System and method for controlling variable valve time

REFERENCE TO RELATED APPLICATION

This application claims the priority of the Chinese patent application of serial no. 201310000189.X, which was filed on Jan. 4, 2013, and the entire content of which is incorporated herein by reference.

FIELD OF THE INVENTION

The present disclosure relates to the control technology of an actuator, and in particular relates to a variable valve time control system and method for control of an actuator, controlling time error caused by motion delay in the actuator.

BACKGROUND OF THE INVENTION

Because the accuracy of engine variable valve system to control time has higher requirements, but the core of an electro-hydraulic control system control valve is control by a high-speed switch valve, the control precision of a high-speed switch valve is influenced by temperature (because the coil impedance varies with temperature) and voltage change (the operating voltage varies as a car operates electric appliances to a varying degree), therefore the operating current of the high-speed switch valve varies even under the same control signal, causing inaccuracy in operation of the electromagnetic valve. At the same time, the operation of the high-speed switch valve and hydraulic cylinder is influenced by system pressure, temperature-dependent oil viscosity, resulting in variation in the resistance to the movement and thus variation in the operation of the electromagnetic valve.

A direct consequence of these variations is that the opening time of the engine variable valve is influenced by above-mentioned various factors, resulting in a time delay between the actual motion of the variable valve and the reference signal, and this delay time also varies with the change of external factors.

By calculation the switch time and inter-cycle corrective control action, it is possible to effectively eliminate, on the variable valve switch time, the impact from variations or fluctuations in the system pressure, the viscosity of the working fluid and the voltage, resulting in good precision in time control of the engine variable valve system. The above mentioned cycle refers to the thermodynamic cycle of an engine.

The prior art systems typically utilize a closed-loop control method to solve consistency and repeatability issues in the variable valve switch time. Closed-loop control imposes stringent requirements on its hardware (such as a displacement sensor with high precision and fast response, thus high cost, low controllability and deficiency in robustness), which directly affects the engine variable valve system of industrialization.

SUMMARY OF THE INVENTION

In light of the above problems, the present disclosure is to provide a variable valve time control system, which is to achieve precise control over the time error caused by actuator motion delay.

The present disclosure is also to provide a variable valve time control method.

The present disclosure provides a variable valve time control system, which comprises an actuator, an actuation switch valve, an electronic control unit and a displacement sensor. The electronic control unit controls the actuation switch valve. The displacement sensor detects the engine valve displacement signal. The electronic control unit calculates the engine valve opening and closing times based on the displacement signal from the engine valve displacement sensor and converts the engine valve opening and closing times into the corresponding switch times of the actuation-switch-valve power output.

In one embodiment of the present disclosure, the variable valve time control system also comprises a start switch valve, and the electronic control unit also controls the start-switch-valve power output.

In another embodiment of the present disclosure, the variable valve time control system further comprises an electrically-driven fluid pump source, and the electronic control unit provides an electrically-driven fluid pump source power output, thereby controlling the electrically-driven fluid pump source.

In another embodiment of the present disclosure, the electronic control unit also comprises an engine control unit and a valve control unit; based on the engine valve displacement signal from the displacement sensor, the engine control unit calculates the engine valve opening and closing times, which are further translated by the valve control unit into the corresponding switch times of the actuation-switch-valve power output.

In another embodiment of the present disclosure, the electronic control unit detects the working state and power requirement of the engine, determines the engine valve opening and closing times, and, according to the state of the actuator system and its environment, translates the engine valve opening and closing times into the switch times of the actuation-switch-valve power output, resulting in a switch in the power state of the actuation switch valve, thereby controlling the opening and closing actions of the actuator and the engine valve.

In another embodiment of the present disclosure, at least one state parameter of the actuator system and its environment comprises the temperature of the working medium.

In another embodiment of the present disclosure, at least one state parameter of the actuator system and its environment comprises the pressure of the working medium.

