JP5534320B2 - Valve timing control device - Google Patents

Description

  The present invention relates to a drive-side rotating body that rotates synchronously with a crankshaft of an internal combustion engine, and a driven-side rotation that is arranged coaxially with the drive-side rotating body and rotates synchronously with a camshaft for opening / closing a valve of the internal combustion engine. The fluid pressure is formed by the body, the drive-side rotator, and the driven-side rotator, and is partitioned into the retard chamber and the advance chamber by a partition provided on at least one of the drive-side rotator and the driven-side rotator. A urging mechanism for urging the driven-side rotator in the advance phase direction with respect to the drive-side rotator, supply of the working fluid from the pump that supplies the working fluid to the fluid pressure chamber, and from the fluid pressure chamber And a fluid control mechanism that controls the discharge of the working fluid.

  As the valve opening / closing timing control device as described above, there is a valve opening / closing timing control device as disclosed in Patent Document 1. Some fluid control mechanisms used in such valve opening / closing timing control devices have a region called a dead zone with low responsiveness. There is a known problem that the control becomes unstable in such a dead zone.

  In order to solve this problem, there is a technique for determining a use signal region based on the center value of the dead zone and setting a control signal for the fluid control mechanism in the use signal region (see Patent Document 2). In the technique of this patent document 2, by setting the area for setting the control signal for controlling the fluid control mechanism to be a controllable range, the use signal is avoided from being located in the dead zone, and the controllability is improved.

  Moreover, when controlling a fluid control mechanism, an apparatus is known that sets different offset amounts according to the control direction (advance angle direction or retard angle direction) (see Patent Document 3). In the apparatus disclosed in Patent Document 3, offsets having different signs are added when displaced in the advance direction and when displaced in the retard direction. Thereby, the biasing force by the spring (biasing mechanism) is corrected, and appropriate control is realized.

JP 2009-74384 A JP 2008-115729 A JP 2009-138650 A

  However, in the valve opening / closing timing control device of Patent Document 1, the biasing force by the biasing mechanism is not taken into account when controlling the fluid control mechanism, so the relative angle of the driven-side rotating body with respect to the driving-side rotating body is set to a desired angle. It is difficult to make. Moreover, since the apparatus of Patent Document 2 does not have an urging mechanism, it is difficult to apply the apparatus to an apparatus having an urging mechanism. On the other hand, in the apparatus of Patent Document 3, the urging mechanism has been studied, but the sign of the offset is simply changed according to the control direction (advance angle direction or retard angle direction). Insufficient examination of the urging power of

  In view of such a problem, an object of the present invention is to provide a valve opening / closing timing control device that enables control in consideration of an urging force of an urging mechanism.

In order to solve the above problem, a valve opening / closing timing control device according to the present invention includes a driving side rotating body that rotates synchronously with a crankshaft of an internal combustion engine, and a coaxial arrangement with respect to the driving side rotating body, A driven-side rotating body that rotates synchronously with a camshaft for opening and closing a valve of an internal combustion engine, the driving-side rotating body, and the driven-side rotating body, and formed on at least one of the driving-side rotating body and the driven-side rotating body A fluid pressure chamber partitioned into a retard chamber and an advance chamber by a provided partition, and an urging mechanism for urging the driven-side rotator in an advance phase direction with respect to the drive-side rotator; A fluid control mechanism for controlling the supply of the working fluid from the pump that supplies the working fluid to the fluid pressure chamber and the discharge of the working fluid from the fluid pressure chamber, and the driving of the driven rotor Relative phase with respect to the side rotor An angle information detection unit that detects a counter angle, a fluid temperature measurement unit that measures the temperature of the working fluid, a discharge pressure measurement unit that measures the discharge pressure of the pump, and the relative angle to be a predetermined target angle A control unit that controls the fluid control mechanism based on the relative angle, the oil temperature of the working fluid, and the discharge pressure of the pump, and the fluid control mechanism is configured to supply the fluid pressure chamber to the fluid pressure chamber. A closing control position where supply of the working fluid or discharge of the working fluid from the fluid pressure chamber is stopped, and supply of the working fluid to the fluid pressure chamber or discharge of the working fluid from the fluid pressure chamber are performed. Between the retard angle control position and the advance angle control position, and when the duty output to the fluid control mechanism is 0%, it is located at one of the retard angle control position and the advance angle control position, and the duty Increased output Displacement to the other of the retard angle control position and the advance angle control position via the blockage control position and, when displaced to the blockage control position, a torque corresponding to the duty output at the blockage control position. value has been corrected to a value of torque corresponding to the duty output in the retarded angle control position or the advanced angle control position, the duty output on the basis of the value of the torque after correction Ru is determined.

