams”) are changed steplessly independently of each other. Specifically, since it is configured in the same manner as the intake cam phase variable mechanism 80 described above, The description is omitted here.
[0106]
On the other hand, the exhaust valve drive mechanism 100 is configured in the same manner as the intake valve drive mechanism 50, and includes main and sub exhaust cams 93 and 94, a rocker arm 101 that opens and closes the exhaust valve 7, and a link mechanism 102 that supports the rocker arm 101. Etc. The main and auxiliary exhaust cams 93 and 94 have cam profiles similar to those of the main and auxiliary intake cams 43 and 44, respectively. Further, since the rocker arm 101 and the link mechanism 102 are respectively configured in the same manner as the rocker arm 51 and the link mechanism 52 described above, detailed description thereof is omitted, but the end portion of the rocker arm 101 opposite to the main exhaust cam 93 is omitted. Is attached with an adjusting bolt 101b similar to the adjusting bolt 51b described above. The rocker arm 101 is rotatably supported by the first link 103.
[0107]
Next, the operation of the variable exhaust valve driving device 90 configured as described above will be described. In the following description, the main and auxiliary exhaust cams 93 and 94 will be described by taking the one for the first cylinder # 1 as an example. FIG. 17 is a diagram for explaining the cam profiles of the main and auxiliary exhaust cams 93 and 94, and shows an example of operation when the auxiliary exhaust cam phase variable mechanism 120 sets the auxiliary exhaust cam phase θmse = 0 deg. ing.
[0108]
A curve indicated by a one-dot chain line in the figure represents a variation amount and timing of a contact point between the main exhaust cam 93 and the exhaust rocker arm 101 when the main exhaust cam 93 is rotated. The fluctuation amount and the fluctuation timing of the first link 103 when the auxiliary exhaust cam 94 rotates are shown. This also applies to FIGS. 18 to 21 below.
[0109]
Further, for comparison, a curve indicated by a two-dot chain line shows the amount of fluctuation of the adjustment bolt 101b and its fluctuation due to an exhaust cam of a general engine operated in an Otto cycle (hereinafter referred to as "Otto exhaust cam"). Represents timing. Since this curve with the valve clearance added corresponds to the valve lift curve of the exhaust valve by the Otto exhaust cam, this curve will be referred to as a valve lift curve as appropriate in the following description.
[0110]
As shown in the figure, the main exhaust cam 93 is configured as a so-called quick-open cam that has the same valve closing timing as the Otto exhaust cam and opens at an earlier timing of the expansion stroke. And has a cam profile in which the maximum valve lift amount continues in a predetermined range (for example, a cam angle of 90 deg).
[0111]
Further, the sub exhaust cam 94 has a cam profile that lasts longer than the main exhaust cam 93 in a predetermined range (for example, a cam angle of 150 deg) in which the valve opening time is longer and the maximum valve lift amount is longer. is doing.
[0112]
The operation when the exhaust valve 7 is actually driven by the main / sub exhaust cams 93 and 94 having the above cam profile will be described with reference to FIGS. FIG. 18 shows an operation example when the sub exhaust cam phase θmse = 0 deg. In addition, the curve shown with the continuous line of the figure shows the actual fluctuation amount of the adjustment bolt 101b and its fluctuation timing, and substantially corresponds to the valve lift curve of the exhaust valve 7 as described above. . Therefore, in the following description, this curve is appropriately referred to as a valve lift curve of the exhaust valve 7, and the actual variation amount and timing of the adjustment bolt 101b are referred to as the valve lift amount and valve timing of the exhaust valve 7. This also applies to FIGS. 19 to 21.
[0113]
When the sub exhaust cam phase θmse = 0 deg, the sub exhaust cam 94 hits the first link 103 at the low cam peak while the main exhaust cam 93 is in contact with the rocker arm 101 at the high cam peak. It will be in contact. As a result, as shown in FIG. 18, the amount of fluctuation of the adjustment bolt 101b becomes extremely small, and its maximum value is slightly smaller than the valve clearance. Accordingly, when θmse = 0 deg, the exhaust valve 7 is not driven by the adjustment bolt 101b, so that the exhaust valve 7 is kept closed.
[0114]
FIGS. 19 to 21 show operation examples when the sub exhaust cam phase θmse is set to 45 deg, 90 deg and 150 deg by the sub exhaust cam phase varying mechanism 120, respectively. In other words, an operation example is shown in which the phase of the auxiliary exhaust camshaft 92 is shifted to the advance side by the cam angles of 45 deg, 90 deg and 150 deg with respect to the main exhaust camshaft 91.
[0115]
Due to the configuration of the exhaust valve drive mechanism 100 described above, the larger the θmse is, that is, the more the phase of the sub exhaust camshaft 92 is advanced with respect to the main exhaust camshaft 91, the higher the main exhaust cam 93 is at the higher cam portion. During the time when it is in contact with the rocker arm 101, the time during which the sub exhaust cam 94 is in contact with the first link 103 at a portion where the cam peak is high becomes longer. As a result, as shown in FIGS. 19 to 21, the opening timing of the exhaust valve 7 becomes earlier as θmse increases.
[0116]
Specifically, in the case of θmse = 45 deg shown in FIG. 19, the valve closing timing is the same and the valve opening timing is later than that of the Otto exhaust cam, and the exhaust valve 7 is driven by the slow opening cam. It becomes. Further, in the case of θmse = 90 deg shown in FIG. 20, the valve timing of the exhaust valve 7 is the same as the valve timing by the Otto exhaust cam. Further, when θmse is larger than 90 deg, for example, when θmse = 150 deg shown in FIG. 21, the valve closing timing is the same and the valve opening timing is earlier than that of the Otto exhaust cam. The valve 7 is driven. Although not shown, the valve lift amount of the exhaust valve 7 is configured to increase as θmse increases in the range of θmse = 0 to 60 deg.
[0117]
On the other hand, as shown in FIG. 3, an ECU 2 is connected to an intake pipe temperature sensor 34, an accelerator opening sensor 35 (load detection means), and an ignition switch (hereinafter referred to as “IG · SW”) 36. The intake pipe temperature sensor 34 outputs a detection signal indicating the air temperature TB in the intake pipe 8 to the ECU 2, and the accelerator opening sensor 35 is a depression amount of an accelerator pedal (not shown) of the vehicle (hereinafter referred to as “accelerator opening”). ) A detection signal representing the AP is output to the ECU 2. Further, the IG / SW 36 is turned ON / OFF by operating an ignition key (not shown), and outputs a signal indicating the ON / OFF state to the ECU 2.
[0118]
Next, the ECU 2 will be described. The ECU 2 includes a microcomputer including an I / O interface, a CPU, a RAM, a ROM, and the like. The operation state of the engine 3 is determined according to the detection signals of the various sensors 20 to 35 and the IG / SW 36 described above. In addition to the determination, various control processes described later are executed in accordance with a control program stored in advance in the ROM, data stored in the RAM, and the like.
[0119]
In the present embodiment, the ECU 2 constitutes a load detection means, a target boost pressure setting means, a supercharging control means, a target valve closing timing setting means, a valve timing control means, and a fuel injection ratio setting means.
[0120]
As shown in FIG. 22, the control device 1 includes a DUTY_th calculation unit 200, a Gcyl calculation unit 210, an auxiliary intake cam phase controller 220, and an inter-intake cam phase controller 230, all of which are specifically ECU2. It is comprised by. In this DUTY_th calculation unit 200, as will be described later, a target opening TH_cmd that is a target value of the throttle valve opening TH is calculated according to the target intake air amount Gcyl_cmd, and further, according to the target opening TH_cmd, A control input DUTY_th to the throttle valve mechanism 16 is calculated.
[0121]
The Gcyl calculation unit 210 calculates a cylinder intake air amount Gcyl that is estimated to be taken into the cylinder according to the equation (1) shown in FIG. In this equation (1), VB represents the intake pipe internal volume, and R represents a predetermined gas constant. The symbol n represents the discretized time, and each discrete data (time series data) with the symbols (n), (n-1), etc. has a predetermined period (for example, input synchronization of TDC signal, constant value, etc.) Indicates that the data is sampled. The data with the symbol (n) indicates the current value, and the data with the symbol (n-1) indicates the previous value. This also applies to the other discrete data in the present specification below. Further, in the description in the present specification, symbols (n), (n−1) and the like representing discrete data are omitted as appropriate.
[0122]
Further, the auxiliary intake cam phase controller 220 calculates a control input DUTY_msi to the auxiliary intake cam phase variable mechanism 70 according to the cylinder intake air amount Gcyl calculated by the Gcyl calculation unit 210, and the details thereof. Will be described later.
[0123]
Further, the inter-intake cam phase controller 230 is for correcting the variation in the intake air amount between the cylinders, and a detailed description thereof will be omitted here. The phase θssi # i_cmd (# i = # 2 to # 4) is calculated, and the control inputs DUTY_ssi # 2 to # 4 are calculated accordingly. Then, by outputting these control inputs DUTY_ssi # 2 to 4 to the three intake cam phase variable mechanisms 80, the phase between the intake cams of the second to fourth cylinders # 2 to # 4 with respect to the first cylinder # 1. θssi # 2 to # 4 are controlled.
[0124]
Next, the auxiliary intake cam phase controller 220 will be described. As shown in FIG. 23, the auxiliary intake cam phase controller 220 calculates a first SPAS controller 221 that calculates a target auxiliary intake cam phase θmsi_cmd, and a control input DUTY_msi. A second SPAS controller 225.
[0125]
The first SPAS controller 221 (valve closing timing setting means) is configured to perform cylinder intake air amount Gcyl, target intake air amount Gcyl_cmd, and required drive torque according to an adaptive predictive response designation control algorithm described below. The target auxiliary intake cam phase θmsi_cmd is calculated in accordance with TRQ_eng, and includes a state predictor 222, an onboard identifier 223, and a sliding mode controller 224.
[0126]
First, the state predictor 222 will be described. The state predictor 222 predicts (calculates) a predicted intake air amount Pre_Gcyl, which is a predicted value of the cylinder intake air amount Gcyl, using a prediction algorithm described below.