The present disclosure also provides a variable valve time control method, applied to the above mentioned variable valve time control system, comprising the following steps:

- to measure the engine-valve displacement profile;
- to determine respectively the engine valve opening and closing startup times according to the engine valve opening-time and closing-time position-thresholds;
- to issue control signal by the valve control unit according to the engine intake and exhaust control algorithm;
- to determine the engine valve opening and closing delays;
- to determine respectively the engine valve opening and closing target-times;
- to obtain respectively the engine valve opening and closing control-deviations according to the engine intake and exhaust control algorithm;
- to determine respectively the engine valve opening and closing signal-times;
- to determine respectively the engine valve opening and closing estimated-delays;
- to determine respectively the engine valve opening and closing time delay-corrections through an inter-cycle iterative algorithm using the engine valve opening and closing time control-deviations; and
- to determine respectively the engine valve opening and closing signal-times of the current cycle.

In another embodiment of the present disclosure, the engine valve control signal is issued by the electronic control unit according to the need of the engine intake and exhaust control algorithm, and the control signal comprises a high-low level switch at the engine valve opening signal time and another high-low level switch at the engine valve closing signal time.

In another embodiment of the present disclosure, the engine valve opening delay is the difference between the engine valve opening start-up time and the engine valve opening signal time; the engine valve closing delay is the difference between the engine valve closing start-up time and the engine valve closing signal time; the engine valve opening time control deviation is the difference between the engine valve opening target time and the engine valve opening start-up time; the engine valve closing time control deviation is the difference between engine valve closing target time and the engine valve closing start-up time; the engine valve opening estimated-delay is the difference between the engine valve opening target time and the engine valve opening signal time; the engine valve closing estimated-delay is the difference between the engine valve closing target time and the engine valve closing signal time.

In another embodiment of the present disclosure, the variable valve time control method is implemented by a proportional-integral-derivative control algorithm.

In another embodiment of the present disclosure, the variable valve time control method is completed by a proportional integral derivative control algorithm. The engine valve opening time control deviation ddt1 is related to the engine valve opening time delay correction ddt3 in the following formula:

$ddt 3 (t) = K_{p} ddt 1 (t) + K_{p} K_{i} \int_{0}^{t} ddt 1 (t) \partial t + K_{p} K_{d} \frac{d \partial t 1 (t)}{\partial t}$

The engine valve closing time control deviation ddt2 is related to the engine valve closing time delay correction ddt4 in the following formula:

$ddt 4 (t) = K_{p} ddt 2 (t) + K_{p} K_{i} \int_{0}^{t} ddt 2 (t) \partial t + K_{p} K_{d} \frac{d \partial t 2 (t)}{\partial t}$

Among them, t is the time variable, Kp the proportional gain, Ki the integral gain and Kd the differential gain.

In another embodiment of the present disclosure, within the inter-cycle iterative algorithm, the engine valve opening estimated-delay is calculated as dt10 (k+1)=dt10 (k)+ddt3 (k), where ddt3 (k) is obtained in the Kth iteration within the control algorithm; the engine valve closing estimated-delay is calculated as dt20 (k+1)=dt20 (k)+ddt4 (k), where ddt4 (k) is obtained in the Kth iteration within the control algorithm, where K is an integer.

The variable valve time control system and method of the present disclosure is able to accurately calculate the engine valve opening/closing time delay, and thus guarantee to reduce to a minimum the difference between the engine valve opening/closing target times and the engine valve opening/closing start-up times and to eliminate disturbances to the engine valve time control, so as to ensure an accurate control of the variable valve opening/closing time.

What described above is only a summary of the technical solutions of the present disclosure. In order to provide a better understanding of the technical means of the present disclosure so that the present disclosure can be practiced according to this description and in order to make the aforesaid and other objectives, features and advantages of the present disclosure more apparent, the present disclosure will be detailed hereinbelow with reference to embodiments thereof in conjunction with the attached drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic diagram of the control system of hydraulic actuator a preferred embodiment of the present disclosure.

FIG. 2 is the spring control cylinder is shown in FIG. 1 of the actuator hydraulic in high lift conditions, while engine valve is opened to the maximum when the schematic diagram.

FIG. 3 is shown in FIG. 1 of the actuator hydraulic spring cylinder in high lift conditions, while engine valve is in the closed state of the schematic structure.

FIG. 4 is the engine valve control with variable delay time calculation diagram of principle.

FIG. 5 is a variable valve turn on delay time algorithm block diagram.

FIG. 6 is a block diagram off delay time algorithm with variable valve.

