JP2015124716A - Fuel supply device of internal combustion engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a fuel supply device of an internal combustion engine capable of improving control accuracy of fuel pressure and capable of reducing power consumption of a low-pressure pump.SOLUTION: A low-pressure pump discharge amount is controlled according to detected operation states NE, TREQ of an internal combustion engine (step 2), and a high-pressure pump is controlled so that a fuel pressure PF to be a fuel pressure on a fuel injection valve side with respect to the high-pressure pump becomes a set target fuel pressure PFOBJ (steps 8, 10, 12). Also, the low-pressure pump discharge amount is suppressed so that the maximum value of the set target fuel pressure PFOBJ becomes an attainable amount by control of the high-pressure pump.

Description

  The present invention relates to a fuel supply device for an internal combustion engine that supplies fuel to a fuel injection valve of the internal combustion engine.

  Conventionally, as this type of fuel supply device, for example, one disclosed in Patent Document 1 is known. The fuel supply device includes an electric low pressure pump, a piston type high pressure pump, and a delivery pipe connected to the high pressure pump and the fuel injection valve. In the fuel supply device, the fuel in the fuel tank is sucked by the low pressure pump and discharged to the high pressure pump side, and the fuel discharged from the low pressure pump is discharged to the delivery pipe side in a pressurized state by the high pressure pump. Is done. Further, the amount of fuel discharged from the high-pressure pump (hereinafter referred to as “high-pressure pump discharge amount”) is controlled so that the fuel pressure that is the pressure of the fuel in the delivery pipe becomes the target fuel pressure.

  Further, the amount of fuel discharged from the low-pressure pump (hereinafter referred to as “low-pressure pump discharge amount”) is controlled in accordance with the operating state of the internal combustion engine. Specifically, the discharge amount of the low-pressure pump is controlled to a smaller value as the rotational speed of the internal combustion engine is lower and as the load on the internal combustion engine is smaller. Furthermore, the low-pressure pump discharge amount is controlled to a smaller value as the high-pressure pump discharge amount is smaller.

JP2009-221906 (pages 5-7)

  The discharge amount of the low-pressure pump does not change immediately with respect to the change in the power supplied to the low-pressure pump, but changes with a certain delay. As described above, in the conventional fuel supply device, the low-pressure pump discharge amount is changed each time according to the rotation speed and load of the internal combustion engine and the high-pressure pump discharge amount. For this reason, for example, when the target fuel pressure rapidly increases, the low pressure pump to the high pressure pump with respect to the high pressure pump discharge amount necessary for controlling the fuel pressure (pressure of fuel in the delivery pipe) to become the target fuel pressure. There is a possibility that the amount of fuel discharged will be insufficient. In that case, even if the discharge amount of the high-pressure pump is controlled, the fuel pressure cannot be controlled to the target fuel pressure, and the control accuracy is lowered.

  The present invention has been made to solve the above-described problems, and provides a fuel supply device for an internal combustion engine that can improve the control accuracy of the fuel pressure and reduce the power consumption of the low-pressure pump. The purpose is to provide.

  In order to achieve the above object, the invention according to claim 1 is a fuel supply device 1 for an internal combustion engine that supplies fuel to a fuel injection valve 4 of the internal combustion engine 3, and injects fuel in a fuel tank 11. An electric low-pressure pump 12 that discharges to the valve 4 side, and an operating state detecting means that detects the operating state of the internal combustion engine 3 (in the embodiment (hereinafter, the same applies in this section)) ECU 2, crank angle sensor 32, accelerator opening The low pressure pump discharge amount QLPO, which is the amount of fuel discharged from the low pressure pump 12, is controlled according to the sensor 33, step 1) and the detected operating state of the internal combustion engine 3 (engine speed NE, required torque TREQ). The low pressure pump control means (ECU 2, step 2) and the high pressure pump 2 that sucks the fuel discharged from the low pressure pump 12 and discharges it to the fuel injection valve 4 side in a pressurized state. Target fuel pressure setting means (ECU 2, step 6) for setting a target fuel pressure PFOBJ, which is a target value of the fuel pressure on the fuel injection valve 4 side of the high-pressure pump 20, in accordance with the operating state of the internal combustion engine 3, and high pressure High-pressure pump control means (ECU 2, steps 8, 10, 12) for controlling the high-pressure pump 20 so that the fuel pressure PF that is the fuel pressure on the fuel injection valve 4 side of the pump 20 becomes the set target fuel pressure PFOBJ; The low-pressure pump control means suppresses the low-pressure pump discharge amount QLPO so that the maximum value of the set target fuel pressure PFOBJ can be realized by the control of the high-pressure pump 20 by the high-pressure pump control means. To do.

  According to this configuration, the fuel discharged from the electric low-pressure pump is discharged to the fuel injection valve side while being pressurized by the high-pressure pump. Further, the high pressure pump is controlled by the high pressure pump control means so that the fuel pressure that is the fuel pressure on the fuel injection valve side of the high pressure pump becomes the target fuel pressure set according to the detected operating state of the internal combustion engine. Is done. Furthermore, the low-pressure pump discharge amount, which is the fuel discharge amount from the low-pressure pump, is controlled by the low-pressure pump control means in accordance with the operating state of the internal combustion engine. In this case, unlike the above-described conventional case, the discharge amount of the low-pressure pump is suppressed so that the maximum value of the target fuel pressure can be realized by the control of the high-pressure pump by the high-pressure pump control means. Even in this case, the fuel can be sufficiently discharged from the low-pressure pump to the high-pressure pump, whereby the fuel pressure can be appropriately controlled so as to become the target fuel pressure, and the control accuracy can be improved.