[0127]
First, Sub-intake cam phase θmsi and cylinder intake air amount Gcyl 24 is modeled as an ARX model (auto-regressive model with exogeneous input), which is a discrete-time system model, Equation (2) shown in FIG. can get. In the formula (2), d represents a dead time determined by the characteristics of the controlled object. Further, a1, a2 and b1 represent model parameters, which are sequentially identified by the onboard identifier 223 as will be described later.
[0128]
Next, when the equation (2) is shifted to the future side by the discrete time [d−1], the equation (3) in FIG. 24 is obtained. Furthermore, the matrices A and B are defined as shown in the equations (4) and (5) shown in FIG. 24 using the model parameters a1, a2, and b1, and the future value [Gcyl ( In order to eliminate (n + d−2), Gcyl (n + d−3)], the equation (3) is modified by repeatedly using the recurrence equation of the above equation (3) to obtain the equation (6) shown in FIG. It is done.
[0129]
Although it is possible to calculate the predicted intake air amount Pre_Gcyl by using this equation (6), a steady deviation and a predicted deviation in the predicted intake air amount Pre_Gcyl due to insufficient model order, nonlinear characteristics of the control target, and the like. Modeling errors can occur.
[0130]
In order to avoid this, in the state predictor 222 of the present embodiment, the predicted intake air amount Pre_Gcyl is calculated by the equation (7) shown in FIG. 24 instead of the equation (6). This equation (7) is obtained by adding a compensation parameter γ1 for compensating for a stationary deviation and a modeling error to the right side of the equation (6).
[0131]
Next, the on-board identifier 223 will be described. The on-board identifier 223 uses the sequential identification algorithm described below to minimize the identification error ide that is the deviation between the predicted intake air amount Pre_Gcyl and the cylinder intake air amount Gcyl (that is, the predicted intake air). This is to identify the matrix parameter α1, α2, βj of the model parameter and the vector θs of the compensation parameter γ1 in the above-described equation (7) so that the amount Pre_Gcyl matches the cylinder intake air amount Gcyl).
[0132]
Specifically, the vector θs (n) is calculated by the equations (8) to (13) shown in FIG. The transposed matrix of this vector θs (n) is defined as in equation (12) in the figure. In Equation (8), KPs (n) represents a gain coefficient vector, and the gain coefficient KPs (n) is calculated by Equation (9). Ps (n) in this equation (9) is a d + second-order square matrix defined by equation (10), and ζs (n) is a vector whose transpose matrix is defined as in equation (13). is there. Further, the identification error ide (n) of equation (8) is calculated by equation (11).
[0133]
In the identification algorithm as described above, one of the following four identification algorithms is selected according to the setting of the weight parameters λ1 and λ2 of Expression (10).
That is,
λ1 = 1, λ2 = 0; Fixed gain algorithm
λ1 = 1, λ2 = 1; least square algorithm
λ1 = 1, λ2 = λ; gradually decreasing gain algorithm
λ1 = λ, λ2 = 1; weighted least squares algorithm
However, λ is a predetermined value set to 0 <λ <1.
In the present embodiment, a weighted least square algorithm is employed in order to optimally ensure both the identification accuracy and the convergence speed of the vector θs to the optimum value.
[0134]
Next, the sliding mode controller (hereinafter referred to as “SLD controller”) 224 will be described. Based on the sliding mode control algorithm, the SLD controller 224 sets the target auxiliary intake cam so that the cylinder intake air amount Gcyl converges to the target intake air amount Gcyl_cmd and the auxiliary intake cam phase θmsi is constrained to the basic value θmsi_base. The phase θmsi_cmd is calculated, and this sliding mode control algorithm will be described below.
[0135]
First, in this sliding mode control algorithm, Expression (14) shown in FIG. 26 is used as a control target model. This equation (14) is obtained by shifting the above-described equation (6) of FIG. 24 to the future side by the discrete time “1”.
[0136]
When the controlled object model shown in the equation (14) is used, the switching function σs is set as follows. That is, as shown in the equation (15) in FIG. 26, when the tracking error Es is defined as a deviation between the cylinder intake air amount Gcyl and the target intake air amount Gcyl_cmd, the switching function σs is as shown in the equation (16) in FIG. Is set as a linear function of time-series data (discrete data) of the tracking error Es. In addition, Ss shown in Formula (16) represents the switching function setting parameter.
[0137]
In the sliding mode control algorithm, when the switching function σs is composed of two state variables [Es (n), Es (n−1)] as in the present embodiment, as shown in FIG. The phase space composed of variables becomes a two-dimensional phase plane having the vertical axis and the horizontal axis, respectively. On this phase plane, a combination of values of two state variables satisfying σs = 0 is expressed by the formula [Es. (N) = − Ss · Es (n−1)] is placed on a straight line called a switching straight line.
[0138]
Since this equation [Es (n) = − Ss · Es (n−1)] represents a first-order lag system without input, the switching function setting parameter Ss is set to, for example, −1 <Ss <1. When this first-order lag system is stabilized, the combination of the two state variables [Es (n), Es (n-1)] converges to an equilibrium point where the value becomes 0 with the passage of time. That is, by converging the tracking error Es to the value 0 in this way, the cylinder intake air amount Gcyl can be converged to the target intake air amount Gcyl_cmd. The time until the two state variables [Es (n), Es (n-1)] asymptotically approach the switching straight line is referred to as a reaching mode, and the behavior of these sliding to the equilibrium point is referred to as a sliding mode.
[0139]
In this case, when the switching function setting parameter Ss is set to a positive value, the first-order lag system expressed by the formula [Es (n) = − Ss · Es (n−1)] is a vibration stable system, and thus the state variable It is not preferable as the convergence behavior of [Es (n), Es (n-1)]. Therefore, in this embodiment, the switching function setting parameter Ss is set as shown in the equation (17) in FIG. When the switching function setting parameter Ss is set in this way, as shown in FIG. 29, the smaller the absolute value of the switching function setting parameter Ss is, the smaller the convergence speed of the tracking error Es to the value 0, that is, the cylinder intake air amount Gcyl. The convergence speed to the target intake air amount Gcyl_cmd increases. As described above, in the sliding mode control, the convergence behavior and the convergence speed of the cylinder intake air amount Gcyl to the target intake air amount Gcyl_cmd can be arbitrarily designated by the switching function setting parameter Ss.
[0140]
Further, the control input Uspas (n) [= θmsi_cmd (n)] for placing the combination of these state variables [Es (n), Es (n−1)] on the switching line is expressed by the equation (18 ) Is defined as the sum of the equivalent control input Ueq (n), the reaching law input Urch (n), and the valve control input Uvt (n).
[0141]
This equivalent control input Ueq (n) is for constraining the combination of [Es (n), Es (n-1)] on the switching straight line. Specifically, the equation shown in FIG. It is defined as (19). This equation (19) is derived as follows. That is, when the formula (22) shown in FIG. 27 is modified based on the above-described formula (16), the formula (23) shown in FIG. 27 is obtained. Next, the recurrence formula is repeated for this formula (23). When deformed by use, the equation (24) shown in FIG. 27 is obtained. Further, in this equation (24), when the terms of the auxiliary intake cam phase θmsi are collectively changed, equation (25) shown in FIG. 27 is obtained. Next, in this equation (25), the sub intake cam phase θmsi (n) on the left side is replaced with the equivalent control input Ueq (n), and at the same time, based on the relationship of Pre_Gcyl (n) ≈Gcyl (n + d−1) described above, The formula (19) is derived by replacing the future value Gcyl (n + d-1) or the like of the cylinder intake air amount on the right side with the predicted value Pre_Gcyl.
[0142]
In addition, the reaching law input Urch (n) is displayed on the switching line when the combination of [Es (n), Es (n-1)] deviates from the switching line due to disturbance or modeling error. This is for convergence, and is specifically defined as shown in Expression (20) shown in FIG.
[0143]
Further, the valve control input Uvt (n) is a feedforward input for constraining the auxiliary intake cam phase θmsi to its basic value θmsi_base. Specifically, as shown in the equation (21) of FIG. It is defined as a value equal to the value θmsi_base. The basic value θmsi_base is calculated according to the required drive torque TRQ_eng, as will be described later.
[0144]
As described above, in the first SPAS controller 221, the state predictor 222 calculates the predicted intake air amount Pre_Gcyl by the state prediction algorithm to which the compensation parameter γ1 is added, and the compensation parameter γ1 is determined by the onboard identifier 223. Thus, the predicted intake air amount Pre_Gcyl can be accurately calculated while compensating for the above-described steady-state deviation and modeling error.
[0145]
Further, in the SLD controller 224, the tracking error Es can be converged to a value of 0 by the reaching law input Urch and the equivalent control input Ueq. That is, the cylinder intake air amount Gcyl can be converged to the target intake air amount Gcyl_cmd, and at the same time, the convergence behavior and the convergence speed can be arbitrarily designated by setting the switching function setting parameter Ss. Therefore, the convergence speed of the cylinder intake air amount Gcyl to the target intake air amount Gcyl_cmd can be set to an appropriate value according to the characteristics of the control target (intake system including the auxiliary intake cam phase variable mechanism 70). Thereby, controllability can be improved.
[0146]
In addition, the valve control input Uvt can constrain the auxiliary intake cam phase θmsi to its basic value θmsi_base, and at the same time, the compensation parameter γ1 is included in the equivalent control input Ueq. The cylinder intake air amount Gcyl can be appropriately converged to the target intake air amount Gcyl_cmd while compensating for the influence of Uvt.
[0147]
Next, the second SPAS controller 225 described above will be described. The second SPAS controller 225 calculates a control input DUTY_msi according to the auxiliary intake cam phase θmsi and the target auxiliary intake cam phase θmsi_cmd by the same control algorithm as that of the first SPAS controller 221 described above with some exceptions. 30, the state predictor 226, the on-board identifier 227, and the sliding mode controller 228 are included.