DETAILED DESCRIPTION OF THE INVENTION

Hereinbelow, some embodiments of the present disclosure will be detailed with reference to the attached drawings so that advantages and features of the present disclosure can be better understood by those skilled in the art. However, the scope of the present disclosure is not limited to these embodiments.

Please refer to FIGS. 1 to 3, the engine valve control system, one embodiment of this present disclosure, includes an actuator 10, a displacement sensor 961, an electronic control unit 90, an actuation switch valve 81, a start switch valve 82, a high-pressure hydraulic fluid source 70, and a low-pressure hydraulic fluid source (not shown).

The actuator 10 comprises a housing 200. In the housing and in the second direction along a longitudinal axis 110, i.e. in the direction from the top to the bottom, the actuator 10 further comprises a start port 260, a cavity 250, a first control passage 271, a first upper port 211, a second upper port 212, an actuation cylinder 230, a fluid bypass 240, a second lower port 222, a first lower port 221, and a second control passage 272. The actuator 10 further comprises a first spring system situated in the cavity 250, a first piston rod 410 in the first control passage 271, an actuation piston 300 situated in the actuation cylinder 230 and the fluid bypass 240, a second piston rod 420 situated in the second control passage 272, a second spring system, and an engine valve 700.

The first piston rod 410 comprises, in order of closeness to the actuation piston 300 (namely in the first direction, i.e., from the bottom towards the top in the drawings), a first-piston-rod first neck 411, a first-piston-rod first shoulder 412, a first-piston-rod second neck 413, and a first-piston-rod second shoulder 414. The second piston rod 420 comprises, in order of closeness to the actuation piston 300 (namely in second direction, i.e. from the top towards the bottom in the drawings), a second-piston-rod first neck 421, a second-piston-rod first shoulder 422, a second-piston-rod second neck 423, and a first-piston-rod second shoulder 424.

The actuation cylinder 230 includes a first fluid space defined by an actuation-cylinder first end 231 and an actuation-piston first surface 310, and a second fluid space defined by an actuation-cylinder second end 232 and an actuation-piston second surface 320.

The actuation cylinder 230 is in-between the actuation-cylinder first and second ends 231 and 232, the fluid bypass 240 is in-between a first edge 241 and a second edge 242, and the fluid bypass 240 provides a hydraulic short circuit in the majority of the length of the actuation cylinder 230. Fluid is able to flow between the first and second fluid spaces with a substantially low resistance because of the hydraulic short circuit, with the entire actuation cylinder 230 under a generally equal hydraulic pressure. The hydraulic short circuit ceases to function when the actuation-piston first surface 310 moves over the first edge 241 of the fluid bypass in the first direction, or when the actuation-piston second surface 320 moves over the second edge 242 of the fluid bypass in the second direction. The longitudinal space between the first edge 241 of the fluid bypass and the actuation-cylinder first end 231 is a first effective hydraulic chamber. The longitudinal space between the second edge 242 of the fluid bypass and the actuation-cylinder second end 232 is a second effective hydraulic chamber. The fluid bypass 240 is in effect when the actuation piston 300 does not engage with any of the first and second effective hydraulic chambers.

The first spring system comprises a top actuation spring 512, a spring retainer 511, a spring-control cylinder body 513 and a plunger 514. The top actuation spring 512 is installed between the spring retainer 511 and the spring-control cylinder body 513. The spring retainer 511 is connected to the first piston rod 410 and fixed by valve keys 515. There is a fluid chamber 5133 in the spring-control cylinder body 513. The plunger 514 extends into the fluid chamber 5133. In the plunger 514, there is a flow passage 5141 providing fluid communication between the fluid chamber 5133 and the start port 260.

In this embodiment the top actuation spring 512 is designed overhead and concentric with the first piston rod 410; and the plunger 514, stationary and extending inside the flow passage 5141, is designed to guide the reciprocating motion of the spring-control cylinder body 513 and to distribute hydraulic fluid as the top actuation spring 512 is compressed. The advantages are as follows: it can avoid lengthwise over-extension of the first piston rod 410 caused by the spring-control mechanism (the spring retainer 511) and the effective spring work stroke when the top actuation spring 512 and the first piston rod 410 are not only concentric but also overlapped, so that one can reduce the length of the first piston rod 410 and also its diameter and mass, which leads to a reduction in the mass of the moving parts of the whole actuator, an increase in actuator velocity and a decrease in energy consumption. The control mechanism of the top actuation spring is compact, and the guidance is stable and reliable, thus to avoid lateral force in its process of compressing the top actuation spring 512. Both end segments of the piston rods are supported the housing so as to maximize the support length of the piston rods and minimize the lateral torque on the piston rods during their travel, thus improving the stability of the actuator.