  In addition, the low-pressure pump discharge amount is not always controlled to a large constant value, but is controlled according to the operating state of the internal combustion engine, and the maximum value of the target fuel pressure is suppressed to a realizable size. The power consumption of the low pressure pump can be reduced according to the operating state. For this reason, for example, when the generator for charging the power source of the low-pressure pump is driven by the internal combustion engine, the fuel consumption of the internal combustion engine can be improved by reducing the power consumption of the low-pressure pump described above.

  According to a second aspect of the present invention, in the fuel supply device 1 for the internal combustion engine according to the first aspect, the low-pressure pump control means is a predetermined plurality defined by the number of revolutions of the internal combustion engine 3 as the operating state of the internal combustion engine 3. The low-pressure pump discharge amount QLPO is controlled so as to increase as the rotational speed region to which the detected rotational speed of the internal combustion engine 3 (engine rotational speed NE) belongs is higher. In addition, in each of the plurality of rotation speed regions, the maximum value of the target fuel pressure PFOBJ set in each rotation speed region can be realized by the control of the high pressure pump 20 by the high pressure pump control means (first discharge amount QLPO1, 2 discharge amount QLPO2 and third discharge amount QLPO3), the low pressure pump discharge amount QLPO is suppressed.

  In general, the higher the number of revolutions of the internal combustion engine, the shorter the period during which fuel can be injected. Therefore, a higher fuel pressure is required, and therefore a larger amount of low-pressure pump discharge is required. According to the above-described configuration, the internal combustion engine as the operating state of the internal combustion engine in which the low-pressure pump discharge amount is detected in a plurality of predetermined rotational speed regions defined by the rotational speed of the internal combustion engine as the operating state of the internal combustion engine. The higher the rotation speed region to which the rotation number belongs, the higher the rotation speed region. Further, in each of the plurality of rotation speed regions, the discharge amount of the low pressure pump is suppressed to a value that can be achieved by controlling the high pressure pump by the high pressure pump control means with the maximum value of the target fuel pressure set in each rotation speed region. The As described above, the discharge amount of the low-pressure pump can be appropriately controlled according to the rotational speed of the internal combustion engine when the rotational speed of the internal combustion engine is in any of the plurality of rotational speed regions, and the control accuracy of the fuel pressure can be improved. Can be improved.

  In addition, the discharge amount of the low-pressure pump is not controlled to a value that can achieve the maximum value of the target fuel pressure set in the whole of the plurality of rotation speed regions in each of the plurality of rotation speed regions. The maximum value of the target fuel pressure set in several regions is suppressed to a realizable value. Therefore, the power consumption of the low pressure pump can be appropriately reduced according to the rotational speed of the internal combustion engine.

  According to a third aspect of the present invention, in the fuel supply device 1 for the internal combustion engine according to the first aspect, the low-pressure pump control means includes a plurality of predetermined plural prescribed by the load of the internal combustion engine 3 as the operating state of the internal combustion engine 3. In the load region, the low-pressure pump discharge amount QLPO is controlled so as to increase as the load region to which the detected load (requested torque TREQ) of the internal combustion engine 3 belongs is higher, and a plurality of loads In each region, the maximum value of the target fuel pressure PFOBJ set in each load region can be realized by the control of the high-pressure pump 20 by the high-pressure pump control means (first discharge amount QLPO1, second discharge amount QLPO2, third The low-pressure pump discharge amount QLPO is suppressed to the discharge amount QLPO3).

  In general, as the load on the internal combustion engine increases, the amount of fuel injected by the fuel injection valve increases, so that a higher fuel pressure is required, and therefore a larger amount of low-pressure pump discharge is required. According to the configuration described above, the load of the internal combustion engine as the operating state of the internal combustion engine in which the low-pressure pump discharge amount is detected in a plurality of predetermined load regions defined by the load of the internal combustion engine as the operating state of the internal combustion engine. It is controlled to increase as the load region to which is belongs is the region on the high load side. Further, in each of the plurality of load regions, the low-pressure pump discharge amount is suppressed to a value that can be achieved by controlling the high-pressure pump by the high-pressure pump control means with the maximum value of the target fuel pressure set in each load region. As described above, when the load of the internal combustion engine is in any of the plurality of load regions, the discharge amount of the low-pressure pump can be appropriately controlled according to the load of the internal combustion engine, thereby improving the control accuracy of the fuel pressure. Can do.

  Further, the discharge amount of the low pressure pump is not controlled to a value at which the maximum value of the target fuel pressure set in the whole of the plurality of load regions can be realized in each of the plurality of load regions. The maximum value of the set target fuel pressure is suppressed to a realizable value. Therefore, the power consumption of the low pressure pump can be appropriately reduced according to the load of the internal combustion engine.

  According to a fourth aspect of the present invention, in the fuel supply device 1 for an internal combustion engine according to any one of the first to third aspects, the amount of fuel discharged from the high pressure pump 20 is adjusted on the suction side of the high pressure pump 20. Are provided with normally open solenoid valves (suction check valve 22, electromagnetic actuator 23), and the high pressure pump control means sets the fuel discharge amount from the high pressure pump 20 so that the fuel pressure PF becomes the target fuel pressure PFOBJ. Further, the control is performed through an electromagnetic valve.

  According to this configuration, the normally open solenoid valve for adjusting the amount of fuel discharged from the high pressure pump is provided on the suction side of the high pressure pump. In this type of high-pressure pump, the fuel pressure boosted by the high-pressure pump acts as a back pressure for the solenoid valve on the suction side. Therefore, once the solenoid valve is closed once by energization, the solenoid valve is energized. Without this, the back pressure can keep the solenoid valve closed. For this reason, in changing the amount of fuel discharged from the high pressure pump (hereinafter referred to as “high pressure pump discharge amount”), it is only necessary to control the closing timing of the solenoid valve, and the solenoid valve for closing the solenoid valve is sufficient. The energization time (period) to the valve can be set to a certain relatively short time. Therefore, coupled with the above-described effect of the invention according to claim 1, the power consumption of the entire apparatus can be reduced.