[0148]
The state predictor 226 predicts (calculates) a predicted sub-intake cam phase Pre_θmsi, which is a predicted value of the sub-intake cam phase θmsi, using a prediction algorithm similar to the state predictor 222 described above.
[0149]
Specifically, Expression (26) shown in FIG. 31 is used as the control target model. In the equation (26), dx represents a dead time determined by the characteristics of the controlled object, and a1 ′, a2 ′, and b1 ′ represent model parameters. The symbol m represents a discretized time, and each discrete data with a symbol (m) or the like is data sampled at a predetermined cycle shorter than the discrete data with the symbol (n) described above. ing. This point is the same for the other discrete data in the present specification below, and in the description in the present specification, the symbol (m) indicating discrete data is omitted as appropriate. As described above, the reason why the sampling period of each discrete data in equation (26) is set to a period shorter than each discrete data in equation (2) described above is that the second intake by the second SPAS controller 225 If the convergence speed of the cam phase θmsi to the target auxiliary intake cam phase θmsi_cmd is slower than the convergence speed of the cylinder intake air amount Gcyl to the target intake air amount Gcyl_cmd by the first SPAS controller 221, the controllability is reduced. This is to avoid this and ensure good controllability.
[0150]
The matrices A ′ and B ′ are defined as the equations (27) and (28) shown in FIG. 31 using the model parameters a1 ′, a2 ′, and b1 ′, and the equation (26) is the state predictor described above. By modifying in the same manner as in the case of 222, the equation (29) shown in FIG. 31 is derived. In this equation (29), γ ′ is a compensation parameter for compensating for a steady-state deviation and a modeling error, similar to the compensation parameter γ described above.
[0151]
Further, the on-board identifier 227 also minimizes the identification error ide ′ that is the deviation between the predicted auxiliary intake cam phase Pre_θmsi and the auxiliary intake cam phase θmsi by the same sequential identification algorithm as the on-board identifier 223 described above. As described above (that is, the predicted auxiliary intake cam phase Pre_θmsi matches the auxiliary intake cam phase θmsi), the model parameter matrix components α1 ′, α2 ′, βj ′ and the compensation parameter γ1 ′ in the above equation (29) This identifies θs ′.
[0152]
Specifically, the vector θs ′ (m) is calculated by the equations (30) to (35) shown in FIG. Since these equations (30) to (35) are configured in the same manner as the aforementioned equations (8) to (13), the description thereof will be omitted.
[0153]
Next, the sliding mode controller (hereinafter referred to as “SLD controller”) 228 will be described. The SLD controller 228 calculates the control input DUTY_msi so that the auxiliary intake cam phase θmsi converges to the target auxiliary intake cam phase θmsi_cmd based on the sliding mode control algorithm.
[0154]
Specifically, the control input DUTY_msi is calculated by the algorithm shown in equations (36) to (41) in FIG. That is, as shown in the equation (36) in the figure, if the tracking error Es ′ is defined as a deviation between the auxiliary intake cam phase θmsi and the target auxiliary intake cam phase θmsi_cmd, the switching function σs ′ and the switching function setting parameter Ss ′. Are defined as shown in equations (37) and (38) of FIG. Further, the control input DUTY_msi is defined as the sum of the equivalent control input Ueq ′ and the reaching law input Urch ′, as shown in the equation (39) of the figure, and the equivalent control input Ueq ′ and the reaching law input Urch ′ are respectively It is defined as shown in equations (40) and (41) in FIG. As shown in the equation (39), in the SLD controller 228, the auxiliary intake cam phase θmsi may be controlled to converge to the target auxiliary intake cam phase θmsi_cmd. Therefore, the valve control input Uvt described above is equal to the control input DUTY_msi. Omitted from the input component.
[0155]
As described above, also in the second SPAS controller 225, the state predictor 226 calculates the predicted auxiliary intake cam phase Pre_θmsi by the state prediction algorithm to which the compensation parameter γ1 ′ is added, and the compensation parameter γ1 ′ is identified on the board. Thus, the predicted auxiliary intake cam phase Pre_θmsi can be accurately calculated while compensating for the steady-state deviation and the modeling error.
[0156]
Further, in the SLD controller 227, the auxiliary intake cam phase θmsi can be converged to the target auxiliary intake cam phase θmsi_cmd by the reaching law input Urch ′ and the equivalent control input Ueq ′. It can be arbitrarily specified by setting the switching function setting parameter Ss ′. Therefore, the convergence speed of the auxiliary intake cam phase θmsi to the target auxiliary intake cam phase θmsi_cmd can be set to an appropriate value according to the characteristics of the control target (system including the auxiliary intake cam phase variable mechanism 70). Thereby, controllability can be improved.
[0157]
In the above two switching function setting parameters Ss and Ss ′, when these are set to values that satisfy the relationship of −1 <Ss <Ss ′ <0, the speed of control by the second SPAS controller 225 is set to the first SPAS. The control by the controller 221 can be enhanced, and the controllability of the auxiliary intake cam phase controller 220 can be improved.
[0158]
Hereinafter, the engine control process including the valve timing control of the intake valve 6 executed by the ECU 2 will be described with reference to FIG. This figure shows the main control contents of the engine control process. In this program, first, the fuel control process is executed in step 1 (abbreviated as “S1” in the figure. The same applies hereinafter). This fuel control process calculates the required drive torque TRQ_eng, the main fuel injection rate Rt_Pre, the cylinder intake air amount Gcyl, the target intake air amount Gcyl_cmd, the fuel injection amounts TOUT_main, TOUT_sub, and the like according to the operating state of the engine 3. Yes, the specific contents will be described later.
[0159]
Next, in step 2, a supercharging pressure control process is executed. This supercharging pressure control process calculates the control input Dut_wg to the wastegate valve 10d according to the operating state of the engine 3, and the specific contents thereof will be described later.
[0160]
Next, in step 3, an intake valve control process is executed. This intake valve control process calculates the above-described various control inputs DUTY_mi, DUTY_msi, and DUTY_ssi # 2 to # 4 according to the operating state of the engine 3, and the specific contents thereof will be described later.
[0161]
Next, in step 4, an exhaust valve control process is executed. Although the exhaust valve control process is not described in detail, the main exhaust cam phase variable mechanism 110, the sub exhaust cam phase variable mechanism 120, and the exhaust cam phase variable mechanism 130 described above according to the operating state of the engine 3. Control inputs DUTY_me, DUTY_mse, and DUTY_sse # 2 to # 4 are respectively calculated.
[0162]
Next, in step 5, throttle valve control processing is executed. This throttle valve control process calculates the above-described control input DUTY_th according to the operating state of the engine 3, and the specific contents thereof will be described later.
[0163]
Next, in step 6, after executing the ignition timing control process, this program is terminated. The ignition timing control process calculates the ignition timing of the air-fuel mixture by the spark plug 5 in accordance with the operating state of the engine 3 although a detailed description thereof is omitted.
[0164]
Next, the fuel control process in step 1 will be described with reference to FIG. As shown in the figure, in this program, first, in step 10, it is determined whether or not an intake / exhaust valve failure flag F_VLVNG or a throttle valve failure flag F_THNG is “1”. The intake / exhaust valve failure flag F_VLVNG is set to “1” when the variable intake valve drive device 40 or the variable exhaust valve drive device 90 is out of order, and is set to “0” when both are normal. It is. Further, the throttle valve failure flag F_THNG is set to “1” when the throttle valve mechanism 16 has failed, and is set to “0” when it is normal.
[0165]
When the determination result in step 10 is NO and the variable intake valve driving device 40, the variable exhaust valve driving device 90, and the throttle valve mechanism 16 are all normal, the process proceeds to step 11 and the required drive torque TRQ_eng (represents a load) Parameter) is calculated by searching a map shown in FIG. 36 according to the engine speed NE and the accelerator pedal opening AP.
[0166]
The predetermined values AP1 to AP3 of the accelerator opening AP in the figure are set so that the relationship AP1>AP2> AP3 is established, and the predetermined value AP1 is the maximum value of the accelerator opening AP, that is, the maximum depression amount. Is set. As shown in the figure, in this map, the required drive torque TRQ_eng is set to a larger value as the engine speed NE is higher and the accelerator pedal opening AP is larger in a range of NE ≦ NER2 (predetermined value). Has been. This is because the required engine torque increases as the load on the engine 3 increases. When AP = AP1, the required drive torque TRQ_eng is set to the maximum value in the range of NER1 (predetermined value) <NE ≦ NER2. Further, in the range of NER2 <NE, the required drive torque TRQ_eng is set to a larger value as the accelerator opening AP is larger, and at the same time, is set to a smaller value as the engine speed NE is higher. . This is due to the output characteristics of the engine torque with respect to the engine speed NE.
[0167]
In step 12 following step 11, it is determined whether or not the required drive torque TRQ_eng calculated in step 11 is smaller than a predetermined stratified combustion operation threshold value TRQ_disc. Note that the stratified charge combustion operation represents an operation in which the air-fuel mixture is stratified by performing fuel injection into the cylinder by the main fuel injection valve 4 during the compression stroke.
[0168]
If the determination result in step 12 is YES and the engine 3 is to be stratified combustion operation, the routine proceeds to step 13 where the target air-fuel ratio KCMD_disc for stratified combustion operation is searched in a table (not shown) according to the required drive torque TRQ_eng. To calculate. In this table, the target air-fuel ratio KCMD_disc for stratified charge combustion operation is set to a value in a predetermined extreme lean region (for example, A / F = 30 to 40).
[0169]
Next, the routine proceeds to step 14, where the target air-fuel ratio KCMD is set to the target air-fuel ratio KCMD_disc for stratified combustion operation, and then at step 15, the main fuel injection rate Rt_Pre as the first fuel injection ratio is set to a predetermined maximum value Rtmax ( 100%). Thereby, the fuel injection by the sub fuel injection valve 15 is stopped as will be described later. Proceeding to step 16, the cylinder intake air amount Gcyl and the target intake air amount Gcyl_cmd are calculated.