In this configuration, the driven-side rotator is urged in the advance phase direction with respect to the drive-side rotator by the urging mechanism. This urging force varies depending on the amount of deformation of the spring or the like constituting the urging mechanism, that is, the relative phase (hereinafter referred to as a relative angle) of the driven side rotator with respect to the drive side rotator. In general, when the oil temperature of the working fluid increases, the viscosity of the working fluid decreases and the response time of the solenoid of the fluid control mechanism increases. On the other hand, the greater the discharge force of the pump, the faster the response time of the fluid control mechanism. For this reason, the above configuration includes a control unit that controls the fluid control mechanism based on the relative angle, the oil temperature of the working fluid, and the discharge pressure of the pump so that the relative angle becomes a predetermined target angle. Here, the target angle is a relative angle at which the opening / closing state of the valve of the internal combustion engine is appropriate, and the opening / closing state of the valve of the internal combustion engine can be made appropriate by controlling the relative angle as described above. This can improve the fuel consumption of the internal combustion engine.
In addition, the correlation between the duty output for the actual fluid control mechanism and the torque of the driven rotor is hysteresis between when it is operated from the advance side to the retard side and when it is moved from the advance side to the retard side. Occurs. This is because of the presence or absence of urging of the urging mechanism in the advance direction and the influence of counter torque on the camshaft. Therefore, in the above-described configuration, the actual correlation is corrected instead of the actual correlation. Thereby, the control in the dead zone where the operation of the fluid control mechanism becomes unstable can be stabilized.

  As a control method of the fluid control mechanism, there are map control and PID control. When the control is performed only by the map control, it cannot be sufficiently applied to the individual difference, so that the control varies from product to product, which is not desirable. On the other hand, when the control of the fluid control mechanism is performed only by PID control, it is possible to sufficiently cope with individual differences. However, if the characteristics of the object to be controlled fluctuate due to variations in the oil temperature or hydraulic pressure, the tuning effort increases, which is not desirable. Therefore, in a preferred embodiment of the valve timing control device of the present invention, the control unit is based on the relative angle output by map control using a map in which the relative angle and the correction amount are associated with each other. And a duty based on the oil temperature, the number of rotations of the pump, and the control based on the oil temperature and the discharge pressure output by the PID control that receives the difference between the target angle and the actual angle. Take control.

  In this configuration, map control is performed for parameters suitable for map control, and PID control is performed for other parameters. Accordingly, appropriate control can be performed while balancing the application to individual differences and the tuning effort. A parameter suitable for map control is a simple relationship in which the relationship between input and output is linear or the like. In the present invention, the relationship between the relative angle and the correction amount for correcting the biasing force of the biasing mechanism is 1. Since there is a relationship that can be approximated by the following equation or the quadratic equation, map control is used for the control using the relative angle as an input. On the other hand, the control based on the oil temperature and the discharge force of the pump uses PID control. Since the discharge force of the normal pump is proportional to the rotation speed of the pump, the above configuration uses the rotation speed of the pump instead of the discharge force of the pump as an input for PID control. In this configuration, when a mechanical hydraulic pump driven by the driving force of the engine is used as the pump, the pump rotation speed becomes the engine rotation speed, so that a sensor for measuring the pump rotation speed is not necessary. . Thus, in this configuration, stable control can be performed by using three input parameters, and the control parameter tuning effort is reduced by combining map control and PID control according to the input parameters. can do.

  In one preferred embodiment of the valve timing control apparatus of the present invention, the target angle is determined from a map based on the engine torque and the engine speed estimated based on the engine speed and the throttle opening. The

  The throttle state clearly represents the operation of the accelerator pedal that is directly operated by the driver. Therefore, the engine torque desired by the driver at that time can be calculated more accurately by using the engine speed and the throttle opening. Therefore, with the above configuration, when the driver desires, the maneuverability with the highest priority on the driver's feeling of operation can be achieved.

  In one preferred embodiment of the valve timing control apparatus of the present invention, the control constant is changed in the PID control.

  With this configuration, it is possible to improve the response of changing the relative angle by a simple process of changing the control constant.

In one preferred embodiment of the valve timing control apparatus of the present invention, the control unit, upon operating the pre Symbol fluid control mechanism used Quito torque before being corrected according to the relative angle.

As described above, in the valve opening / closing timing control device of the present invention, the driven side rotating body is urged in the advance direction by the urging mechanism. Therefore, as the relative angle of the driven-side rotating member is retarded, the torque of the driven-side rotating member for urging force becomes stronger the biasing mechanism is easily obtained. On the other hand, becomes torque is difficult to obtain a driven rotor for urging force of the biasing mechanism as the relative angle of the driven-side rotating member is in the advance side is weakened. Therefore, it is necessary to operate the fluid control mechanism more quickly and increase the operation of the driven side rotating body as the driven side rotating body is on the advance side. Therefore, in the above arrangement, the torque is corrected in accordance with the relative angle control is performed of the fluid control mechanism based on the corrected correlation using the corrected torque. As a result, the valve opening / closing timing can be controlled while reducing the influence of fluctuations in the urging force of the urging mechanism.