The second spring system comprises a valve spring retainer 521, an engine valve return spring (also called bottom actuation spring) 522, a valve guide 524 and a cylinder head block 523. The valve spring retainer 521 is connected to one end of a valve stem 730, and the other end of the valve stem 730 is connected to the engine valve head 710. The cylinder head block 523 is located in-between the valve spring retainer 521 and the engine valve head 710, the valve guide 524 is installed in the cylinder head block, and the valve stem 730 goes through the valve guide. The bottom actuation spring 522 is installed around the valve stem 730 and supported by the cylinder head block 523 and the valve spring retainer 521.

The second upper port 212 is in connection with a first snubber. The first snubber comprises, in parallel, a first check valve 612, a first throttle orifice 613 and a first relief valve 614. The second lower port 222 is in connection with a second snubber. The second snubber comprises, in parallel, a second check valve 622, a second throttle orifice 623 and a second relief valve 624. The check valves are intended to supply pressurized fluid in their respective open directions and to cut off the backflow in the opposite direction, thus to form a snubbing chamber. The throttle orifices are intended to throttle for snubbing. One is to set-up a reasonable cross-section area for the throttle orifices in order to obtain soft and stable seating at the final stage of snubbing for the piston rod, and also to reduce the sensitivity of snubbing to temperature. The relief valves are intended to limit the peak pressure in the snubber, regulating snubbing time. A relief valve with an adjustable relief pressure may be preferred so that the peak pressure of the snubber can be controlled according to load conditions, to avoid the effects of high pressure pulse on the system. At the same time, the dynamic response of the relief valves has to be fast to effectively regulate the velocity of the first piston rod as the rod enters the snubbing phase.

In order to simplify the design, one can eliminate the first and second relief valves 614, 624, so that the first snubber comprises only the first check valve 612 and the first throttle orifice 613 in parallel, and the second snubber comprises only the second check valve 622 and the second throttle orifice 623 in parallel.

At least one first throttle slot 4121 is cut on the first-piston-rod first shoulder 412 next to the end surface of the first-piston-rod second neck 413. The throttle area of the first throttle slot 4121 is variable, being gradually smaller in the second direction. At the end of the second-piston-rod first shoulder 422, close to the second-piston-rod second neck 423, there is at least one second throttle slot 4221. The throttle area of the second throttle slot 4221 is variable, being gradually smaller in the first direction. The throttle area of the throttle slots is designed to be variable so as to achieve stable snubbing process for the piston rod.

The first throttle slot 4121 works with the first snubber to help achieve a stable snubbing process as the first piston rod 410 terminates its travel in the first direction. The second throttle slot 4221 works with the second snubber to help achieve a stable snubbing process as the second piston rod 420 terminates its travel in the second direction.

The high-pressure hydraulic fluid source 70 comprises a pump 71, a high-pressure regulator 73, a high-pressure accumulator 74, a high-pressure supply line 75 and a tank 72. The high-pressure hydraulic fluid source 70 supplies necessary hydraulic flow at high P_H. The pump 71 delivers, via the high-pressure supply line 75, the hydraulic fluid from the tank 72 to the rest of the system. The high pressure P_H was controlled by the pressure control valve 73. The high-pressure accumulator 74 helps reduce pressure and flow fluctuations, and it is optional according to the total capacity or elasticity of the system, flow distribution balance and/or functional needs. The pump 71 can be of a variable or fixed displacement, with the former providing better energy efficiency. The pressure control valve 73 regulates pressure according to the functional needs and/or energy efficiency.

The low-pressure hydraulic fluid source (not shown) refers to the downstream part of the system or fluid return line, and the pressure P_L is either a pressure arising from flow passage resistance or a pressure regulated by a back-up pressure regulation means, and is slightly higher than the atmospheric pressure or the tank pressure.