[0170]
These cylinder intake air amount Gcyl and target intake air amount Gcyl_cmd are specifically calculated by a program shown in FIG. That is, first, in step 30 in the figure, the cylinder intake air amount Gcyl is calculated by the above-described equation (1).
[0171]
Next, at step 31, the basic value Gcyl_cmd_base of the target intake air amount is calculated by searching a map shown in FIG. 38 according to the engine speed NE and the required drive torque TRQ_eng. The predetermined drive torque predetermined values TRQ_eng1 to 3 in this map are set so that the relationship of TRQ_eng1>TRQ_eng2> TRQ_eng3 is established. As shown in the figure, the basic value Gcyl_cmd_base of the target intake air amount is set to a larger value as the engine speed NE is higher or the required drive torque TRQ_eng is larger. This is because a larger engine output is required as the load of the engine 3 is larger, and a larger amount of intake air is required.
[0172]
Next, in step 32, the air-fuel ratio correction coefficient Kgcyl_af is calculated by searching the table shown in FIG. 39 according to the target air-fuel ratio KCMD. In this table, the air-fuel ratio correction coefficient Kgcyl_af increases as the target air-fuel ratio KCMD is richer. small Is set to a value. This is because the necessary intake air amount becomes smaller as the air-fuel ratio of the air-fuel mixture is controlled to be richer. Note that the value KCMDST in the figure is a value corresponding to the theoretical air-fuel ratio.
[0173]
Next, the routine proceeds to step 33, where the product of the basic value of the target intake air amount and the air-fuel ratio correction coefficient (Kgcyl_af · Gcyl_cmd_base) is set as the target intake air amount Gcyl_cmd, and then this program ends.
[0174]
Returning to FIG. 35, after executing step 16 as described above, the routine proceeds to step 17, where the fuel injection control process is executed. In this fuel injection control process, specifically, control inputs to the main / sub fuel injection valves 4 and 15 are calculated as follows.
[0175]
First, the main fuel injection amount TOUT_main that is the fuel injection amount of the main fuel injection valve 4 and the sub fuel injection amount TOUT_sub that is the fuel injection amount of the sub fuel injection valve 15 are calculated. That is, the final total fuel injection amount TOUT for each cylinder is calculated for each cylinder based on the operating state of the engine 3 and the target air-fuel ratio KCMD described above, and then the main fuel injection amount TOUT is calculated by the following equations (42) and (43). -Sub fuel injection amounts TOUT_main and TOUT_sub are calculated, respectively.
TOUT_main = [TOUT · Rt_Pre] / 100 (42)
TOUT_sub = [TOUT · (100−Rt_Pre)] / 100 (43)
Referring to this equation (43), the second fuel injection ratio [(100−Rt_Pre) / 100] by the sub fuel injection valve 15 becomes 0 when Rt_Pre = Rtmax (100%). That is, it is understood that TOUT_sub = 0 and fuel injection by the auxiliary fuel injection valve 15 is stopped.
[0176]
Next, a control input to the main / sub fuel injection valves 4 and 15 is calculated by searching a table (not shown) according to the calculated main / sub fuel injection amounts TOUT_main, TOUT_sub. After executing step 17 as described above, this program is terminated.
[0177]
On the other hand, when the determination result in step 12 is NO, it is determined that the engine 3 should be the premixed lean operation of the uniform combustion operation instead of the stratified combustion operation, and the process proceeds to step 18 where the target empty for the premixed lean operation The fuel ratio KCMD_lean is calculated by searching a table (not shown) according to the required drive torque TRQ_eng. In this table, the target air-fuel ratio KCMD_lean for the premixed lean operation is set to a predetermined lean region value (for example, A / F = 18 to 21).
[0178]
Next, the routine proceeds to step 19, where the target air-fuel ratio KCMD is set to the target air-fuel ratio KCMD_lean for premixed lean operation, and then at step 20, the main fuel injection rate Rt_Pre is shown in FIG. 40 according to the required drive torque TRQ_eng. Calculate by searching the table. In the following tables and maps including the same figure, various predetermined values TRQ_idle, TRQ_disc, TRQott, and TRQ1 to TRQ4 of the required driving torque TRQ_eng have a relationship of TRQ_idle <TRQ_disc <TRQ1 <TRQott <TRQ2 <TRQ3 <TRQ4, respectively. It is set to a value that holds. TRQ_idle represents a predetermined idle operation value. Further, the predetermined value TRQ2 corresponds to a predetermined load.
[0179]
As shown in the figure, in this table, the main fuel injection rate Rt_Pre is set to a smaller value as the required drive torque TRQ_eng is larger in the range of TRQ1 <TRQ_eng <TRQ4. This is due to the following reason. That is, as the required drive torque TRQ_eng is larger, the supercharging pressure Pc is controlled to be higher, so that the temperature of the intake air rises and knocking is likely to occur. Therefore, in order to avoid such knocking, it is necessary to increase the cooling effect of the intake air by the fuel vaporization cooling device 12 described above by increasing the fuel injection amount TOUT_sub of the auxiliary fuel injection valve 15. The fuel injection rate Rt_Pre is set as described above.
[0180]
Further, in this table, the main fuel injection rate Rt_Pre is set to a predetermined minimum value Rtmin (10%) when the required drive torque TRQ_eng is in the range of the predetermined value TRQ4 or more, and in the range of the predetermined value TRQ1 or less, the above-mentioned maximum value is set. The value Rtmax is set.
[0181]
After execution of this step 20, after executing steps 16 and 17 described above, this program is terminated.
[0182]
On the other hand, when the determination result in step 10 is YES and any of the variable intake valve driving device 40, the variable exhaust valve driving device 90, and the throttle valve mechanism 16 has failed, the process proceeds to step 21 and the required drive torque TRQ_eng. Is set to a predetermined failure time value TRQ_fs. Thereafter, the routine proceeds to step 22 where the main fuel injection rate Rt_Pre is set to the aforementioned maximum value Rtmax. Next, as described above, after executing steps 16 and 17, the program is terminated.
[0183]
Next, the above-described supercharging pressure control process will be described with reference to FIG. As shown in the figure, in this program, first, in step 40, it is determined whether or not the intake / exhaust valve failure flag F_VLVNG or the throttle valve failure flag F_THNG is “1”.
[0184]
When the determination result is NO and the variable intake valve drive device 40, the variable exhaust valve drive device 90, and the throttle valve mechanism 16 are all normal, the routine proceeds to step 41, where the engine start flag F_ENGSTART is “1”. It is determined whether or not. This engine start flag F_ENGSTART is set in a determination process (not shown) by determining whether or not the engine is being started, that is, cranking, according to the engine speed NE and the output state of the IG / SW 36. Yes, specifically, “1” is set when the engine is starting, and “0” is set otherwise.
[0185]
If the decision result in the step 41 is YES and the engine is being started, the process proceeds to a step 43, the control input Dut_wg to the waste gate valve 10d is set to a predetermined full open value Dut_wgmax, and then this program is ended. Thereby, the wastegate valve 10d is controlled to a fully open state, and the supercharging operation by the turbocharger device 10 is substantially stopped.
[0186]
On the other hand, when the determination result of step 41 is NO and the engine is not starting, the routine proceeds to step 42 where it is determined whether or not the catalyst warm-up flag F_CATHOT is “1”. This catalyst warm-up flag F_CATHOT is set to “1” when the catalyst warm-up control after engine start is being executed, and is set to “0” otherwise. The catalyst warm-up control is for rapidly activating the catalyst after the engine is started, and is started immediately after the engine start is completed. In this catalyst warm-up control, when the execution time Tcat reaches a predetermined value, the control ends and the catalyst warm-up flag F_CATHOT is set to “0”.
[0187]
If the decision result in the step 42 is YES and the catalyst warm-up control is being performed, the process proceeds to a step 44, and the control input Dut_wg to the waste gate valve 10d is set to the full open value Dut_wgmax as in the above step 43. Then, this program is terminated.
[0188]
On the other hand, if the determination result in step 42 is NO and the engine is not being started and the catalyst warm-up control is not being performed, the process proceeds to step 45 where the basic value Dut_wg_bs of the control input Dut_wg is shown in FIG. It is calculated by searching the table shown.
[0189]
As shown in the figure, in this table, the basic value Dut_wg_bs is set to a smaller value as the required drive torque TRQ_eng is larger in the range of TRQ1 <TRQ_eng <TRQ2. This is because as the required drive torque TRQ_eng is larger, it is necessary to increase the supercharging pressure Pc for the purpose of increasing the charging efficiency due to supercharging. Further, the basic value Dut_wg_bs is set to a predetermined fully closed value Dut_wgmin in the range of TRQ2 ≦ TRQ_eng ≦ TRQ3, and this increases the supercharging effect according to the load of the engine 3 being in the high load range. This is to get the maximum. Further, the basic value Dut_wg_bs is set to a smaller value as the required drive torque TRQ_eng is larger in the range of TRQ3 <TRQ_eng, in order to avoid the occurrence of knocking.
[0190]
Next, at step 46, the target boost pressure Pc_cmd is calculated by searching the table shown in FIG. 43 according to the required drive torque TRQ_eng. As shown in the figure, in this table, the target boost pressure Pc_cmd is set to a larger value as the required drive torque TRQ_eng is larger in the range of TRQ_idle <TRQ_eng <TRQ2. As described above, this is to further increase the charging efficiency by supercharging. Further, the target supercharging pressure Pc_cmd is set to a predetermined fully closed value Dut_wgmin in the range of TRQ2 ≦ TRQ_eng ≦ TRQ3, and this is for obtaining the maximum supercharging effect as described above. Further, the target boost pressure Pc_cmd is set to a smaller value as the required drive torque TRQ_eng is larger in the range of TRQ3 <TRQ_eng <TRQ4, and this is to avoid the occurrence of knocking. In addition, Patm in the same figure has shown atmospheric pressure, and this point is the same also below.
[0191]
Next, the process proceeds to step 47, the control input Dut_wg is calculated by the IP control algorithm shown in the following equation (44), and then this program is terminated. Thereby, feedback control is performed so that the supercharging pressure Pc converges to the target supercharging pressure Pc_cmd.