1 is a cutaway sectional view showing an overall configuration of a valve timing control apparatus according to the present invention. It is the II-II sectional view of Drawing 1 in one operation state of a valve timing control device. It is a schematic diagram of the valve opening / closing timing control device in the most retarded phase. It is a schematic diagram of the valve opening / closing timing control device in the most advanced angle phase. It is a functional block diagram showing each function part of a control part. It is a flowchart showing the flow of a duty output determination process. It is the figure which represented typically how to obtain | require a duty output. It is a figure showing the correlation of the measured duty output and the torque of a driven side rotary body.

  Hereinafter, an embodiment of a valve timing control apparatus of the present invention will be described with reference to the drawings. FIG. 1 is a cross-sectional view schematically showing the configuration of the valve timing control apparatus of the present invention. 2 is a cross-sectional view taken along the line II-II in FIG. 1, and is a plan view schematically showing the state of the phase conversion mechanism 1 in one operating state. The phase conversion mechanism 1 includes a drive-side rotator 12 that rotates synchronously with a crankshaft of an engine (internal combustion engine), and a camshaft that is coaxially disposed with respect to the drive-side rotator 12 and that opens and closes the valve of the internal combustion engine. 10 and a driven side rotating body 11 that rotates synchronously. In the present embodiment, the driven side rotating body 11 is disposed inside the driving side rotating body 12. Moreover, although the drive side rotary body 12 uses the sprocket in this embodiment, it is good also as other structures, such as a pulley. Rotational force from the crankshaft of the engine is transmitted to the drive-side rotator 12 via a transmission mechanism such as a belt or chain (not shown). The driven-side rotator 11 is fixed to the camshaft 10 with bolts 14. When the driven-side rotator 11 rotates in conjunction with the drive-side rotator 12, the camshaft 10 is rotated, and the intake valve or exhaust of the engine The valve opens and closes.

  As shown in FIG. 2, a space 2 (an example of a fluid pressure chamber of the present invention) is formed between the driving side rotating body 12 and the driven side rotating body 11. This space 2 is divided into two types of pressure chambers 2A and 2B by a vane 13 (an example of a partition portion in the present invention) which is a movable partition. Since the volume of the space 2 is unchanged, the volume of the pressure chambers 2A and 2B is changed complementarily by changing the position of the vane 13 in the space 2. When the volumes of the pressure chambers 2A and 2B change, the relative phase (hereinafter referred to as a relative angle) of the driven-side rotator 11 with respect to the drive-side rotator 12 is changed. By changing the relative angle, the opening / closing timing of the intake valve and the exhaust valve with respect to the piston-moving engine is changed. The partition between the pressure chambers 2A and 2B is not limited to the block-shaped vane 13 as shown in FIG. Further, in the present embodiment, the space 2 is formed by the concave portion of the driving side rotating body 12 and the vane 13 is provided in the driven side rotating body 11, but the space 2 is formed by the concave portion of the driven side rotating body 11, The vane 13 may be provided on the drive side rotating body 12.

  In the present embodiment, the entire phase conversion mechanism 1 rotates clockwise. FIG. 2 shows an intermediate lock phase set as a relative angle suitable for starting the internal combustion engine. The intermediate lock phase includes the most retarded phase state in which the relative angle of the driven-side rotator 11 is most delayed with respect to the drive-side rotator 12 and the relative angle of the driven-side rotator 11 with respect to the drive-side rotator 12. It is set in an intermediate region between the most advanced advance angle phase state and is fixedly held by the lock mechanism 6.

  When the lock mechanism 6 is released from the state of FIG. 2 and engine oil (an example of the working fluid of the present invention) is supplied to the pressure chamber 2A and the engine oil is discharged from the pressure chamber 2B, the pressure chamber for the pressure chamber 2B is discharged. The relative volume of 2A increases. As a result, the relative angle of the driven-side rotator 11 changes to the retard side. Conversely, when engine oil is supplied to the pressure chamber 2B and the engine oil is discharged from the pressure chamber 2A, the phase of the driven-side rotator 11 changes to the advance side. For this reason, in the following description, the pressure chamber 2A is referred to as a retard chamber, and the pressure chamber 2B is referred to as an advance chamber. In addition, the flow path leading to the retard chamber 2A and the flow path leading to the advance chamber 2B shown in FIG. 1 are referred to as a retard channel 21 and an advance channel 22, respectively. The retard chamber 2A and the advance chamber 2B are not completely sealed. Therefore, when engine oil exceeding the respective capacities is supplied, the engine oil leaks outside the phase conversion mechanism 1. The leaked engine oil is collected together with the engine oil supplied to each part of the engine.