The actuation switch valve 81 and the start switch valve 82 feed the ports of the hydraulic actuator 10 via suitable flow supply passages. As shown in FIG. 1, the start switch valve 82 is a two-position three-way valve. It is three-way because the valve has three external ports: one connected with a fluid line 190, and the other two connected with the high pressure P_H pipeline and the low pressure P_L pipeline. The start switch valve 82 is a two-position valve because it has two stable control positions, symbolized by the left and right boxes in its graphical symbol. When its solenoid is not energized, it is secured at the left position by the spring force of the return spring, and this is also known as the natural or default position. The right position is secured by energizing the solenoid. On the left and right positions, the start switch valve 82 connects the fluid line 190 with the low pressure P_L and high pressure P_H pipelines, respectively.

The actuation switch valve 81 is a two-position four-way valve, which is connected with four external hydraulic lines: the low pressure P_L pipeline, the high pressure P_H pipeline, a fluid line 192 and a fluid line 194. The default position of the actuation switch valve 81 is the right position secured by its return spring, and its other position is the left position secured by an electromagnetic force. In its default or right position, the actuation switch valve 81 connect the fluid line 192 and the fluid line 194, with the low pressure P_L pipeline and the high pressure P_H pipeline respectively. When the actuation switch valve 81 is in the left position, the connection order is switched.

The electronic control unit 90 controls the actuation switch valve 81, the start switch valve 82 and an electrically-driven fluid pump source (not shown). The electronic control unit 90 comprises an engine control unit (ECU or Engine Control Unit) 91 and a valve control unit 92. The engine control unit 91 controls the whole engine operation, and the engine control unit 91 is also responsible for calculating engine valve opening and closing times, whose signal is transferred to the valve control unit 92 through the wire harness 931. The valve control unit 92 is responsible for converting the control signal from the engine control unit 91 into a power current- or voltage-output, where the power output comprises three components: the first one being the actuation-switch-valve power output (Output_S1) 951, the second one being the start-switch-valve power output (Output_S2) 952, and the third one being electrically-driven fluid pump source power output (Output_M1) 953.

In any embodiment without a start switch valve, the start-switch-valve power output (Output_S2) 952 can be eliminated; in any embodiment without an electrically-driven fluid pump source, the electrically-driven fluid pump source power output (Output_M1) 953 can be eliminated.

The division between the engine control unit 91 and the valve control unit 92 in their functions and circuits is not absolute; if needed, the valve control unit 92 can be combined into the engine control unit 91.

In this embodiment, the displacement sensor 961 transmits the detected displacement signal through the input of the electronic control unit and to the valve control unit 92, in order to obtain the engine valve opening or closing start-up time. The displacement sensor 961 may use various sensing mechanisms, and also be placed in various locations within the actuator. For example, one may use a differential variable reluctance sensor (DVRT), and install it as shown in FIG. 1 near the spring retainer 511 to detect its displacement variation, with the spring retainer 511 moving together with the engine valve 700. Also the displacement sensor 961 may be placed in a proper location to directly measure the displacement of the valve stem 730.

In other embodiments (not shown), one may utilize other (i.e. non-displacement type) sensors to indirectly infer the engine valve displacement information. For example, two pressure sensors can be used to detect pressure variation at one upper port and one lower port of the actuator, thereby determining the engine valve position according to the correlation between this pressure variation and the engine valve displacement, and thus obtaining the engine valve start-up opening and closing times.

The pump 71 is preferably directly driven by the engine, which can be realized through a mechanical transmission (not shown) appropriately connected with the output shaft or crankshaft of the engine. At the engine start-up, the pump 71 needs to work effectively to pressurize the hydraulic circuit; and if necessary the above-mentioned electrically-driven fluid pump source may be added. The electrically-driven fluid pump source can be an electrical power assisted device (not shown) to assist the pump 71, or an electrically-driven fluid pump (not shown). The electrically-driven fluid pump may be either in series or parallel with the pump 71. Operational control of the electrically-driven fluid pump source is based on the operation state of the engine, which is obtained by the engine control unit 91. The engine control unit 91 issues the demand pressure signal via the wire harness 931, the valve control unit 92 converts the input signal into the electrically-driven fluid pump source power output (Output_M1) 953 to control the operation of the electrically-driven fluid pump source. The system can dynamically control flow through the electrically-driven fluid pump source, thereby controlling the system pressure or high pressure P_H, replacing the high-pressure regulator 73 and its pressure regulating function.