1 is a diagram schematically showing a fuel supply device for an internal combustion engine according to an embodiment of the present invention, together with an internal combustion engine to which the fuel supply device is applied. It is a block diagram which shows ECU etc. of a fuel supply apparatus. It is sectional drawing which shows the high pressure pump at the time of completion | finish of a suction stroke. It is sectional drawing which shows the high pressure pump in a spill process. It is sectional drawing which shows the high pressure pump at the time of completion | finish of a discharge stroke (at the time of the start of a suction stroke). It is a flowchart which shows the pump control process performed by ECU. It is an example of the ELOBJ map used by pump control processing. It is a figure which shows an example of transition of a low pressure pump discharge amount, energization start timing, and a request torque.

  Hereinafter, preferred embodiments of the present invention will be described in detail with reference to the drawings. An internal combustion engine (hereinafter referred to as “engine”) 3 shown in FIG. 1 is a four-cycle gasoline engine for a vehicle (not shown), and includes four cylinders 3a (# 1 to # 4). In addition, the engine 3 is provided with a fuel injection valve (hereinafter referred to as “injector”) 4 and a spark plug (not shown) for each cylinder 3 a, and a fuel supply device 1 that supplies fuel to each injector 4. ing.

  The fuel of the engine 3 is directly injected into the corresponding cylinder 3a from each injector 4, and the air-fuel mixture generated in the cylinder 3a is ignited by a spark plug. That is, the engine 3 is an in-cylinder injection type engine. The opening and closing of the injector 4 is controlled by a control signal from an ECU 2 (see FIG. 2) described later, whereby the fuel injection timing is controlled by the valve opening timing and the fuel injection amount is controlled by the valve opening time. In FIG. 2, only one injector 4 is shown for convenience.

  The fuel supply apparatus 1 includes a fuel tank 11 for storing fuel, a low-pressure pump 12 provided in the fuel tank 11, and a high-pressure pump 20.

  The low-pressure pump 12 is an electric type controlled by the ECU 2, is connected to a 12V battery (not shown) as a power source, and is always operated while the engine 3 is operating. The amount of fuel discharged from the low-pressure pump 12 increases as the power supplied from the 12V battery increases. The 12V battery is charged by a generator that reduces the power of the engine 3. In addition, a fuel suction passage 13, a low pressure delivery pipe 14, and a fuel return passage 15 are connected to the low pressure pump 12. The low-pressure pump 12 sucks the fuel in the fuel tank 11 through the fuel suction passage 13, raises the fuel to a predetermined low-pressure feed pressure (for example, 392 kPa), and then discharges the excess fuel to the low-pressure delivery pipe 14. It returns to the fuel tank 11 through the fuel return path 15. The high-pressure pump 20 is connected to the downstream end of the low-pressure delivery pipe 14, and the low-pressure fuel discharged from the low-pressure pump 12 to the low-pressure delivery pipe 14 is supplied to the high-pressure pump 20.

  The high-pressure pump 20 is a positive displacement pump coupled to a crankshaft (not shown) of the engine 3 and is connected to the high-pressure delivery pipe 16. The high-pressure pump 20 is driven by the crankshaft to further increase the pressure of the low-pressure fuel supplied from the low-pressure pump 12 and discharge it to the high-pressure delivery pipe 16. Details of the high-pressure pump 20 will be described later.

  The high pressure delivery pipe 16 is provided with the above-described four injectors 4 in parallel. The high-pressure fuel discharged from the high-pressure pump 20 to the high-pressure delivery pipe 16 is supplied to each injector 4 and is injected into the corresponding cylinder 3a when the injector 4 is opened. Further, the high-pressure delivery pipe 16 is provided with a fuel pressure sensor 31, and the fuel pressure (hereinafter referred to as “fuel pressure”) PF in the high-pressure delivery pipe 16 is detected by the fuel pressure sensor 31, and the detection signal is input to the ECU 2. Is done.

  In addition, the fuel supply device 1 includes a bypass pipe 17 that bypasses the high-pressure pump 20, and a relief valve 18 is provided in the bypass pipe 17. The relief valve 18 is of a mechanical type and opens when the fuel pressure PF in the high pressure delivery pipe 16 reaches a predetermined relief pressure (for example, 25 MPa), and fuel is supplied from the high pressure delivery pipe 16 to the low pressure delivery pipe 14. By letting it escape, the fuel pressure PF is limited so as not to exceed the relief pressure.

  As shown in FIGS. 3 to 5, the high-pressure pump 20 includes a pump body 21, a suction check valve 22 and a discharge check valve 24 housed in the pump body 21, and an electromagnetic actuator for driving the suction check valve 22. 23, a plunger 25 driven by a drive cam 19, and the like. The drive cam 19 has four cam peaks 19a arranged at equal intervals in the circumferential direction, and is provided integrally with an exhaust camshaft (not shown) of the engine 3, so that the crankshaft rotates twice. Rotate once.

  A booster chamber 21a for boosting fuel is formed inside the pump body 21, and this booster chamber 21a communicates with the low-pressure delivery pipe 14 via a suction port 21b and via a discharge port 21c. It communicates with the high pressure delivery pipe 16. The suction check valve 22 opens and closes the inlet of the boosting chamber 21a, is accommodated in the boosting chamber 21a, and includes a valve body 22a and a coil spring 22b. The valve body 22a is freely movable between a valve opening position for opening the inlet of the pressure increasing chamber 21a (position shown in FIG. 3) and a valve closing position for closing the inlet of the pressure increasing chamber 21a (position shown in FIG. 5). And is biased toward the valve closing position by the coil spring 22b.