Dut_wg = Dut_wg_bs + Kpwg · Pc + Kiwg · Σ (Pc-Pc_cmd) (44)
Here, Kpwg represents the P term gain, and Kiwg represents the I term gain.
[0192]
On the other hand, when the determination result in step 40 is YES and any of the variable intake valve driving device 40, the variable exhaust valve driving device 90, and the throttle valve mechanism 16 has failed, the process proceeds to step 48, and the above-described step 43 is performed. , 44, the control input Dut_wg to the wastegate valve 10d is set to the fully open value Dut_wgmax, and then this program is terminated.
[0193]
Next, the above-described intake valve control processing in step 3 will be described with reference to FIGS. As shown in the figure, in this program, first, in step 60, it is determined whether or not the intake / exhaust valve failure flag F_VLVNG is "1". The determination result is NO, and the variable intake valve drive is determined. When both the device 40 and the variable exhaust valve drive device 90 are normal, the routine proceeds to step 61, where it is determined whether or not the engine start flag F_ENGSTART described above is “1”.
[0194]
If the determination result is YES and the engine is starting, the routine proceeds to step 62, where the target main intake cam phase θmi_cmd, which is the target value of the main intake cam phase θmi, is set to a predetermined idle value θmi_idle.
[0195]
Next, the routine proceeds to step 63, where the target auxiliary intake cam phase θmsi_cmd is set to a predetermined starting value θmsi_st. This starting value θmsi_st is set as a predetermined value for the slow closing of the intake valve 6. Thereafter, the routine proceeds to step 64 where the target intake cam phase θssi # i_cmd (# i = # 2 to # 4) is set to a value of 0.
[0196]
Next, the routine proceeds to step 65 in FIG. 45, where the control input DUTY_mi to the main intake cam phase variable mechanism 60 is calculated by searching a table (not shown) according to the target main intake cam phase θmi_cmd. Thereafter, in step 66, the control input DUTY_msi to the auxiliary intake cam phase variable mechanism 70 is calculated by searching a table (not shown) according to the target main intake cam phase θmi_cmd. In step 66, the control input DUTY_msi may be calculated by the same method as in step 74 described later.
[0197]
Next, in step 67, the control input DUTY_ssi # i to the intake cam inter-camera phase varying mechanism 80 is calculated by searching a table (not shown) according to the target intake cam inter-phase θssi # i_cmd, and then the program is terminated. To do.
[0198]
Returning to FIG. 44, if the determination result in step 61 is NO and the engine is not being started, the process proceeds to step 68, where it is determined whether or not the catalyst warm-up flag F_CATHOT is "1". If the determination result is YES and the catalyst warm-up control is being performed, the routine proceeds to step 69, where the target main intake cam phase θmi_cmd is set to the aforementioned predetermined idle value θmi_idle.
[0199]
Next, the routine proceeds to step 70, where the catalyst warm-up value θmsi_cw of the target auxiliary intake cam phase is calculated by searching the table shown in FIG. 46 according to the execution time Tcat of the catalyst warm-up control described above. A value θmsiott (a value corresponding to a predetermined timing) in the figure shows the Otto phase value (= cam angle 90 deg) of the auxiliary intake cam phase θmsi at which the valve timing of the intake valve 6 is the same as that of the Otto intake cam. The same applies to the following description.
[0200]
Next, in step 71, the target auxiliary intake cam phase θmsi_cmd is set to the catalyst warm-up value θmsi_cw, and then in step 72, as in step 64, the target intake cam phase θssi # i_cmd (# i = # 2 to # 4) are all set to the value 0.
[0201]
Next, the routine proceeds to step 73 in FIG. 45, where the control input DUTY_mi to the main intake cam phase variable mechanism 60 is calculated according to the target main intake cam phase θmi_cmd and the main intake cam phase θmi. This control input DUTY_mi is calculated by the same algorithm as the control algorithm by the second SPAS controller 225 described above.
[0202]
Next, at step 74, the control input DUTY_msi to the auxiliary intake cam phase varying mechanism 70 is calculated by the control algorithm of the second SPAS controller 225. That is, the control input DUTY_msi is calculated by applying the above-described prediction algorithm of Equation (29), the identification algorithm of Equations (30) to (35), and the sliding mode control algorithm of Equations (36) to (41), respectively. To do.
[0203]
Next, at step 75, in accordance with the target intake cam phase θssi # i_cmd and the intake cam phase θssi # i calculated at step 72, the control input DUTY_ssi # i (# i = After calculating # 2 to # 4), the program is terminated. The control input DUTY_ssi # i is calculated by the same algorithm as the control algorithm used for calculating the control input DUTY_msi.
[0204]
Returning to FIG. 44, when the determination result of step 68 is NO and the engine is not being started and the catalyst warm-up control is not being performed, the routine proceeds to step 76 where the normal operation value θmi_drv of the target main intake cam phase is set to the required drive torque TRQ_eng and It is calculated by searching the map shown in FIG. 47 according to the engine speed NE.
[0205]
In the figure, predetermined values NE1 to NE3 of the engine speed NE are set so that the relationship NE1>NE2> NE3 is established, and this also applies to the following. In this map, the normal operation value θmi_drv is set to a more advanced value as the required drive torque TRQ_eng is larger or the engine speed NE is higher. This is because, as the engine load is higher, the main intake cam phase θmi is advanced and the opening / closing timing of the intake valve 6 is advanced, thereby appropriately ensuring the engine output.
[0206]
Next, at step 77, the target main intake cam phase θmi_cmd is set to the normal operation value θmi_drv, and then the routine proceeds to step 78, where the aforementioned basic value θmsi_base of the auxiliary intake cam phase (value corresponding to the target valve closing timing) is set. The calculation is performed by searching the table shown in FIG. 48 according to the required drive torque TRQ_eng.
[0207]
As shown in the figure, in this table, the basic value θmsi_base is set to a constant value on the late closing side in the range of TRQ_eng <TRQ_disc, that is, in the stratified combustion operation region of the engine 3. This is to stabilize the combustion state in the low load region where the stratified combustion operation is executed. Further, the basic value θmsi_base is set such that the degree of delayed closing becomes smaller as the required drive torque TRQ_eng is larger in the range of TRQ_disc ≦ TRQ_eng ≦ TRQott. This is because the larger the required drive torque TRQ_eng, the larger the amount of fuel blown back into the intake manifold due to the degree of late closing of the intake valve 6, which is avoided. Further, when TRQ_eng = TRQott (predetermined first and second load range threshold values), the basic value θmsi_base is set to the Otto phase value θmissiot (a value corresponding to a predetermined timing).
[0208]
Further, the basic value θmsi_base is set so that the degree of early closing increases as the required drive torque TRQ_eng increases in the range of TRQott <TRQ_eng <TRQ2, which increases the combustion efficiency by high expansion ratio cycle operation. Because.
[0209]
In the range of TRQ2 ≦ TRQ_eng <TRQ4, the basic value θmsi_base is set such that the degree of early closing of the intake valve 6 decreases as the required drive torque TRQ_eng increases. This is due to the following reason. That is, in a high load range such as TRQ2 ≦ TRQ_eng <TRQ4, as described later, the supercharging operation is limited in order to avoid the occurrence of knocking, so that the charging efficiency is reduced due to such supercharging operation limitation. In this state, if the degree of early closing of the intake valve 6 is controlled to be large, the generated torque is reduced. Therefore, in order to compensate for such a decrease in the generated torque, the degree of early closing of the intake valve 6 is set to be smaller as the required drive torque TRQ_eng is larger.
[0210]
In step 79 following step 78, the target auxiliary intake cam phase θmsi_cmd is calculated by the control algorithm of the first SPAS controller 221 described above. That is, the target auxiliary intake cam phase θmsi_cmd is obtained by applying the prediction algorithm of Expression (7), the identification algorithm of Expressions (8) to (13), and the sliding mode control algorithm of Expressions (15) to (21). Is calculated.
[0211]
Next, at step 80, the target intake cam phase θssi # i_cmd (# i = # 2 to # 4) is calculated by the control algorithm of the intake cam phase controller 230 described above. Next, after executing steps 73 to 75 in FIG. 45 as described above, the present program is terminated.
[0212]
Returning to FIG. 44, when the determination result in step 60 is YES and the variable intake valve driving device 40 or the variable exhaust valve driving device 90 is out of order, the process proceeds to step 81 to set the target main intake cam phase θmi_cmd to a predetermined value. After setting to the idle value θmi_idle, the routine proceeds to step 82, where the target auxiliary intake cam phase θmsi_cmd is set to a predetermined failure value θmsi_fs.
[0213]
Next, the routine proceeds to step 83, where the target intake cam phase θssi # i_cmd (# i = # 2 to # 4) is set to a value of 0 as in steps 64 and 72 described above. Thereafter, as described above, after executing steps 65 to 67 in FIG. 45, the program is terminated.
[0214]
Next, the throttle valve control process in step 5 described above will be described with reference to FIG. As shown in the figure, in this program, first, in step 120, it is determined whether or not the intake / exhaust valve failure flag F_VLVNG is "1". The determination result is NO, and the variable intake valve drive is determined. When both the device 40 and the variable exhaust valve driving device 90 are normal, the routine proceeds to step 121, where it is determined whether or not the engine start flag F_ENGSTART described above is “1”.
[0215]
If the determination result is YES and the engine is being started, the routine proceeds to step 122, where the target opening TH_cmd is set to a predetermined starting value THcmd_st. The predetermined starting value THcmd_st is set to a value slightly larger than an idle value THcmd_idle described later. Next, the routine proceeds to step 123, where the control input DUTY_th to the throttle valve mechanism 16 is calculated, and then this program is terminated. More specifically, the control input DUTY_th is calculated by searching a table (not shown) according to the target opening TH_cmd.