The actuator 10 provides two steps for the engine valve lift. The small lift is mainly used for engine starting and low speed and low load conditions, the high lift is mainly used in high speed and high load conditions. The switch between the small and high lifts is realized primarily through the start switch valve 82.

Referring to FIG. 1, when the start switch valve 82 is in the default state, the start port 260 is in connected with the low-pressure hydraulic fluid source; the actuator 10 is in the small lift operating mode (namely, the spring-control-cylinder upper surface 5131 and the cavity first limit surface 251 are in contact); the top actuation spring 512 has a certain amount or a limited amount of spring compression; the steady-state net spring force from the top and bottom actuation springs 512, 522 situates the actuation piston 300, the piston rods 410, 420 and the engine valve 700 to or around a position shown in FIG. 1; and the engine valve 20 is set at its closed position, with a desired contact force.

Referring to FIG. 2, when the engine is switched from a low load condition to a high load condition, the engine control unit 91 issues a switching signal to the valve control unit 92; the valve control unit 92 sends a switch signal to the start switch valve 82; the start switch valve 82 is switched to the right position; the start port 260 is connected to the low-pressure hydraulic fluid source 70; the hydraulic fluid flows through the flow passage 5141 into the fluid chamber 5133; the spring-control cylinder body 513 is pressured downward; the engine valve is switched from a small lift mode to a high lift mode; the spring-control-cylinder lower surface 5132 is in contact with the cavity second limit surface 252; and the top actuation spring 512 is compressed to a larger extent. At the engine start-up, the two switch valves 81, 82 are in their default positions.

The start port 260 is connected to the low pressure P_L pipeline, the upper ports (comprising the first upper port 211 and second upper port 212) is connected to the low pressure P_L pipeline, and the lower ports (comprising the first lower port 221 and the second lower port 222) is connected to the high pressure P_H pipeline. Immediately after the system start-up, the lower chamber, i.e., the second fluid space, of the actuator is at the system high pressure P_H, and the engine valve is in the closed state.

Referring to FIG. 4, the opening and closing operations of the engine valve are to be explained. To open the engine valve, the engine control unit 91 checks the operation status and power requirement of the engine and determines the engine valve opening time when the engine thermodynamic cycle necessitates the opening of the engine valve 700. As shown by the control signal 280 in FIG. 4, the engine control unit 91 delivers through the wire harness 931 a trigger signal at the trigger time 281 according to the variable valve time control method. Based on the trigger signal from the wire harness 931 and the displacement sensor signal (Input_D1) 941, the environment temperature signal (Input_T1) 942, the working-medium temperature signal (Input_T2) 943 and the working-medium pressure signal (Input_P) 944 from the wire harness 933, the valve control unit 92 delivers the actuation-switch-valve power output Output_S1, the start-switch-valve power output (Output_S2) 952 and the electrically-driven fluid pump source power output (Output_M1) 953. When the actuation-switch-valve power output Output_S1 is at high electrical potential, the actuation switch valve 81 is switched to an energized state and thus its left position, thereby the upper ports being connected with the high-pressure hydraulic fluid source 70 connectivity, and the lower ports being connected with the low-pressure hydraulic fluid source. The displacement sensor signal (Input_D1) 941 is from the displacement sensor 961, the environment temperature signal (Input_T1) 942 is from the environment temperature sensor (not shown), the working-medium temperature signal (Input_T2) 943 is from the working medium temperature sensor (not shown), and the working-medium pressure signal (Input_P) 944 is from the working-medium pressure sensor (not below). The working medium pressure can be directly the system high pressure P_H or a composite or average value of more than one different pressure values. The system or environmental state parameters of the actuator (comprising the ambient temperature, the medium temperature and the medium pressure) influence to a varying degree the response or working state of the actuator; the electronic control unit will selectively consider their influences in the control process.

As shown by the engine-valve displacement curve 255 in FIG. 4, the engine valve opens across the engine valve opening-time position-threshold X1 (usually, but not limited to, 1% to 3% of the engine valve full lift) at the engine valve opening start-up time t2, and it closes across the engine valve closing-time position-threshold X2 (usually, but not limited to, 99% to 97% of the engine valve full lift) at the engine valve closing start-up time t4.