  As shown in FIG. 2, the lock mechanism 6 is formed on a part of the outermost peripheral surface of the retard-side lock body 6 </ b> A and the advance-angle lock section 6 </ b> B provided on the drive-side rotary body 12 and the driven-side rotary body 11. The locking recess 62 is provided. The retard lock 6A and the advance lock 6B function to regulate the relative angle displacement toward the retard angle side and the relative angle displacement toward the advance angle side, respectively. The retard lock portion 6A and the advance lock portion 6B are respectively provided with lock pieces 60A and 60B and lock pieces 60A and 60B supported on the drive side rotator 12 so as to be slidable in the radial direction. It has a spring 61 that protrudes and biases in the direction. The lock recess 62 is provided along the circumferential direction of the driven-side rotator 11, and a lock groove 62M for performing the original lock function by engaging the lock pieces 60A and 60B, and a lock groove It is a two-stage groove provided with a regulation auxiliary locking groove 62a in which the locking depth by the lock piece 60A is shallower than 62M. As shown in FIG. 2, the regulation auxiliary locking groove 62 a extends from the end portion on the most advanced angle side of the locking groove 62 </ b> M toward the advanced angle side. As the shapes of the lock pieces 60A and 60B, a plate shape, a pin shape, or the like can be appropriately employed.

  The retard lock portion 6A prevents the driven-side rotating body 11 from being displaced from the intermediate lock phase to the retard side by engaging the retard lock piece 60A into the engaging groove 62M. On the other hand, the advance lock portion 6B engages the advance lock piece 60B in the lock recess 62 to prevent the driven-side rotator 11 from being displaced from the intermediate lock phase to the advance side. .

  The width of the locking groove 62M deeper than the restriction auxiliary locking groove 62a is the distance between the side surfaces of the retard lock piece 60A and the advance lock piece 60B that are spaced apart from each other in the circumferential direction of the driven-side rotating body 11. Approximately matched. Therefore, as shown in FIG. 2, the phase angle of the driven-side rotating body 11 is substantially reduced by engaging both the retarding lock piece 60A and the advancement lock piece 60B into the locking groove 62M at the same time. A so-called locked state in which an intermediate lock phase having no allowable width is constrained can be obtained.

  The lock recess 62 communicates with a lock channel 63 formed in the driven-side rotator 11, and the lock channel 63 is connected to a second control valve 75 of the hydraulic circuit 7 described later. When engine oil is supplied from the second control valve 75 to the lock recess 62 via the lock channel 63, the pair of lock pieces 60A and 60B engaged with the lock recess 62 have their tips rotated on the driven side. Retreat to the drive side rotating body 12 side until it is positioned slightly radially outside the outermost peripheral surface of the body 11. As a result, the locked state between the driving side rotating body 12 and the driven side rotating body 11 is released, and the relative phase can be displaced.

  As shown in FIG. 1, the hydraulic circuit 7 is driven by an engine to supply a first pump 71 that supplies engine oil, a second pump 72 provided on the downstream side of the first pump 71, It has between the 2nd pump 72, and the storage part 73 which can store an engine oil is provided.

  The first pump 71 is a mechanical hydraulic pump driven by the driving force of the crankshaft of the engine. The first pump 71 draws in engine oil stored in the oil pan 76 from the suction port and discharges it downstream from the discharge port. The discharge port of the first pump 71 communicates with the engine lubrication system 78 and the reservoir 73 via the filter 77. The engine lubrication system 78 is a concept that includes all parts that require the supply of the engine and surrounding engine oil.

  The second pump 72 is an electric pump driven by power different from the engine, for example, an electric motor. Therefore, the second pump 72 can operate according to the operation signal from the control unit 8 regardless of the operating state of the engine. The second pump 72 sucks the engine oil stored in the storage portion 73 from the suction port and discharges it downstream from the discharge port. The discharge port of the second pump 72 communicates with the first control valve 74 and the second control valve 75. Further, the hydraulic circuit 7 includes a bypass flow path 79 that communicates the upstream flow path and the downstream flow path of the second pump 72 so as to be parallel to the second pump 72. The bypass flow path 79 is provided with a check valve 79a.

  The storage part 73 is provided between the first pump 71 and the second pump 72 and has a storage chamber 73a capable of storing a certain amount of engine oil. The storage unit 73 is provided at a position lower than the first communication port 73 b and the first communication port 73 b that allow the storage chamber 73 a to communicate with the flow path on the downstream side of the first pump 71, and the storage chamber 73 a is connected to the second pump 72. A second communication port 73c that communicates with the upstream flow path, and a lubrication system communication port 73d that is provided at a position higher than the first communication port 73b and communicates the storage chamber 73a with the engine lubrication system 78.