  The electromagnetic actuator 23 constitutes a spill valve mechanism together with the suction check valve 22, and includes an actuator body 23a, a coil 23b, an armature 23c, and a coil spring 23d. The coil 23b is accommodated in the actuator body 23a and is electrically connected to the ECU 2. The coil 23b is excited by energization and is held in a non-excited state by stopping energization. Energization of the coil 23b is controlled by the ECU 2.

  In addition, the armature 23c has a predetermined origin position (a position shown in FIGS. 3 and 4) whose tip protrudes toward the suction check valve 22 side, and a predetermined operation position (see FIG. 5) that retreats from the suction check valve 22 side. Between the actuator body 23a and the actuator body 23a. The armature 23c is held at the origin position by the urging force of the coil spring 23d when the coil 23b is in a non-excited state, and is applied to the urging force of the coil spring 23d by the electromagnetic force when the coil 23b is excited. While resisting, it is sucked to the operating position side.

  Further, the urging force of the coil spring 23d of the electromagnetic actuator 23 is set to a value larger than the urging force of the coil spring 22b of the suction check valve 22, so that the coil 23b is not excited in the suction check valve 22. At this time, the valve is kept open by the armature 23c at the origin (see FIG. 4).

  The discharge check valve 24 opens and closes the outlet of the boosting chamber 21a, is accommodated in the valve chamber 21d between the boosting chamber 21a and the discharge port 21c, and includes a valve body 24a and a coil spring 24b. . This valve body 24a is located between a valve opening position (the position shown in FIG. 5) for opening the outlet of the pressure increasing chamber 21a and a valve closing position (the position shown in FIGS. 3 and 4) for closing the outlet of the pressure increasing chamber 21a. And is urged toward the valve closing position by the coil spring 24b.

  In addition, the plunger 25 has a predetermined protruding position [a position shown in FIG. 5 (top dead center position)] at which one end protrudes into the boosting chamber 21a and a predetermined retracted position [see FIG. It is accommodated in the plunger barrel 21e of the pump body 21 so as to be slidable between the position shown in FIG. A spring seat 26 is attached to the other end portion of the plunger 25, and the plunger 25 and the spring seat 26 are in contact with the drive cam 19 via a spring holder 28.

  Further, a coil spring 27 is provided between the spring seat 26 and the pump body 21, and the plunger 25 is urged toward the retracted position by the coil spring 27. With the above configuration, the plunger 25 is held so as to abut against the cam surface of the drive cam 19 via the spring holder 28 by the biasing force of the coil spring 27 during the rotation of the drive cam 19, thereby operating the engine 3. In the middle, it is always driven by the drive cam 19 between the protruding position and the retracted position.

  Next, the operation of the high-pressure pump 20 configured as described above will be described. In the high-pressure pump 20, the suction stroke, the spill stroke, and the discharge stroke are executed once in order in accordance with the rotation of the drive cam 19 during one operation cycle. Hereinafter, the suction stroke, the spill stroke, and the discharge stroke will be described in order.

[Suction stroke]
In the suction stroke, the plunger 25 moves from the protruding position to the retracted position as the drive cam 19 rotates clockwise from the rotational angle position shown in FIG. 5 toward the rotational angle position shown in FIG. Then, the fuel pressure in the pressure increasing chamber 21a is lowered, whereby the suction check valve 22 is opened, and the fuel from the low pressure pump 12 is sucked into the pressure increasing chamber 21a. The rotation angle position of the drive cam 19 shown in FIGS. 5 and 3 indicates the rotation angle position at the start of the suction stroke (at the end of the discharge stroke) and at the end of the suction stroke, respectively. Is executed from the protruding position (top dead center position) to the retracted position (bottom dead center position).

[Spill process]
In the spill stroke following the suction stroke, the plunger 25 moves from the retracted position toward the protruding position as the drive cam 19 rotates from the rotation angle position shown in FIG. 3 toward the rotation angle position shown in FIG. At that time, the electromagnetic actuator 23 is controlled to be in an OFF state by stopping energization of the coil 23b, whereby the suction check valve 22 is held in an open state. Is returned to the low-pressure pump 12 side.

[Discharge process]
In the discharge stroke following the spill stroke, the drive cam 19 rotates from the rotation angle position shown in FIG. 4 toward the rotation angle position shown in FIG. 5, and the electromagnetic actuator 23 is controlled to be turned on by energizing the coil 23b. As a result, the suction check valve 22 is closed. As a result, the fuel pressure in the boosting chamber 21a rises, so that the discharge check valve 24 is opened, and the high-pressure fuel in the boosting chamber 21a is discharged to the high-pressure delivery pipe 16. During this discharge stroke, the coil 23b is energized from an energization start timing TSHP described later to an energization end timing TEHP, whereby the electromagnetic actuator 23 is controlled to be in an on state. Further, the high-pressure fuel in the pressure increasing chamber 21a acts as a back pressure on the valve body 22a of the suction check valve 22, whereby the suction check valve 22 is held in the closed state.

  As described above, in the high-pressure pump 20, the energization start timing TSHP of the electromagnetic actuator 23, that is, the closing timing of the suction check valve 22 is controlled during the spill stroke, so that the high-pressure pump 20 returns to the low-pressure pump 12 side. The amount of fuel is changed. Thereby, the fuel pressure PF in the high-pressure delivery pipe 16 is controlled by adjusting the discharge amount of the fuel discharged from the high-pressure pump 20 to the high-pressure delivery pipe 16 side. In this case, as apparent from the operation of the high-pressure pump 20 described so far, the earlier the closing timing of the suction check valve 22 in the spill stroke, that is, the earlier the energization start timing TSHP in the spill stroke, the faster the boost chamber 21a. Therefore, the amount of fuel returned to the low pressure pump 12 side decreases, and the amount of fuel discharged from the high pressure pump 20 increases.