The closing process of the engine valve 700 is in fact a reverse of the above described opening process. The closing process is initiated by switching the actuation switch valve 81 to its default or right position as shown in FIG. 1. In the end, the hydraulic actuator 10 and the engine valve 700 return to their default state as shown in FIG. 1.

The present disclosure provides a variable valve time control method, comprising the following step: to determine the engine valve opening start-up time t2 and the engine valve closing start-up time t4 according to the engine valve opening-time position-threshold X1 and the engine valve closing-time position-threshold X2. In FIG. 4, the engine-valve displacement curve 255 is the actual engine-valve displacement curve, which is measured using the displacement sensor 961 in this embodiment. One can also use other methods to obtain the related position points to determine the opening and closing status of the engine valve.

The engine-valve control signal 280 in FIG. 4 is issued by the valve control unit 92 based on the need of the engine intake and exhaust control algorithm. The signal can be in one of a variety of formats, including the TTL format. The control signal comprises a high-low level switch at the engine valve opening signal time t1 and another high-low level switch at the engine valve closing signal time t3.

After the engine valve opening signal is issued, due to the influences from the electromagnetic coils of the actuation switch valve 81, the hydraulic fluid and other factors, the engine valve actually opens up at the engine valve opening start-up time t2, the time delay between the time t1 and the time t2 is the engine valve opening delay dt1=t2−t1. Similarly, the time delay between the time t3 and the time t4 is the engine valve closing delay dt2=t4−t3.

Before the engine valve start-up opening or closing, the engine intake control algorithm determines the engine valve opening target time t20 or the engine valve closing target time t40. By the time t20, the engine valve needs to start the opening process. By the time t40, the engine valve needs to start the closing process.

Due to the existence of various factors, the time the engine valve reaches the engine valve opening-time position-threshold X1 deviates from the target time t20, i.e., there is generally a deviation between the engine valve opening start-up time t2 and the engine valve opening target time t20. This deviation is the engine valve opening time control deviation ddt1, calculated as ddt1=t20−t2. Similarly, the engine valve closing time control deviation ddt2 is calculated as ddt2=t40−t4.

In the engine-valve opening control algorithm, the input to the engine valve time control algorithm is the engine valve opening target time t20, from which one calculates the engine valve opening signal time t1. The difference between the engine valve opening target time t20 and the engine valve opening signal time t1 is the engine valve opening estimated-delay dt10, i.e. dt10=t20−t1.

In the engine-valve closing control algorithm, the input to the engine valve time control algorithm is the engine valve closing target time t40, from which one calculates the engine valve closing signal time t3. The difference between the engine valve closing target time t40 and the engine valve closing signal time t3 is the engine valve closing estimated-delay dt20, i.e. dt20=t40−t3. The main factors influencing the delays dt1, dt2 are: the environment temperature T1, the working medium temperature T2, and the working medium pressure P. Variation in these factors is typically continuous, not abrupt, so that the opening estimated-delay dt10 and the closing estimated-delay dt20 from the last control cycle may be used to correct deviations in a new cycle.

FIGS. 5 and 6 are the block diagrams for the engine valve opening and closing algorithms respectively, their cores are the engine-valve opening control algorithm 258 and the engine-valve closing control algorithm 259 respectively, and their implementation is performed in the valve control unit 92. In this embodiment, the control algorithms 258, 259 use an adaptive PID (proportional integral differential) control, and the PID control comprises the proportion unit (P), the integral unit (1) and the differential unit (D). The inputs to the engine valve opening and closing control algorithms 258, 259 comprise the engine valve opening time control deviation ddt1 and the engine valve closing time control deviation ddt2 respectively, and the outputs comprise the engine valve opening time delay correction ddt3 and the engine valve closing time delay correction ddt4 respectively, with their formulae in equations (1-1) and (1-2). Within the control algorithm, the proportional, differential and integral units interact with each other, and one can adjust the proportional gain Kp, the integral gain Ki and the differential gain Kd to adjust the response, error and overshoot performance of the control algorithm.