  When the engine is stopped, that is, when the first pump 71 is stopped, the second pump 72 performs an operation of supplying engine oil to the space 2 and the lock mechanism 6. Therefore, the capacity of the region lower than the first communication port 73 b and higher than the second communication port 73 c is the capacity of the space 2 and the lock recess 62 of the lock mechanism 6, and the capacity of the flow path between these and the second pump 72. The capacity of the storage chamber 73a of the storage unit 73 is set so as to be equal to or greater than the combined capacity. Thereby, even if the 1st pump 71 is a stop state, it becomes possible to displace the relative angle of the driven side rotary body 11 to a target relative angle with the 2nd pump 72. FIG.

  The hydraulic circuit 7 includes a first control valve 74 (an example of a fluid control mechanism of the present invention) that controls the supply of engine oil to the space 2 and a second control that controls the supply of engine oil to the lock mechanism 6. And a valve 75. Further, the hydraulic circuit 7 has a control unit 8 that controls the operation of the second pump 72, the first control valve 74, and the second control valve 75. In the present embodiment, the control unit 8 is configured with an EUC (Electronic Control Unit) as a core.

  As the first control valve 74, for example, a variable electromagnetic spool valve that displaces a spool that is slidable in the sleeve by energizing the solenoid from the control unit 8 against the spring can be used. The first control valve 74 includes an advance port that communicates with the advance channel 22, a retard port that communicates with the retard channel 21, and a supply port that communicates with a downstream channel of the second pump 72. And a drain port communicating with the oil pan 76. The first control valve 74 communicates the advance port and the supply port, communicates the retard port and the drain port, communicates the retard port and the supply port, and advances the advance port and the drain port. Is a three-position control valve that can perform three state controls: a retard control position that communicates with each other, and a hold control position that closes the advance port and the retard port. The first control valve 74 controls the engine oil in the space 2 and controls the relative angle of the driven-side rotator 11 by sliding the solenoid according to a control signal (duty output) from the control unit 8. . Specifically, when the duty output is 0%, the first control valve 74 is in the retard control position, and as the duty output increases, the first control valve 74 is displaced to the advance control position via the closing control position. That is, the first control valve 74 is closed while the driven-side rotator 11 is displaced from the most retarded phase state to the most advanced angle phase state. In this state, since the supply of engine oil to the space 2 is interrupted, even if the duty output is increased, the phase of the driven rotor 11 hardly changes. This section is called a dead zone (see FIG. 8). Due to the presence of this dead zone, the engine operating state cannot be smoothly switched. In addition, when this duty output is in this dead zone, the ideal valve opening / closing timing cannot be obtained, so that sufficient fuel consumption cannot be improved.

  Similar to the first control valve 74, the second control valve 75 can be constituted by a variable electromagnetic spool valve. The second control valve 75 includes a lock port that communicates with a lock channel 63 that is a hydraulic fluid channel of the lock mechanism 6, a supply port that communicates with a channel downstream of the second pump 72, and an oil pan 76. A drain port communicating therewith. The second control valve 75 is a two-position control valve capable of performing two state controls: a lock release control position for communicating the lock port and the supply port, and a lock control position for communicating the lock port and the drain port. It has become. The second control valve 75 is controlled by a control signal from the control unit 8 and operates to control the lock mechanism. The lock flow path 63 that connects the second control valve 75 and the lock mechanism 6 is a flow path that connects the advance flow path 22 or the retard flow path 21 and the first control valve 75 in the phase conversion mechanism 1. The hydraulic oil supply / discharge control for the lock mechanism 6 can be performed independently of the hydraulic oil supply / discharge control for the retard chamber 2A and the advance chamber 2B.

  Between the drive side rotary body 12 and the driven side rotary body 11, a torsion spring 3 is provided as a biasing mechanism that applies a biasing force (assist torque) that biases the driven side rotary body 11 to the advance side. Yes. The driven side rotator 11 tends to be delayed with respect to the drive side rotator 12 due to the resistance received from the valve springs of the intake and exhaust valves and the phase conversion mechanism 1. Therefore, the torsion spring 3 functions to suppress the phase toward the delay, that is, the retard side of the driven-side rotator 11.

  As shown in FIGS. 1 and 3, the torsion spring 3 has one end 3 a fixed to the driving side rotating body 12 and the other end 3 b provided to the driven side rotating body 11. Locked to the radial locking portion 15.

  3 and 4 show the transition of the relative phase between the driving side rotating body 12 and the driven side rotating body 11. 3 and 4 show the most retarded phase and the most advanced phase, respectively. That is, as the driven-side rotator 11 rotates clockwise with respect to the drive-side rotator 12, the most advanced angle phase is reached from the most retarded phase through the intermediate lock phase (see FIG. 2).