  The crankshaft of the engine 3 is provided with a crank angle sensor 32 composed of a magnet rotor and an MRE pickup (both not shown) (see FIG. 2). The crank angle sensor 32 outputs a CRK signal and a TDC signal that are pulse signals as the crankshaft rotates.

  The CRK signal is generated and output at every predetermined crank angle. The ECU 2 calculates the engine speed (hereinafter referred to as “engine speed”) NE of the engine 3 based on the CRK signal. The TDC signal is at a predetermined crank angle position (hereinafter referred to as “reference crank angle position”) near the TDC (top dead center) at the start of the intake stroke of the piston (not shown) of the engine 3 in any cylinder 3a. It is a signal representing this. In the present embodiment, since the engine 3 has four cylinders 3a, a TDC signal is generated and output at every crank angle of 180 °. The engine 3 is provided with a cylinder discrimination sensor (not shown). The cylinder discrimination sensor inputs a cylinder discrimination signal, which is a pulse signal for discriminating the cylinder 3a, to the ECU 2. The ECU 2 calculates the rotational angle position of the drive cam 19 described above according to these CRK signal, TDC signal and cylinder discrimination signal.

  Further, the ECU 2 receives from the accelerator opening sensor 33 a detection signal indicating an operation amount (hereinafter referred to as “accelerator opening”) AP of an accelerator pedal (not shown) of the vehicle.

  The ECU 2 is constituted by a microcomputer (not shown) including a CPU, a RAM, a ROM, an input / output interface (all not shown), and the like. In accordance with the detection signals from the various sensors 31 to 33 described above, the ECU 2 determines the amount of fuel discharged from the low pressure pump 12 and the high pressure pump 20 (hereinafter referred to as “low pressure pump discharge amount”, respectively) according to the control program stored in the ROM. In order to control the “high pressure pump discharge amount”), the pump control process shown in FIG. 6 is executed.

  This pump control process is repeatedly executed in synchronism with the generation of the CRK signal described above during the operation of the engine 3. First, in step 1 of FIG. 6 (illustrated as “S1”, the same applies hereinafter), a predetermined map (not shown) is searched according to the calculated engine speed NE and the detected accelerator pedal opening AP. Then, the required torque TREQ is calculated. This required torque TREQ is a torque required for the engine 3.

  Next, the target supply power ELOBJ is calculated by searching a predetermined ELOBJ map shown in FIG. 7 according to the engine speed NE and the calculated required torque TREQ (step 2). This target supply power ELOBJ is a target value of power (hereinafter referred to as “low pressure pump supply power”) supplied to the low pressure pump 12 from the 12V battery described above. As shown in FIG. 7, in this ELOBJ map, three predetermined rotation speed load regions for setting the target supply power ELOBJ are defined in accordance with the engine speed NE and the required torque TREQ. Specifically, the target supply power ELOBJ is set to the first target supply power ELOBJ1 when the engine speed NE and the required torque TREQ are in the low rotation / low load region. Further, the target supply power ELOBJ is set to the second target supply power ELOBJ2 that is larger than the first target supply power ELOBJ1 when the engine speed NE and the required torque TREQ are in the middle rotation / medium load region, When in the high load region, the third target supply power ELOBJ3 is set to be larger than the second target supply power ELOBJ2.

  In the ELOBJ map, the first to third target supply electric powers ELOBJ1 to ELOBJ3 will be described later with the maximum value of the target fuel pressure PFOBJ calculated when the engine speed NE and the required torque TREQ are in the corresponding engine speed load range. A value that can be realized by controlling the discharge amount of the high-pressure pump is set in advance by experiments or the like. This target fuel pressure PFOBJ is a target value of the aforementioned fuel pressure PF (pressure of fuel in the high-pressure delivery pipe 16). That is, the first target supply power ELOBJ1 is set to a value that can realize the maximum value of the target fuel pressure PFOBJ calculated when NE and TREQ are in the low rotation / low load region. In addition, the second target supply power ELOBJ2 is set to a value that can achieve the maximum value of the target fuel pressure PFOBJ calculated when NE and TREQ are in the middle rotation / medium load region, and the third target supply power ELOBJ3 is set to NE and The maximum value of the target fuel pressure PFOBJ calculated when TREQ is in the high rotation / high load region is set to a value that can be realized. Further, when the target supply power ELOBJ is calculated (set) by executing Step 2, the low-pressure pump supply power is controlled to be the calculated target supply power ELOBJ.

  In step 3 following step 2, it is determined whether or not the calculated target supply power ELOBJ is equal to the previous value ELOBJZ. When the answer is YES and the target supply power ELOBJ is equal to the previous value ELOBJZ, the target supply power change flag F_CHAN is set to “0” to indicate that the target supply power ELOBJ has not changed (step 4). Then, the process proceeds to Step 6 described later. On the other hand, when the answer to step 3 is NO and the target supply power ELOBJ is not equal to the previous value ELOBJZ, the target supply power change flag F_CHAN is set to “1” to indicate that the target supply power ELOBJ has changed. (Step 5), go to Step 6.

  In step 6, the above-mentioned target fuel pressure PFOBJ is calculated by searching a predetermined map (not shown) according to the engine speed NE and the required torque TREQ. In this map, the target fuel pressure PFOBJ is set to a larger value as the engine speed NE is higher and the required torque TREQ is larger. This is because the higher the engine speed NE, the shorter the period during which fuel can be injected, and the greater the required torque TREQ (load of the engine 3), the greater the amount of fuel injected by the injector 4. This is because a higher fuel pressure PF is required.