[0216]
On the other hand, when the determination result of step 121 is NO and the engine is not starting, the routine proceeds to step 124, where it is determined whether or not the above-mentioned catalyst warm-up flag F_CATHOT is “1”. If the determination result is YES and the catalyst warm-up control is being performed, the routine proceeds to step 125, where the catalyst warm-up value THcmd_ast of the target opening is shown in FIG. 50 in accordance with the execution time Tcat of the catalyst warm-up control described above. It is calculated by searching the table shown.
[0217]
A value THcmd_idle in the figure indicates an idle value used during idling. As shown in the figure, in this table, the catalyst warm-up value THcmd_ast is set to a larger value as the execution time Tcat becomes shorter until the execution time Tcat reaches a predetermined value Tcat1, and the execution time Tcat is increased. After reaching the predetermined value Tcat1, the idle value THcmd_idle is set.
[0218]
Next, the routine proceeds to step 126, where the target opening TH_cmd is set to the catalyst warm-up value THcmd_ast, and after executing step 123 as described above, this program is terminated.
[0219]
On the other hand, if the determination result in step 124 is NO and the engine is not being started and the catalyst warm-up control is not being performed, the routine proceeds to step 127 where the normal operation value THcmd_drv of the target opening is set to the required drive torque TRQ_eng and the engine speed NE. Accordingly, the map is calculated by searching the map shown in FIG.
[0220]
As shown in the figure, in this map, the normal operation value THcmd_drv is set to a larger value as the required drive torque TRQ_eng is higher or the engine speed NE is higher. This is because a larger amount of intake air is required to ensure a larger engine output as the load of the engine 3 is higher.
[0221]
Next, at step 128, the target opening TH_cmd is set to the normal operation value THcmd_drv, and then, as described above, after executing step 123, the program is terminated.
[0222]
On the other hand, if the determination result in step 120 is YES and the variable intake valve driving device 40 or the variable exhaust valve driving device 90 is malfunctioning, the routine proceeds to step 129, where the target opening degree failure value THcmd_fs is It is calculated by searching the map shown in FIG. 52 according to the degree AP and the engine speed NE. As shown in the figure, in this map, the failure value THcmd_fs is set to a larger value as the accelerator opening AP is larger or the engine speed NE is higher. This is due to the same reason as described in the calculation of the normal operation value THcmd_drv.
[0223]
Next, the routine proceeds to step 130, where the target opening TH_cmd is set to the failure value THcmd_fs. Next, as described above, after executing step 123, this program is terminated.
[0224]
By the above control processing, various control inputs DUTY_mi, DUTY_msi, DUTY_ssi # i, DUTY_me, DUTY_mse, DUTY_sse # i, and DUTY_th are any of a pulse signal, a current signal, and a voltage signal having a duty ratio according to the calculation result. Set to one.
[0225]
The operation when the engine control as described above is executed will be described for each range of the required drive torque TRQ_eng (L1) to (L6) below with reference to FIG.
[0226]
(L1) TRQ_idle ≦ TRQ_eng <TRQ_disc range
In this range, the auxiliary intake cam phase θmsi is controlled to a substantially constant value on the late closing side by setting the basic value θmsi_base described above. Further, by controlling the intake air amount so that it is not throttled by the throttle valve 17, the intake pipe absolute pressure PBA is controlled to a substantially constant value slightly lower than the atmospheric pressure Patm. Further, the cylinder intake air amount Gcyl is substantially constant. System It is controlled. Further, the main fuel injection rate Rt_Pre is set to the maximum value Rtmax, the target air-fuel ratio KCMD is set to the above-described extremely lean region value, and the stratified combustion operation is performed.
[0227]
(L2) TRQ_disc ≦ TRQ_eng ≦ TRQ1 range
In this range, the setting of the basic value θmsi_base described above causes the auxiliary intake cam phase θmsi to be controlled to a value on the delay closing side that is considerably larger than the value in the range (L1), and the required drive torque TRQ_eng is larger. The degree of late closing is controlled so as to become smaller. Further, the cylinder intake air amount Gcyl is controlled to a value smaller than the value in the range of (L1), and is controlled to be larger as the required drive torque TRQ_eng is larger. Further, the target air-fuel ratio KCMD is controlled so as to maintain the above-described lean region value on the rich side with respect to the value in the range (L1), and the intake pipe absolute pressure PBA and the main fuel injection rate Rt_Pre are both Control is performed so as to hold the value in the range of (L1).
[0228]
(L3) Range of TRQ1 <TRQ_eng ≦ TRQott
In this range, the auxiliary intake cam phase θmsi is controlled in the same tendency as in the range (L2) by setting the basic value θmsi_base described above. In particular, when TRQ_eng = TRQott, the auxiliary intake cam phase θmsi is controlled to become the Otto phase value θmsiot. That is, the engine 3 is operated in an Otto cycle. Further, the target air-fuel ratio KCMD and the cylinder intake air amount Gcyl are also controlled to have the same tendency as in the range (L2). Further, in this range, the supercharging operation by the turbocharger device 10 is executed, whereby the intake pipe absolute pressure PBA is controlled to be higher as the required drive torque TRQ_eng is larger. Further, the main fuel injection rate Rt_Pre is controlled to be a smaller value as the required drive torque TRQ_eng is larger. That is, the control is performed so that the fuel injection amount TOUT_sub of the auxiliary fuel injection valve 15 becomes larger as the required drive torque TRQ_eng is larger. This is for obtaining the cooling effect of the intake air by the fuel vaporization cooling device 12.
[0229]
(L4) Range of TRQott <TRQ_eng <TRQ2
In this range, the auxiliary intake cam phase θmsi is controlled such that the degree of early closing increases as the required drive torque TRQ_eng increases. This is because the combustion efficiency is increased by the high expansion ratio cycle operation as described above. Further, the cylinder intake air amount Gcyl, the target air-fuel ratio KCMD, the main fuel injection rate Rt_Pre, and the intake pipe absolute pressure PBA are controlled so as to show the same tendency as in the range (L3). In particular, the intake pipe absolute pressure PBA is controlled to be higher as the required drive torque TRQ_eng is larger, as described above. This is because when the auxiliary intake cam phase θmsi is controlled to the early closing side, the generated torque is reduced, so that the charging efficiency is increased by supercharging and the generated torque is increased for the purpose of compensation.
[0230]
(L5) TRQ2 ≦ TRQ_eng <TRQ4 range
In this range, the auxiliary intake cam phase θmsi is controlled such that the degree of early closing decreases as the required drive torque TRQ_eng increases, and as a result, the effective compression volume increases. As described above, when the charging efficiency is reduced due to the limitation of the supercharging operation and the degree of early closing of the intake valve 6 is controlled to be large, the generated torque is reduced. This is to compensate for a decrease in generated torque by controlling the auxiliary intake cam phase θmsi.
[0231]
The intake pipe absolute pressure PBA is controlled to maintain a constant value in the range of TRQ2 ≦ TRQ_eng ≦ TRQ3, and becomes lower as the required drive torque TRQ_eng is larger in the range of TRQ3 <TRQ_eng <TRQ4. To be controlled. Further, the main fuel injection rate Rt_Pre is controlled to be a smaller value as the required drive torque TRQ_eng is larger, similarly to the range of (L3) above. As described above, in the range of (L5), as the required drive torque TRQ_eng is larger, the supercharging operation by the turbocharger device 10 is restricted, and at the same time, the cooling effect by the fuel vaporization cooling device 12 is increased. Thus, the occurrence of knocking can be avoided without performing retard control of the ignition timing. In the case of a conventional engine with a turbocharger device, knocking occurs within the range of (L5) unless ignition timing retard control is executed.
[0232]
(L6) Range of TRQ4 ≦ TRQ_eng
In this range, since it is an extremely high load region, the restriction of the supercharging operation by the turbocharger device 10 and the cooling effect by the fuel vaporization cooling device 12 cannot avoid the occurrence of knocking. Is executed. That is, the target air-fuel ratio KCMD is controlled to become richer as the required drive torque TRQ_eng is larger. At the same time, the auxiliary intake cam phase θmsi is controlled to become the Otto phase value θmsiott, the cylinder intake air amount Gcyl is controlled to be substantially constant, the main fuel injection rate Rt_Pre is controlled to the minimum value Rtmin, The intake pipe absolute pressure PBA is controlled to maintain a substantially constant value.
[0233]
As described above, according to the control device 1 of the present embodiment, the auxiliary intake cam phase θmsi is controlled to be the basic value θmsi_base of the target auxiliary intake cam phase, and the boost pressure Pc is set to the target boost pressure. It is controlled to be Pc_cmd. In particular, in the conventional control, the basic value θmsi_base of the target auxiliary intake cam phase is the required drive in the range of TRQ2 <TRQ_eng <TRQ4, which is a high load region in which knocking occurs unless ignition timing retard control is executed. The higher the torque TRQ_eng, the smaller the degree of early closing. That is, the basic value θmsi_base is set such that the effective compression volume increases as the load increases. Further, the target boost pressure Pc_cmd is set to a constant value in the range of TRQ2 <TRQ_eng ≦ TRQ3, and is set to a smaller value in the range of TRQ3 <TRQ_eng <TRQ4 as the required drive torque TRQ_eng is larger. That is, the supercharging pressure Pc is controlled so as to be suppressed (maintained or decreased) without increasing in the range of TRQ2 <TRQ_eng <TRQ4, and in the range of TRQ3 <TRQ_eng <TRQ4, the required drive torque. Control is performed such that the larger TRQ_eng is, the smaller TRQ_eng is. As described above, the auxiliary intake cam phase θmsi is controlled so that the effective compression volume increases as the load of the engine 3 increases, so that the required engine output is ensured even in a high load range. It is possible to control the supply pressure Pc to be suppressed (maintained or decreased) without increasing it, thereby suppressing the increase in intake air temperature. Accordingly, the limit at which knocking starts to occur can be further expanded without performing the retard control of the ignition timing, and both the combustion efficiency and the engine output can be improved. As a result, drivability and merchantability can be improved.