$\begin{matrix} ddt 3 (t) = K_{p} ddt 1 (t) + K_{p} K_{i} \int_{0}^{t} ddt 1 (t) \partial t + K_{p} K_{d} \frac{d \partial t 1 (t)}{\partial t} & (1 - 1) \\ ddt 4 (t) = K_{p} ddt 2 (t) + K_{p} K_{i} \int_{0}^{t} ddt 2 (t) \partial t + K_{p} K_{d} \frac{d \partial t 2 (t)}{\partial t} & (1 - 2) \end{matrix}$

Due to a relatively high sensitivity of the engine valve actuator 10, the actuation switch valve 81 and the start switch valve 82 to the external environment temperature T1, the medium temperature T2, and the medium pressure P, the characteristics of the system will change with varying environmental conditions. In order to maintain control stability, within the control algorithms 258, 259, one may selectively modify the proportional gain Kp, the integral gain Ki and the differential gain Kd in response to variation in the external environment temperature T1, the working medium temperature T2 and/or the working medium pressure P. Different temperatures and pressures vary in their influence, and variables with smaller influence can be ignored.

As shown in FIGS. 5 and 6, in the control algorithms 258, 259, the parameter tuning in the PID algorithm is obtained in the development process using the common engineering tuning method, which is further revised through experiment, resulting in an pre-stored parameter adjustment table. This table is used in application to adjust proportional, integral and differential gain parameters in response to variation in the external environment temperature T1, the working medium temperature T2, the working medium pressure P.

In this embodiment, the control algorithms 258, 259 are a PID control algorithm. In practice, one can also use other control algorithms to adjust the engine valve opening time control deviation ddt1 and the engine valve closing time control deviation ddt2.

With the control algorithms 258, 259, one is able to estimate, with the current cycle information and for the next cycle, the engine valve opening time delay correction (or iterative weight) ddt3 and the engine valve closing time delay correction (or iterative weight) ddt4.

With an inter-cycle iterative computation method, the engine valve opening estimated-delay dt10 is calculated as: dt10(k+1)=dt10(k)+ddt3(k), where ddt3 (k) is calculated in the Kth iteration within the control algorithm 258. Similarly, the engine valve closing estimated-delay dt20 is calculated as: dt20(k+1)=dt20(k)+ddt4(k), where ddt4 (k) is calculated in the Kth iteration within the control algorithm 259.

Thus, knowing the engine valve opening target time t20 and the engine valve closing target time t40, one can calculate new value of the engine valve opening signal time t1: t1 (k+1)=t20 (k+1)−dt10 (k+1), and new value of the engine valve closing signal-time t3: t3 (k+1)=t40 (k+1)−dt20 (k+1).

In summary, using the variable engine valve time control system and method in this disclosure to accurately calculate the engine valve opening and closing time delays, one may minimize the difference between the engine valve opening and closing target times (t20 and t40) and the engine valve opening and closing start-up time (t2 and t4), and eliminate the disturbing factors to the engine valve time control, so as to ensure precise control of the variable valve opening and closing times.

The actuator 10 in this disclosure can be substituted, in its structure or composition, with the actuators disclosed in the United States patents numbered as U.S. Pat. No. 7,302,920, U.S. Pat. No. 7,194,991, U.S. Pat. No. 7,156,058, U.S. Pat. No. 7,290,509, U.S. Pat. No. 7,213,549, and U.S. Pat. No. 7,370,615 and entitled as “variable valve actuator,” the Chinese patent numbered as 201210095184.5 and entitled as “variable valve actuator” (or its equivalent U.S. patent application Ser. No. 13/850,372 and entitled as “variable valve actuator”), the Chinese patent application numbered 201210323412.X (replacing an earlier number 201210095178.X) and entitled as “variable valve actuator switch time measurement method,” and the Chinese patent application numbered 201210201497.4 (replacing an earlier number 01210161831.8) and entitled as “variable valve actuation system and control methods,” using the control process and displacement measurement method similar to those in this disclosure, details of which are not to be repeated here again.

What described above are only embodiments of the present disclosure, but are not intended to limit the present disclosure in any form. Although the present disclosure has been described above with reference to the embodiments thereof, these embodiments are not intended to limit the present disclosure. People skilled in the art can make slight alterations or modifications as equivalent embodiments on the basis of the above disclosures without departing from the scope of the present disclosure. However, any alterations, equivalent changes and modifications made to the above embodiments according to the technical spirits of the present disclosure and without departing from the scope of the present disclosure shall all be covered within the scope of the present disclosure.

System and method for controlling variable valve time

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