  Next, a fuel pressure deviation DEPF that is a deviation between the calculated target fuel pressure PFOBJ and the detected fuel pressure PF is calculated (step 7). Next, based on the calculated fuel pressure deviation DEPF, an FB correction term COFB is calculated using a predetermined feedback control algorithm (for example, PI control algorithm) (step 8). This FB correction term COFB is used as a correction term in the calculation of the energization start timing TSHP in order to perform feedback control so that the fuel pressure PF becomes the target fuel pressure PFOBJ.

  Next, it is determined whether or not the target supply power change flag F_CHAN is “1” (step 9). When the answer to step 9 is NO (F_CHAN = 0), that is, when the target supply power ELOBJ has not changed, the energization start timing TSHP of the high-pressure pump 20 is calculated by the normal-time TSHP calculation processing (step 10) The process is terminated.

  In step 10, the energization start timing TSHP is calculated as follows. That is, the energization start timing TSHP is calculated by adding the FB correction term COFB calculated in step 8 to the energization start timing TSHP (previous value) obtained at that time. When the engine 3 is started and the energization start timing TSHP is not calculated, a predetermined initial value is set as the previous value. The energization start timing TSHP is represented by the rotational angle position of the drive cam 19 based on the end of the discharge stroke described above, and is calculated at the timing during the spill stroke. In this case, the energization start timing TSHP is calculated as a larger rotational angle position as it is earlier (as it is on the more advanced side).

  On the other hand, when the answer to step 9 is YES (F_CHAN = 1) and the target supply power ELOBJ changes, an offset correction term COFF is calculated (step 11). This offset correction term COFF is used as a correction term in the calculation of the energization start timing TSHP in order to compensate for the influence on the high pressure pump discharge amount due to the change in the low pressure pump discharge amount accompanying the change in the target supply power ELOBJ. .

  Specifically, the offset correction term COFF is calculated (set) as follows. That is, when the target supply power ELOBJ increases from the first target supply power ELOBJ1 to the second target supply power ELOBJ2, and when the target supply power ELOBJ2 decreases from the second target supply power ELOBJ2 to the first target supply power ELOBJ1, the offset correction term COFF is The predetermined first offset correction term COFF1 (see FIG. 8) is set. When the target supply power ELOBJ increases from the second target supply power ELOBJ2 to the third target supply power ELOBJ3 and when the target supply power ELOBJ decreases from the third target supply power ELOBJ3 to the second target supply power ELOBJ2, the offset correction term COFF is The predetermined second offset correction term COFF2 (see FIG. 8) is set.

  In step 12 following step 11 above, the energization start timing TSHP is calculated by the TSHP calculation process for changing the ELOBJ, and this process ends. The calculation method of the energization start timing TSHP in Step 12 is different between when the target supply power ELOBJ is increased and when it is decreased. Specifically, when the target supply power ELOBJ increases, the sum of the energization start timing TSHP (previous value) obtained at that time and the FB correction term COFB calculated in step 8 is calculated in step 11. The energization start timing TSHP is calculated by subtracting the calculated offset correction term COFF (TSHP ← TSHP previous value + COFB−COFF). On the other hand, when the target supply power ELOBJ decreases, the energization start timing TSHP is calculated by adding the offset correction term COFF to the sum of the previous value of the energization start timing TSHP and the FB correction term COFB (TSHP ← TSHP last time). Value + COFB + COFF).

  Further, the energization end timing TEHP of the high-pressure pump 20 is retarded by a predetermined value from the energization start timing TSHP by subtracting a predetermined value from the energization start timing TSHP, regardless of whether or not the target supply power ELOBJ has changed. Is calculated at the rotation angle position. This predetermined value is set to a very small value. Thus, the energization end timing TEHP is calculated at a timing during the discharge stroke.

  FIG. 8 shows an example of changes in the low-pressure pump discharge amount QLPO, the energization start timing TSHP, and the required torque TREQ. As described above, the target supply power ELOBJ is set to the first target supply power ELOBJ1 when the required torque TREQ is in the low load region, and the second target supply power ELOBJ2 when the required torque TREQ is in the medium load region and the high load region. The third target supply power ELOBJ3 is set (step 2, FIG. 7). Thereby, as shown in FIG. 8, the low pressure pump discharge amount QLPO is controlled to a constant first discharge amount QLPO1 according to ELOBJ1 when TREQ is in the low load region, and is set to ELOBJ2 when it is in the medium load region. It is controlled to a constant second discharge amount QLPO2 corresponding to the constant, and when in the high load region, it is controlled to a constant third discharge amount QLPO3 corresponding to ELOBJ3. The magnitude relationship between the first to third discharge amounts QLPO1 to QLPO3 is QLPO3> QLPO2> QLPO1.

  By setting the target supply power ELOBJ in the execution of step 2, the low-pressure pump discharge amount QLPO is set so that the engine speed NE belongs to the high speed side in the predetermined three engine speed load areas (see FIG. 7). Control is performed such that the greater the region is, the more the load region to which the required torque TREQ belongs is the region on the high load side. Further, the low-pressure pump discharge amount QLPO is set to a constant value that can be realized by the control of the high-pressure pump 20 described above, with the maximum value of the target fuel pressure PFOBJ set in each rotation speed load area in each of the three rotation speed load areas. Be controlled.