[0234]
Further, in the range of TRQott <TRQ_eng <TRQ4, the engine 3 is operated in a high expansion ratio cycle by the early closing of the intake valve 6, so that the case where the intake valve 6 is driven by the Otto intake cam or the case where it is delayed closed is different. Blowing back of fuel into the intake manifold can be avoided, and an increase in the amount of staying fuel in the intake manifold and an increase in the amount of fuel adhering to the inner wall of the intake manifold due to this can be avoided. Thereby, the control accuracy of the air-fuel ratio control and the torque control in the high load region can be improved, and in particular, the control accuracy in the transient operation state can be remarkably improved. As a result, exhaust gas characteristics and operability can be improved, and the life of an intake system device such as the intake valve 6 can be extended.
[0235]
Further, in the range of TRQ_idle ≦ TRQ_eng <TRQott, that is, in a low load range, the engine 3 is operated at a high expansion ratio cycle due to the late closing of the intake valve 6, so that it is not necessary to throttle the intake air amount by the throttle valve 17. In addition, while avoiding the pumping loss, the intake air amount can be set to an appropriate value according to the low load, and the fuel efficiency can be improved. In addition to this, unlike the case where the high expansion ratio cycle operation is realized by the early closing of the intake valve 6, the occurrence of fuel liquefaction in the cylinder and the combustion in the low load range when the intake air temperature and the engine temperature are low Destabilization of the state can be avoided, and a good combustion state can be ensured.
[0236]
In addition to this, since the variable intake valve drive device 40 is configured by a hydraulic drive type, for example, a variable intake valve drive device of a type that drives the valve body of the intake valve 6 by electromagnetic force of a solenoid is provided. Compared with the case where it is used, the intake valve 6 can be reliably opened and closed even in a higher load range, power consumption can be reduced, and operating noise of the intake valve 6 can be reduced.
[0237]
In the variable intake valve driving device 40, not only the closing timing of the intake valve 6 but also the valve lift amount can be changed in the range of the auxiliary intake cam phase θmsi = 120 to 180 deg. By controlling the lift amount to a smaller value side, the flow rate of the intake air flowing into the combustion chamber can be increased and the in-cylinder flow can be increased. Thereby, combustion efficiency can be improved.
[0238]
Further, a combination of the intake valve drive mechanism 50 including the main / sub intake cams 43, 44, the main / sub intake cam shafts 41, 42, the link mechanism 50 and the rocker arm 51, and the auxiliary intake cam phase variable mechanism 70, A configuration in which the intake cam phase θmsi can be freely changed, that is, a configuration in which the valve closing timing and the valve lift amount of the intake valve 6 can be freely changed can be realized.
[0239]
In the range of TRQ1 <TRQ_eng <TRQ4, fuel is injected from the auxiliary fuel injection valve 15 toward the lipophilic film plate 14 in the fuel vaporization cooling device 12 and supplied into the cylinder. At that time, the fuel is thinned on the lipophilic film of the lipophilic film plate 14 and is vaporized by the heat of the intake air whose temperature has been increased by the supercharging operation by the turbocharger device 10 to generate an air-fuel mixture. The intake air is cooled by heat. In this way, it is possible to obtain an intake air cooling action with the generation of the air-fuel mixture, thereby further expanding the limit at which knocking starts without performing ignition timing retard control. .
[0240]
Further, in the range of TRQ1 <TRQ_eng <TRQ4, the main fuel injection rate Rt_Pre, which is the ratio of the main fuel injection amount TOUT_main injected from the main fuel injection valve 4 to the total fuel injection amount TOUT, is the required drive torque TRQ_eng. A larger value is set to a smaller value. In other words, the ratio of the sub fuel injection amount TOUT_sub injected from the sub fuel injection valve 15 [(100−Rt_Pre) as the required drive torque TRQ_eng is large and the load is high and the degree of increase in the intake air temperature due to supercharging increases. / 100] is set to a larger value, the intake air cooling action by the fuel evaporative cooling device 12 can be obtained more effectively and appropriately according to the degree of increase in the intake air temperature.
[0241]
In the embodiment, the intake valve 6 is controlled to the slow closing side in the range of TRQ_eng <TRQott (predetermined third load range). Instead, the intake valve 6 is controlled to the early closing side in this range. May be. In that case, in the ECU 2 (valve lift amount determining means), the above-described auxiliary intake cam phase θmsi is set in the range of θmsi = 120 to 180 deg according to the required drive torque TRQ_eng. What is necessary is just to control to a value smaller than the maximum value. More specifically, the valve lift amount may be controlled to be smaller as the required drive torque TRQ_eng is smaller. When such control is performed, combustion can be accelerated by increasing the flow rate of the intake air flowing into the combustion chamber and increasing the in-cylinder flow. As a result, even when the intake air temperature or the engine temperature is low, the liquefaction of fuel as described above due to the early closing of the intake valve 6 can be avoided in the low load region. Thereby, the combustion state can be stabilized.
[0242]
The embodiment is an example in which the turbocharger device 10 is used as a supercharging device. However, the supercharger is not limited to this, and any device that can perform a supercharging operation of intake air may be used. For example, a mechanical supercharger such as a supercharger may be used.
[0243]
The embodiment is an example in which the main fuel injection valve 4 is used as a fuel injection valve for direct injection, but this may be provided on the intake pipe 8 side and used as a port injection type fuel injection valve.
[0244]
Furthermore, when the auxiliary intake cam phase variable mechanism 70 does not require high responsiveness (for example, in the above-described intake valve control process, the intake valve 6 need only be controlled to one of the late closing side and the early closing side). Instead of the hydraulic piston mechanism 73 and the motor 74, the hydraulic pump 63 and the electromagnetic valve mechanism 64 may be used as in the main intake cam phase variable mechanism 60. In that case, the control device 1 may be configured as shown in FIG.
[0245]
As shown in the figure, the control device 1 is provided with a DUTY_msi calculator 300 and a throttle valve opening controller 301 instead of the DUTY_th calculator 200 and the auxiliary intake cam phase controller 220 in the embodiment. The DUTY_msi calculation unit 300 searches the table according to the required drive torque TRQ_eng to calculate the target auxiliary intake cam phase θmsi_cmd, and then searches the table according to the calculated target auxiliary intake cam phase θmsi_cmd. Thus, the control input DUTY_msi is calculated. Further, the throttle valve opening controller 301 calculates the target opening TH_cmd by the same control algorithm as that of the first SPAS controller 221 described above according to the cylinder intake air amount Gcyl and the target intake air amount Gcyl_cmd. The control input DUTY_th is calculated by the same control algorithm as that of the second SPAS controller 225 according to the opening TH_cmd. When configured as described above, even when the responsiveness of the auxiliary intake cam phase varying mechanism 70 is low, the auxiliary intake cam phase θmsi can be appropriately controlled while avoiding the influence.
[0246]
The embodiment is an example in which both the first SPAS controller 221 and the second SPAS controller 225 are provided as the auxiliary intake cam phase controller 220, but instead of this, an auxiliary intake cam phase controller 220 having only the first SPAS controller 221 is used. Also good. In this case, the control input DUTY_msi may be calculated according to the target auxiliary intake cam phase θmsi_cmd calculated by the first SPAS controller 221 by referring to a table, for example.
[0247]
Furthermore, the control device of the present invention is not limited to the internal combustion engine for vehicles of the embodiment, but can be applied to various internal combustion engines for ships and the like.
[0248]
【The invention's effect】
As described above, according to the control device for an internal combustion engine of the present invention, it is possible to improve both combustion efficiency and engine output while suppressing the occurrence of knocking in a high load range, thereby improving the drivability and Productivity can be improved.
[Brief description of the drawings]
FIG. 1 is a diagram showing a schematic configuration of an internal combustion engine to which a control device according to an embodiment of the present invention is applied.
FIG. 2 is a diagram showing a schematic configuration of a variable intake valve driving device and a variable exhaust valve driving device for an internal combustion engine.
FIG. 3 is a diagram showing a schematic configuration of a control device.
FIG. 4 is a diagram showing a schematic configuration of a fuel vaporization cooling apparatus.
FIG. 5 is a schematic diagram illustrating a schematic configuration of a variable intake valve driving device and a variable exhaust valve driving device in a plan view.
FIG. 6 is a diagram showing a schematic configuration of an intake valve drive mechanism of a variable intake valve drive device.
FIG. 7 is a diagram showing a schematic configuration of a main intake cam phase varying mechanism.
FIG. 8 is a diagram showing a schematic configuration of an auxiliary intake cam phase varying mechanism.
FIG. 9 is a view showing a schematic configuration of a modified example of the auxiliary intake cam phase varying mechanism.
FIG. 10 is a diagram showing a schematic configuration of a phase change mechanism between intake cams.
FIG. 11 is a diagram for explaining cam profiles of a main intake cam and a sub intake cam.
FIGS. 12A and 12B are diagrams showing an operation state of the intake valve drive mechanism in the case of the auxiliary intake cam phase θmsi = 0 deg; and FIG. 12B shows a valve lift curve for explaining the operation of the intake valve.
FIGS. 13A and 13B are views showing an operation state of the intake valve drive mechanism in the case of the auxiliary intake cam phase θmsi = 90 deg; and FIG. 13B shows a valve lift curve for explaining the operation of the intake valve.
14A is a view showing an operation state of an intake valve drive mechanism in the case of a sub intake cam phase θmsi = 120 deg, and FIG. 14B is a view showing a valve lift curve for explaining the operation of the intake valve.
FIGS. 15A and 15B are diagrams showing an operation state of the intake valve driving mechanism and a valve lift curve for explaining the operation of the intake valve when the auxiliary intake cam phase θmsi = 180 deg.
FIG. 16 is a diagram showing changes in valve lift amount and valve timing for explaining the operation of the intake valve when the auxiliary intake cam phase θmsi is changed from 120 deg to 180 deg.
FIG. 17 is a view for explaining cam profiles of a main exhaust cam and a sub exhaust cam.
FIG. 18 is a diagram showing a valve lift curve and the like for explaining the operation of the exhaust valve when the sub exhaust cam phase θmse = 0 deg.