  As shown in FIG. 5, the control unit 8 includes a torque estimation unit 81, a target angle determination unit 82, a relative angle calculation unit 83 (an example of the angle information detection unit of the present invention), a deviation calculation unit 84, and a control constant determination unit. 85, a PID control unit 86, a correction unit 87, and a duty output determination unit 88 are configured by software. Further, the control unit 8 includes an oil temperature sensor 91 for measuring the oil temperature of the engine oil (an example of the fluid temperature measurement unit of the present invention), an engine speed sensor 92 for measuring the engine speed, and the throttle opening. Signals from a throttle sensor 93 for measuring, a crank angle sensor 94 for detecting the crank angle, and a camshaft angle sensor 95 for detecting the angle of the camshaft 10 are inputted.

  The torque estimation unit 81 estimates the engine torque desired by the driver based on the engine speed and the throttle opening. In the present embodiment, the torque estimation unit 81 stores a map that associates the engine speed, the throttle opening, and the engine torque, and the torque estimation unit 81 estimates the engine torque based on the map.

  The target angle determination unit 82 determines a target value (hereinafter referred to as a target angle) of the relative angle of the driven-side rotator 11 based on the engine speed and the engine torque estimated by the torque estimation unit 81. In the present embodiment, the target angle determination unit 82 stores a map in which the engine torque and the target angle are associated with each other, and the target angle determination unit 82 determines the target angle based on the map. As described above, in the present embodiment, the first pump 71 is configured as a mechanical hydraulic pump driven by the driving force of the crankshaft of the engine. Can be used as a substitute for force. Therefore, in this configuration, the engine speed sensor 92 functions as a discharge pressure measuring unit.

  The relative angle calculation unit 83 calculates the relative angle of the driven-side rotator 11 based on the crank angle measured by the crank angle sensor 94 and the angle of the camshaft 10 measured by the camshaft angle sensor 95. .

  The deviation calculation unit 84 calculates a deviation (deviation) between the actual relative angle and the target angle from the target angle determined by the target angle determination unit 82 and the relative angle calculated by the relative angle calculation unit 83.

  The control constant determination unit 85 determines a control constant in PID control based on the deviation calculated by the deviation calculation unit 84.

The PID controller 86 uses the deviation calculated by the deviation calculator 84 and the control constant determined by the control constant determiner 85 to calculate a duty output as a control value. In the present embodiment, the PID control unit 86 performs duty duty by (duty output) = (control constant) × (deviation) + (control constant) × (integral value of deviation) + (control constant) × (differential differential value). Assume the output. The control constant determining unit 85 and the PID control unit 86 realize the PID control of the present invention.

  The correction unit 87 has a function of correcting the urging force of the torsion spring 3. The duty output calculated by the PID control unit 86 does not consider the urging force of the torsion spring 3. Therefore, even if the fluid control mechanism 74 is controlled using the duty output calculated by the PID control unit 86, it is not possible to give an appropriate torque to the driven side rotating body 11. Therefore, the correction unit 87 corrects the torque generated by the duty output calculated by the PID control unit 86 based on the urging force of the torsion spring 3. Note that the map control of the present invention is realized by the correction unit 87.

  The duty output determination unit 88 determines a final duty output based on the torque corrected by the correction unit 87.

  FIG. 6 is a flowchart showing the flow of the duty output determination process in the present embodiment. It is assumed that measured values from the oil temperature sensor 91, the engine speed sensor 92, the throttle sensor 93, the crank angle sensor 94, and the camshaft angle sensor 95 are input to the control unit 8 in real time.

  First, the torque estimation unit 81 acquires the engine speed from the engine speed sensor 92 and the throttle opening from the throttle sensor 93, and estimates the engine torque from the map based on those values (# 01). Since the throttle opening is determined according to the depression of the accelerator pedal, the throttle opening is considered to accurately represent the engine torque desired by the driver. Therefore, it is preferable to use the throttle opening as a factor for estimating the engine torque. The estimated engine torque is sent to the target angle determination unit 82.

  The target angle determination unit 82 acquires the engine torque estimated from the torque estimation unit 81, acquires the engine rotation speed from the engine rotation speed sensor 91, and uses these values to determine the relative of the driven-side rotator 11 from the map. A target angle that is a target value of the angle is determined (# 02). The determined target angle is sent to the deviation calculator 84.

  The relative angle calculation unit 83 acquires the crank angle measured by the crank angle sensor 94 and the angle of the camshaft 10 measured by the camshaft angle sensor 95, and based on these values, the current of the driven-side rotator 11 is obtained. Is calculated (# 03). The calculated relative angle is sent to the deviation calculating unit 84 and the correcting unit 87.

  The deviation calculation unit 84 acquires the target angle from the target angle determination unit 82 and the relative angle from the relative angle calculation unit 83, and calculates the deviation based on these values (# 04). Specifically, (relative angle) − (target angle) is calculated. Thereby, the residual with respect to the target angle of the present relative angle is obtained. The calculated deviation is sent to the PID control unit 86.