  The energization start timing TSHP is determined when the target supply power ELOBJ has not changed (step 9: NO), that is, when the low-pressure pump discharge amount QLPO has not changed (other than the time points t1 and t2). Regardless of the rotational speed load region to which the number NE and the required torque TREQ belong, it is calculated by adding the FB correction term COFB to the previous value (step 10). As a result, the energization start timing TSHP is calculated to be a more advanced value as the required torque TREQ is larger, thereby increasing the discharge amount of the high-pressure pump. As a result, the fuel pressure PF is changed to a larger target fuel pressure PFOBJ. To be increased.

  Further, when the target supply power ELOBJ increases (step 9: YES), that is, when the low-pressure pump discharge amount QLPO increases (time t1, time t2), the energization start timing TSHP is corrected with the previous value and FB correction. It is calculated by subtracting the offset correction term COFF from the sum of the terms COFB (step 12). In this case, the offset correction term COFF is set to the first offset correction term COFF1 when the target supply power ELOBJ increases from the first target supply power ELOBJ1 to the second target supply power ELOBJ2 (corresponding to the time point t1). When the second target supply power ELOBJ2 increases to the third target supply power ELOBJ3 (corresponding to the time point t2), the second offset correction term COFF2 is set (step 11).

  As described above, the energization start timing TSHP is corrected to a value on the retard side (side to decrease the high-pressure pump discharge amount) by the amount of the first offset correction term COFF1 when the target supply power ELOBJ increases from ELOBJ1 to ELOBJ2. When it increases from ELOBJ2 to ELOBJ3, it is corrected to the value on the retard side by the amount of the second offset correction term COFF2. As a result, the increase in the low-pressure pump discharge amount QLPO accompanying the increase in the target supply power ELOBJ is compensated, so that the fuel pressure PFOBJ can be quickly converged to the target fuel pressure PFOBJ.

  When the target supply power ELOBJ decreases (step 9: YES), that is, when the low-pressure pump discharge amount QLPO decreases (time t1, time t2), the energization start timing TSHP is corrected with the previous value and FB correction. The offset correction term COFF is added to the sum of the terms COFB (step 12). In this case, the offset correction term COFF is set to the first offset correction term COFF1 when the target supply power ELOBJ decreases from the second target supply power ELOBJ2 to the first target supply power ELOBJ1 (corresponding to the time point t1). When the third target supply power ELOBJ3 decreases to the second target supply power ELOBJ2 (corresponding to the time point t2), the second offset correction term COFF2 is set (step 11).

  As described above, the energization start timing TSHP is corrected to the value on the advance side (the side that increases the discharge amount of the high-pressure pump) by the amount of the first offset correction term COFF1 when the target supply power ELOBJ decreases from ELOBJ2 to ELOBJ1. When it decreases from ELOBJ3 to ELOBJ2, it is corrected to the value on the advance side by the amount of the second offset correction term COFF2. As a result, the decrease in the low-pressure pump discharge amount QLPO accompanying the decrease in the target supply power ELOBJ is compensated, so that the fuel pressure PFOBJ can be quickly converged to the target fuel pressure PFOBJ.

  FIG. 8 shows the relationship between the low-pressure pump discharge amount QLPO, the energization start timing TSHP and the required torque TREQ (load of the engine 3), but the low-pressure pump discharge amount QLPO, the energization start timing TSHP and the engine speed NE The relationship is the same.

  The correspondence between various elements in the present embodiment and various elements in the present invention is as follows. That is, the ECU 2 in the present embodiment corresponds to the operation state detection means, the low pressure pump control means, the target fuel pressure setting means, and the high pressure pump control means in the present invention, and the suction check valve 22 and the electromagnetic actuator 23 in the present embodiment are It corresponds to a solenoid valve in the present invention. Further, the crank angle sensor 32 and the accelerator opening sensor 33 in the present embodiment correspond to the driving state detecting means in the present invention.

  As described above, according to this embodiment, the discharge amount of the high-pressure pump is controlled via the energization start timing TSHP of the electromagnetic actuator 23 so that the fuel pressure PF becomes the target fuel pressure PFOBJ (steps 8, 10, 12). ). Further, the low-pressure pump discharge amount QLPO is controlled in accordance with the engine speed NE and the required torque TREQ (step 2, FIG. 7).

  Specifically, when the low-pressure pump discharge amount QLPO is in a predetermined three rotational speed load region, the higher the rotational speed region to which the engine rotational speed NE belongs, the more the load region to which the required torque TREQ belongs. Is controlled so as to increase as the region is on the high load side. Further, in each of the three rotational speed load regions, the low-pressure pump discharge amount QLPO is suppressed to a value that can be realized by the control of the high-pressure pump 20 described above by controlling the maximum value of the target fuel pressure PFOBJ set in each rotational speed load region. Is done. As described above, the low-pressure pump discharge amount QLPO can be appropriately controlled according to the engine speed NE and the required torque TREQ when the operating state of the engine 3 is in any of the three engine speed load regions, and as a result, the fuel pressure PF The control accuracy can be improved.

  Further, the low pressure pump discharge amount QLPO is not controlled to a value that can realize the maximum value of the target fuel pressure PFOBJ set in the whole of the three rotation speed load areas in each of the three rotation speed load areas, The maximum value of the target fuel pressure PFOBJ set in each rotation speed load region is suppressed to a value that can be realized. Therefore, the power consumption of the low-pressure pump 12 can be appropriately reduced according to the engine speed NE and the required torque TREQ, and consequently the fuel consumption of the engine 3 can be improved.

  Furthermore, a normally open solenoid valve including a suction check valve 22 and an electromagnetic actuator 23 for adjusting the discharge amount of the high pressure pump is provided on the suction side of the high pressure pump 20. Therefore, coupled with the above-described effect of reducing the power consumption of the low-pressure pump 12, the power consumption of the entire fuel supply device 1 can be reduced.