FIG. 19 is a view showing a valve lift curve and the like for explaining the operation of the exhaust valve when the sub exhaust cam phase θmse = 45 deg.
FIG. 20 is a view showing a valve lift curve and the like for explaining the operation of the exhaust valve when the sub exhaust cam phase θmse = 90 deg.
FIG. 21 is a view showing a valve lift curve and the like for explaining the operation of the exhaust valve when the sub exhaust cam phase θmse = 150 deg.
FIG. 22 is a block diagram showing a configuration for controlling the throttle valve mechanism, the auxiliary intake cam phase varying mechanism, and the intake cam phase varying mechanism in the control device;
FIG. 23 is a block diagram showing a schematic configuration of an auxiliary intake cam phase controller.
FIG. 24 is a diagram illustrating a calculation formula of a cylinder intake air amount Gcyl and a formula of a prediction algorithm of a state predictor in the first SPAS controller.
FIG. 25 is a diagram illustrating a mathematical expression of an identification algorithm of an on-board identifier in the first SPAS controller.
FIG. 26 is a diagram showing mathematical expressions of a sliding mode control algorithm of the sliding mode controller in the first SPAS controller.
FIG. 27 is a diagram illustrating a mathematical expression for explaining a derivation method of Expression (19) in FIG. 26;
FIG. 28 is a diagram illustrating a phase plane and a switching line for explaining a sliding mode control algorithm.
FIG. 29 is a diagram illustrating an example of the convergence behavior of the tracking error Es when the switching function setting parameter Ss is changed in the sliding mode controller.
FIG. 30 is a block diagram showing a schematic configuration of a second SPAS controller.
FIG. 31 is a diagram illustrating a mathematical expression of a prediction algorithm of a state predictor in the second SPAS controller.
FIG. 32 is a diagram illustrating mathematical expressions of an identification algorithm of an on-board identifier in the second SPAS controller.
FIG. 33 is a diagram illustrating mathematical expressions of a sliding mode control algorithm of the sliding mode controller in the second SPAS controller.
FIG. 34 is a flowchart showing main control contents of an engine control process.
FIG. 35 is a flowchart showing a fuel control process.
FIG. 36 is a diagram showing an example of a map used for calculation of required drive torque TRQ_eng.
FIG. 37 is a flowchart showing a calculation process of a cylinder intake air amount Gcyl and a target intake air amount Gcyl_cmd.
FIG. 38 is a diagram illustrating an example of a map used for calculating a basic value Gcyl_cmd_base of a target intake air amount.
FIG. 39 is a diagram illustrating an example of a table used for calculating an air-fuel ratio correction coefficient Kgcyl_af.
FIG. 40 is a diagram showing an example of a table used for calculating a main fuel injection rate Rt_Pre.
FIG. 41 is a flowchart showing a supercharging pressure control process.
FIG. 42 is a diagram illustrating an example of a table used for calculating a basic value Dut_wg_base of a control input to the wastegate valve.
FIG. 43 is a diagram illustrating an example of a table used for calculating a target boost pressure Pc_cmd.
FIG. 44 is a flowchart showing intake valve control processing;
45 is a flowchart showing a continuation of FIG. 44. FIG.
FIG. 46 is a diagram illustrating an example of a table used for calculating a catalyst warm-up value θmsi_cw of a target auxiliary intake cam phase.
FIG. 47 is a diagram showing an example of a table used for calculating a normal operation value θmi_drv of the target main intake cam phase.
FIG. 48 is a diagram illustrating an example of a map used for calculating a basic value θmsi_base of a target auxiliary intake cam phase.
FIG. 49 is a flowchart showing a throttle valve control process.
FIG. 50 is a diagram illustrating an example of a table used for calculating a catalyst warm-up value THcmd_ast of a target opening degree.
FIG. 51 is a diagram illustrating an example of a map used for calculating a normal operation value THcmd_drv of a target opening.
FIG. 52 is a diagram illustrating an example of a map used for calculating a failure value THcmd_fs for a target opening degree.
FIG. 53 is a diagram showing an operation example of engine control by the control device.
FIG. 54 is a block diagram showing a schematic configuration of a modified example of the control device.
[Explanation of symbols]
1 Control device
2 ECU (load detection means, target supercharging pressure setting means, supercharging control means,
Target valve closing timing setting means, valve timing control means,
Fuel injection ratio setting means)
3 Internal combustion engine
# 1 to # 4 1st to 4th cylinder
4 Main fuel injection valve (first fuel injection valve)
6 Intake valve
8 Intake pipe (intake passage)
10 Turbocharger device (supercharger)
12 Fuel evaporative cooling system (intake air cooling system)
14 Lipophilic film board
15 Sub fuel injection valve (second fuel injection valve)
20 Crank angle sensor (load detection means)
35 Accelerator opening sensor (load detection means)
40 Variable intake valve drive device (Variable valve timing device)
41 Main intake camshaft (first intake camshaft)
42 Secondary intake camshaft (second intake camshaft)
43 Main intake cam (first intake cam)
44 Secondary intake cam (second intake cam)
51 Intake rocker arm
51c pin (rotating fulcrum)
70 Sub-intake cam phase variable mechanism (intake cam phase variable mechanism)
221 First SPAS controller 221 (valve closing timing setting means)
θmsi Sub intake cam phase (relative phase between first and second intake camshafts)
θmsi_base Basic value of target auxiliary intake cam phase (value corresponding to target valve closing timing)
θmsiot Otto phase value (value corresponding to a predetermined timing)
TRQ_eng Required drive torque (parameter indicating load)
TRQ2 Predetermined value (predetermined load)
TRQ3 Predetermined value (predetermined high load range threshold)
TRQ4 Predetermined value (predetermined high load range threshold)
TRQott Predetermined value (value that prescribes a predetermined low load range)
TRQ_idle Predetermined value (value that prescribes a predetermined low load range)
Pc boost pressure
Pc_cmd Target supercharging pressure
Psd hydraulic pressure
TOUT Total fuel injection amount
Rt_Pre main fuel injection rate (value defining the first fuel injection rate and the second fuel injection rate)

Claims (10)

Control of the internal combustion engine that controls the supercharging pressure of the intake air via a supercharger provided in the intake passage and variably controls the closing timing with respect to the opening timing of the intake valve via a valve timing variable device A device,
Load detecting means for detecting a load of the internal combustion engine;
Target boost pressure setting means for setting a target boost pressure that is a target of the boost pressure control according to the detected load of the internal combustion engine;
A supercharging control means for controlling the supercharger according to the set target supercharging pressure;
When the load of the detected said internal combustion engine is higher predetermined high-load range than the predetermined load, the target valve-closing timing as a target valve closing timing control of the intake valve, the expansion ratio in the combustion cycle A target valve closing timing setting means for setting the timing such that the expansion ratio approaches the compression ratio as the compression ratio exceeds and the detected load of the internal combustion engine is higher;
Valve timing control means for controlling the valve timing variable device according to the set target valve closing timing;
A control device for an internal combustion engine, comprising: The target valve closing timing setting means sets the intake valve closing timing when the load of the internal combustion engine is in the predetermined high load range, and the expansion ratio during the combustion cycle of the internal combustion engine is equal to the compression ratio. 2. The control device for an internal combustion engine according to claim 1, wherein the timing is set earlier than a predetermined timing. When the load of the internal combustion engine is in a predetermined low load region lower than the predetermined high load region, the target valve closing timing setting means sets the valve closing timing of the intake valve during the combustion cycle of the internal combustion engine. 3. The control device for an internal combustion engine according to claim 1, wherein the expansion ratio is set to a timing at which the expansion ratio exceeds the compression ratio. The target valve closing timing setting means sets the intake valve closing timing when the load of the internal combustion engine is in the predetermined low load region, and the expansion ratio during the combustion cycle of the internal combustion engine becomes equal to the compression ratio. 4. The control apparatus for an internal combustion engine according to claim 3, wherein the timing is set to a timing later than the predetermined timing. The target boost pressure setting means sets the target boost pressure to a smaller value as the load of the internal combustion engine is larger when the load of the internal combustion engine is in the predetermined high load range. The control device for an internal combustion engine according to any one of claims 1 to 4. The valve timing variable device is composed of a hydraulically driven valve timing variable device driven by supply of hydraulic pressure,
6. The control apparatus for an internal combustion engine according to claim 1, wherein the valve timing control means controls a hydraulic pressure supplied to the hydraulically driven valve timing variable device. The control apparatus for an internal combustion engine according to any one of claims 1 to 6, wherein the variable valve timing device is configured to be capable of changing a valve lift amount of the intake valve. The valve timing variable device is:
An intake rocker arm that is pivotally provided and drives the intake valve to open and close by the rotation;
First and second intake camshafts rotating at the same rotational speed;
An intake cam phase variable mechanism that changes a relative phase between the first and second intake camshafts;
A first intake cam that is provided on the first intake camshaft and rotates around the rotation center by rotating in accordance with the rotation of the first intake camshaft;
A second intake cam that is provided on the second intake camshaft and moves according to the rotation of the second intake camshaft to move the rotation center of the intake rocker arm;
The control apparatus for an internal combustion engine according to claim 7, further comprising: The internal combustion engine
A first fuel injection valve that injects fuel and supplies the fuel into the cylinder;
An intake air cooling device that is provided downstream of the supercharger in the intake passage and cools the intake air from the supercharger,
The intake air cooling device
A lipophilic film plate having a lipophilic film having affinity for fuel formed on the surface;
A second fuel injection valve that injects fuel toward the lipophilic film plate and supplies the fuel into the cylinder;
The control apparatus for an internal combustion engine according to any one of claims 1 to 8, characterized by comprising: Of the amount of fuel to be supplied to the cylinder, a first fuel injection ratio to be supplied by the first fuel injection valve and a second fuel injection ratio to be supplied by the second fuel injection valve are respectively set in the internal combustion engine. The fuel injection ratio setting means that sets the second fuel injection ratio according to a load and sets the second fuel injection ratio to a larger value as the load of the internal combustion engine is larger is provided. Control device for internal combustion engine.

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