  The control constant determination unit 85 acquires the oil temperature of the engine oil from the oil temperature sensor 91 and the engine speed from the engine rotation sensor 92, and determines a control constant in PID control based on these (# 05). Since the viscosity of the engine oil is low when the oil temperature is low, there is little leakage from the space 2 and the responsiveness of the driven side rotating body 11 is high. In the present embodiment, the first pump 71 that supplies engine oil to the fluid control mechanism 74 employs a mechanical hydraulic pump that is driven by the driving force of the crankshaft of the engine. There is a characteristic that the response performance of the driven-side rotator 11 is higher as the height is higher. Therefore, it is preferable to determine the control constant based on the oil temperature of the engine oil that determines the response performance of the driven-side rotator 11 and the engine speed. The control constant determined by the control constant determination unit 85 is sent to the PID control unit 86.

The PID control unit 86 acquires the current relative angle deviation from the target angle of the driven rotor 11 from the deviation calculation unit 84, acquires the control constant from the control constant determination unit 85, and based on these values, A duty value as a control value is calculated (# 06). As described above, in the present embodiment, duty output is represented by (duty output) = (control constant) × (deviation) + (control constant) × (integral value of deviation) + (control constant) × (differential value of deviation). Therefore, the duty output can be calculated by substituting the control constant and the deviation into this equation. The calculated duty output is sent to the correction unit 87.

  The torque corresponding to the duty output calculated by the PID control unit 86 is a torque that needs to be actually applied to the driven-side rotator 11. However, as described above, the torsion spring 3 applies a biasing force in the advance direction to the driven-side rotating body 11. Therefore, even if the fluid control mechanism is controlled by the duty output calculated by the PID control unit 86, a desired torque cannot be obtained. Therefore, the correction unit 87 corrects the torque with respect to the duty output calculated by the PID control unit 87 (# 07 to # 08).

  In the present embodiment, the correction unit 87 stores a map in which a relative angle and a correction torque (an example of the correction amount of the present invention) are associated. Accordingly, when the correction unit 86 acquires the relative angle from the relative angle calculation unit 83, the correction torque can be obtained (# 07). The correction torque obtained in this way is added to the torque corresponding to the duty output calculated by the PID controller 86 (# 08).

  FIG. 7 schematically shows this state. The dotted straight line represents the duty output calculated by the PID control unit 86 and the corresponding torque (hereinafter referred to as a pre-correction straight line). In the figure, the duty output D1 is obtained by the PID controller 86, and the torque at that time is T1. On the other hand, the correction unit 87 obtains a correction torque t based on the current relative angle of the driven-side rotator 11. When the straight line before correction is corrected using this correction torque t, it becomes a one-dot chain line in the figure (hereinafter referred to as a straight line after correction). Therefore, the corrected torque (hereinafter referred to as corrected torque) is T2 = T1 + t.

  In the present embodiment, the correction unit 87 stores a correction torque for a relative angle of 5 ° pitch as a map. The correction torque is set so that the amount of change decreases as the relative angle changes from the retard side to the advance side. This is because the urging force of the torsion spring 3 in the advance direction gradually becomes weaker as it is displaced from the retard side to the advance side.

  The correction torque obtained in this way is sent to the duty output determination unit 88.

  The duty output determination unit 88 that has acquired the correction torque from the correction unit 87 determines the duty output based on the correction torque. Therefore, the duty output determining unit 88 stores a correlation created by correcting the correlation between the actually measured duty output and the torque (hereinafter referred to as a corrected correlation), and the corrected torque and the corrected torque are stored. The final duty output is determined using the correlation (# 09).

  FIG. 8 shows the correlation between the actually measured duty output and torque. The duty outputs 0% and 100% are the most advanced phase state and the most retarded phase state, respectively. The correlation shown in the figure represents a state in which the phase is shifted from the most advanced phase state in the retarded direction (dotted line in the figure) and a state in which the phase is shifted from the most retarded phase state in the advanced angle direction (dashed line in the figure). ing. As is apparent from the figure, there is a discontinuity between the displacement in the retard direction and the displacement in the advance direction. In other words, this system has hysteresis. This hysteresis is considered to be an influence of a reaction force such as a cam that the torsion spring 3 and the driven side rotating body 11 receive. If the correlation having such hysteresis is used, the duty control becomes unstable. Therefore, in the present embodiment, a correlation (corrected correlation) in which a portion having the correlation hysteresis in FIG. 8 is replaced with a straight line is used. The solid line in FIG. 7 is the corrected correlation created from the correlation in FIG.

  The duty output determining unit 88 determines a final duty output using the corrected torque and the corrected correlation. Specifically, a duty output serving as a correction torque is obtained from the corrected correlation. In the example of FIG. 7, T2 is obtained as the duty output that becomes the corrected torque T2. Thereby, since it can avoid that a duty output is located in a dead zone, stable control is realizable.