  In addition, this invention can be implemented in various aspects, without being limited to the described embodiment. For example, in the embodiment, both the engine speed NE and the required torque TREQ are used as the operating state of the internal combustion engine in the present invention, but one of both NE and TREQ may be used. Alternatively, other appropriate parameters representing the operating state of the internal combustion engine, for example, an intake air amount or a fuel injection amount may be used. In the embodiment, as shown in FIG. 8, the low-pressure pump discharge amount QLPO is controlled stepwise, but may be controlled so as to change linearly according to the operating state of the internal combustion engine, or You may control so that it may change in a curve. In this case, for example, in each of the plurality of rotation speed load regions, a minimum value of the low pressure pump discharge amount and an intermediate value between the minimum value and the maximum value of the low pressure pump discharge amount are set in each rotation speed load region. The target fuel pressure is controlled (set) so that the maximum value can be realized. Furthermore, in the embodiment, the number of the plurality of rotation speed regions and load regions in the present invention is three, but is arbitrary.

  In the embodiment, the control of the low-pressure pump discharge amount according to claims 2 and 3 is executed in each of the three rotation speed load regions, but is executed only in a predetermined one operation region, and the others In this operating region, the discharge amount of the low-pressure pump may be controlled to be larger as the engine speed NE is higher and the required torque TREQ is larger, without being controlled to a constant value. Furthermore, in the embodiment, the low-pressure pump discharge amount QLPO is controlled in a feedforward manner based on the target supply power ELOBJ, but may be feedback-controlled. In this case, for example, the actual value of the low-pressure pump discharge amount is detected by a sensor or the like, the target value of the low-pressure pump discharge amount is set, and predetermined feedback control is performed based on the deviation between the target value of the low-pressure pump discharge amount and the actual value. The target supply power of the low-pressure pump is calculated using the feedback correction term calculated by the algorithm.

  In the embodiment, the high-pressure pump 20 of the type in which the normally-open suction check valve 22 is provided on the suction side is used. However, other appropriate high-pressure pumps may be used. A high-pressure pump of the type in which the suction check valve is provided on the suction side may be used. Furthermore, in the embodiment, the engine 3 that is a gasoline engine is used as the internal combustion engine in the present invention, but a diesel engine, an LPG engine, or the like may be used. Of course, variations of the above embodiments may be combined as appropriate. In addition, it is possible to appropriately change the detailed configuration within the scope of the gist of the present invention.

DESCRIPTION OF SYMBOLS 1 Fuel supply apparatus 2 ECU (Operating state detection means, low pressure pump control means, target fuel pressure setting means, high pressure pump control means)
3 Engine 4 Injector 11 Fuel tank 12 Low pressure pump 20 High pressure pump 22 Suction check valve (solenoid valve)
23 Electromagnetic actuator (solenoid valve)
32 Crank angle sensor (operating state detection means)
33 Accelerator opening sensor (operating state detection means)
PF Fuel pressure NE Engine speed (operating state of internal combustion engine, speed of internal combustion engine)
TREQ required torque (operating state of internal combustion engine, load of internal combustion engine)
PFOBJ Target fuel pressure QLPO Low pressure pump discharge

Claims (4)

A fuel supply device for an internal combustion engine for supplying fuel to a fuel injection valve of the internal combustion engine,
An electric low pressure pump that discharges fuel in a fuel tank to the fuel injection valve side;
An operating state detecting means for detecting an operating state of the internal combustion engine;
Low-pressure pump control means for controlling a low-pressure pump discharge amount that is a fuel discharge amount from the low-pressure pump according to the detected operating state of the internal combustion engine;
A high-pressure pump that sucks the fuel discharged from the low-pressure pump and discharges the fuel to the fuel injection valve side in a pressurized state;
Target fuel pressure setting means for setting a target fuel pressure, which is a target value of the fuel pressure on the fuel injection valve side of the high-pressure pump, according to the operating state of the internal combustion engine;
High-pressure pump control means for controlling the high-pressure pump so that the fuel pressure that is the fuel pressure on the fuel injection valve side of the high-pressure pump becomes the set target fuel pressure,
The low-pressure pump control means suppresses the discharge amount of the low-pressure pump so that a maximum value of the set target fuel pressure can be realized by controlling the high-pressure pump by the high-pressure pump control means. Engine fuel supply.   The low-pressure pump control means determines the discharge amount of the low-pressure pump as the detected rotational speed of the internal combustion engine in a plurality of predetermined rotational speed regions defined by the rotational speed of the internal combustion engine as the operating state of the internal combustion engine. And the maximum value of the target fuel pressure set in each rotation speed region in each of the plurality of rotation speed regions is controlled so that the rotation speed region to which the engine belongs is higher. The fuel supply device for an internal combustion engine according to claim 1, wherein the discharge amount of the low-pressure pump is suppressed to a value that can be realized by controlling the high-pressure pump by the high-pressure pump control means.   The low-pressure pump control means is configured to determine the discharge amount of the low-pressure pump as a load to which the detected load of the internal combustion engine belongs in a plurality of predetermined load regions defined by the load of the internal combustion engine as the operating state of the internal combustion engine. The higher the load region is, the more the region is controlled so as to increase, and in each of the plurality of load regions, the maximum value of the target fuel pressure set in each load region is set to the high pressure pump control means. 2. The fuel supply device for an internal combustion engine according to claim 1, wherein the discharge amount of the low-pressure pump is suppressed to a value that can be realized by the control of the high-pressure pump. On the suction side of the high-pressure pump, a normally open solenoid valve for adjusting the amount of fuel discharged from the high-pressure pump is provided,
The high-pressure pump control means controls the amount of fuel discharged from the high-pressure pump via the solenoid valve so that the fuel pressure becomes the target fuel pressure. A fuel supply device for an internal combustion engine according to claim 1.

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