  The fluid control mechanism 74 is driven by the duty output thus determined, and the relative angle of the driven-side rotator 11 is displaced (# 10). This process is repeated until the deviation between the relative angle and the target angle becomes a predetermined threshold value or less. Thereby, control is performed so as to obtain an optimum relative angle.

[Another embodiment]
(1) In the above-described embodiment, the torsion spring 3 has a configuration in which the biasing force works only in a part of the section. However, the biasing force may work from the most retarded phase state to the most advanced angle phase state. It is possible to adopt a configuration in which the biasing force works only in a section different from the embodiment.

(2) In the above-described embodiment, the duty control is performed by the map control and the PID control. However, the control by only one of them may be performed.

(3) In the above-described embodiment, the driven-side rotator 11 is in the most advanced phase state when the duty output is 0%. However, the driven-side rotator 11 is in the most retarded phase state when the duty output is 100%. You may comprise so that it may become.

  The present invention relates to a drive-side rotating body that rotates synchronously with a crankshaft of an internal combustion engine, and a driven-side rotation that is arranged coaxially with the drive-side rotating body and rotates synchronously with a camshaft for opening / closing a valve of the internal combustion engine. The fluid pressure is formed by the body, the drive-side rotator, and the driven-side rotator, and is partitioned into the retard chamber and the advance chamber by a partition provided on at least one of the drive-side rotator and the driven-side rotator. A urging mechanism for urging the driven-side rotator in the advance phase direction with respect to the drive-side rotator, supply of the working fluid from the pump that supplies the working fluid to the fluid pressure chamber, and from the fluid pressure chamber And a fluid control mechanism for controlling the discharge of the working fluid.

2: Space (fluid pressure chamber)
2A: Pressure chamber (retard chamber)
2B: Pressure chamber (advance chamber)
3: Torsion spring (biasing mechanism)
10: Camshaft 11: Driven side rotator 12: Drive side rotator 13: Vane (partition)
71: First pump (pump)
74: First control valve (fluid control mechanism)
94: Crank angle sensor (angle information detector)
95: Camshaft angle sensor (angle information detector)
91: Oil temperature sensor (fluid temperature detector)
92: Engine speed sensor (discharge pressure measuring unit)
8: Control unit

Claims (5)

A drive side rotating body that rotates synchronously with the crankshaft of the internal combustion engine;
A driven-side rotating body that is coaxially disposed with respect to the driving-side rotating body and that rotates synchronously with a camshaft for opening and closing the valve of the internal combustion engine;
A fluid formed by the driving side rotating body and the driven side rotating body and partitioned into a retarded angle chamber and an advanced angle chamber by a partition provided in at least one of the driving side rotating body and the driven side rotating body. A pressure chamber,
An urging mechanism for urging the driven-side rotator in the advance phase direction with respect to the drive-side rotator;
A fluid control mechanism for controlling the supply of the working fluid to the fluid pressure chamber from the pump that supplies the working fluid and the discharge of the working fluid from the fluid pressure chamber;
An angle information detector that detects a relative angle that is a relative phase of the driven-side rotator to the drive-side rotator;
A fluid temperature measurement unit for measuring the temperature of the working fluid;
A discharge pressure measuring unit for measuring the discharge pressure of the pump;
A control unit that controls the fluid control mechanism based on the relative angle, the oil temperature of the working fluid, and the discharge pressure of the pump so that the relative angle becomes a predetermined target angle ;
The fluid control mechanism includes a closing control position where supply of the working fluid to the fluid pressure chamber or discharge of the working fluid from the fluid pressure chamber is stopped, supply of the working fluid to the fluid pressure chamber, or Displace between a retard control position or an advance control position where the working fluid is discharged from the fluid pressure chamber,
When the duty output to the fluid control mechanism is 0%, it is located at one of the retard control position and the advance control position, and when the duty output increases, the retard control position via the closing control position And the value of the torque corresponding to the duty output at the block control position when the valve is displaced to the other one of the advance angle control position and the block angle control position. A valve opening / closing timing control device that is corrected to a torque value corresponding to the duty output and that determines the duty output based on the corrected torque value . The control unit is a control value based on the relative angle output by map control using a map in which the relative angle and a correction amount for correcting the urging force by the urging mechanism are associated with each other;
A duty control based on the oil temperature, the number of rotations of the pump, and the control based on the oil temperature and the discharge pressure output by the PID control that receives the difference between the target angle and the actual angle is performed. 2. The valve opening / closing timing control device according to 1.   The valve opening / closing timing control device according to claim 1 or 2, wherein the target angle is determined from a map based on an engine torque estimated based on an engine speed and a throttle opening and an engine speed.   The valve opening / closing timing control device according to claim 2, wherein a control constant is changed in the PID control. Wherein, upon operating a pre Symbol fluid control mechanism, the valve timing control apparatus according to any one of claims 1 to 4 using a Quito torque before being corrected according to the relative angle.

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