Intelligent smooth shifting method and system for bulldozer based on industrial mechanism model

Abstract

The application belongs to the technical field of automatic control of industrial machines, and specifically discloses a bulldozer intelligent smooth gear shifting method and system based on an industrial mechanism model, which comprises the following steps: collecting cylinder pressure and related rotating speed, substituting the collected cylinder pressure and related rotating speed into a longitudinal dynamics model, solving a transient load gradient in real time, and calculating a target engagement pressure reference value based on static pressure balance and oil temperature correction; combining viscosity-temperature characteristics and an orifice throttling model to determine an actual volume flow rate of an electromagnetic valve, and outputting fast oil charging time to eliminate a clutch gap; after entering a sliding friction overlap stage, calculating clutch transmission torque according to the Coulomb friction law, taking the total torque of the turbine as a constraint, and performing closed-loop regulation on the electromagnetic valve of the current and target gears until the oil pressure reaches the reference value to complete gear shifting. The application realizes adaptive power transfer under a full temperature range and complex load, and effectively eliminates gear shifting impact.

Description

Technical Field

[0001] This invention belongs to the field of industrial machinery automatic control technology, and relates to a method and system for intelligent and smooth gear shifting of bulldozers based on an industrial mechanism model. Background Technology

[0002] Bulldozers are typical hydraulic-mechanical transmission engineering vehicles, and their gear shifting process involves the uninterrupted transfer of power from one gear clutch to another. The quality of gear shifting directly affects work efficiency, operator comfort, and the service life of key components in the transmission system.

[0003] In existing technologies, bulldozer shift control generally employs a lookup table method based on empirical calibration. This involves determining fixed-ratio clutch build-up and release curves through extensive real-vehicle calibration tests under different engine speeds, turbine speeds, and oil temperatures. These curves are pre-stored in the transmission controller and directly retrieved during shifting, relying on preset slopes for open-loop control of the clutch oil pressure. Some solutions introduce single-variable feedback compensation on top of feedforward control, such as correcting only the turbine speed change rate, to mitigate shift shock.

[0004] The existing technical solutions described above have the following technical defects: First, the calibration parameters are not adaptable enough to complex and variable load environments. When the cutting resistance of the shovel, the inertial load of the whole machine, or the viscosity of the oil deviates from the calibration conditions, the fixed pressure build-up and pressure relief curves cannot guarantee the real-time balance between the clutch-transmitted torque and the turbine output torque, which can easily lead to shift shock or power interruption. Second, temperature changes have a significant impact on the viscosity of hydraulic oil and the dynamic friction coefficient of the friction plates. The existing lookup table scheme only performs piecewise interpolation processing at a limited number of temperature points, lacking compensation at the physical mechanism level, making it difficult to guarantee shift smoothness under extreme conditions in a wide temperature range. Third, the existing open-loop timing control cannot compensate for the differences in oil pressure response caused by time-varying factors such as changes in clutch disengagement clearance, friction plate wear, and hydraulic system leakage. During the oil filling stage, it is easy to cause rigid impact of the piston on the friction plates, and the shift quality and actuator life will gradually deteriorate after long-term use. Summary of the Invention

[0005] In order to overcome the above-mentioned defects of the prior art and to achieve the above objectives, the present invention proposes the following technical solution: a bulldozer intelligent smooth shifting method based on an industrial mechanism model, comprising: S1, when the shifting command is triggered, collecting the rodless chamber pressure of the bulldozer blade lifting cylinder, the turbine speed of the hydraulic torque converter and the speed of the track drive wheel, and substituting the collected pressure and speed into the longitudinal dynamic mechanism model to solve for the transient load gradient.

[0006] S2. Based on the transient load gradient and the structural parameters of the target gear clutch, calculate the minimum engagement oil pressure of the target gear clutch according to the hydrostatic pressure balance relationship, and use it as the target engagement pressure reference value.

[0007] S3. Collect the hydraulic oil temperature in the transmission oil pan, determine the real-time dynamic viscosity of the hydraulic oil based on the mapping relationship between oil temperature and dynamic viscosity, and calculate the actual volumetric flow rate of the shift ratio solenoid valve under unit control quantity by combining the main oil circuit pressure collection value of the transmission and the orifice throttling flow relationship.

[0008] S4. Obtain the clearance oil filling volume parameter of the target gear clutch, divide it by the actual volumetric flow rate to obtain the fast oil filling time under the current working condition, and output a full duty cycle control signal to the target gear clutch solenoid valve based on the fast oil filling time.

[0009] S5. After the fast oil charging time ends, the slip friction overlap stage is entered. The real-time discharge pressure of the current gear clutch solenoid valve and the real-time pressure build-up pressure of the target gear clutch solenoid valve are obtained. Based on Coulomb's law of friction and the characteristic parameters of the clutch friction pair, the residual transmission torque of the current gear clutch attenuation and the new transmission torque generated by the target gear clutch are calculated respectively.

[0010] S6. Using the transient total torque output by the hydraulic torque converter turbine as the constraint total, the current gear clutch solenoid valve and the target gear clutch solenoid valve are closed-loop regulated to ensure that at any time during the slip-overlap stage, the sum of the residual transmitted torque and the newly transmitted torque remains equal to the transient total torque. When the target gear clutch oil pressure rises to the target engagement pressure reference value, the current gear clutch pressure is forced to zero to complete the gear shift.

[0011] The second aspect of the present invention provides a bulldozer intelligent smooth shifting system based on an industrial mechanism model, comprising: a transient load calculation module, used to collect the rodless chamber pressure of the bulldozer blade lifting cylinder, the turbine speed of the hydraulic torque converter, and the speed of the track drive wheel when the shifting command is triggered, and to substitute the collected pressure and speed into the longitudinal dynamic mechanism model to solve for the transient load gradient.

[0012] The target pressure calculation module is used to calculate the minimum engagement oil pressure of the target gear clutch based on the transient load gradient and the structural parameters of the target gear clutch, according to the hydrostatic pressure balance relationship, and serve as the target engagement pressure reference value.

[0013] The volumetric flow rate calculation module is used to collect the hydraulic oil temperature in the transmission oil pan, determine the real-time dynamic viscosity of the hydraulic oil based on the mapping relationship between oil temperature and dynamic viscosity, and calculate the actual volumetric flow rate of the shift ratio solenoid valve under a unit control quantity by combining the main oil circuit pressure of the transmission and the orifice throttling flow relationship.

[0014] The fast charging timing control module is used to obtain the gap oil filling volume parameter of the target gear clutch, divide it by the actual volume flow rate, obtain the fast charging time under the current working condition, and output a full duty cycle control signal to the target gear clutch solenoid valve based on the fast charging time.

[0015] The slip torque calculation module is used to enter the slip overlap stage after the fast oil charging time ends. It obtains the real-time discharge pressure of the current gear clutch solenoid valve and the real-time pressure build-up pressure of the target gear clutch solenoid valve. Based on Coulomb's law of friction and the friction pair characteristic parameters of the clutch, it calculates the residual transmission torque of the current gear clutch attenuation and the new transmission torque generated by the target gear clutch.

[0016] The torque closed-loop adjustment module is used to perform closed-loop adjustment of the current gear clutch solenoid valve and the target gear clutch solenoid valve, using the transient total torque output by the hydraulic torque converter turbine as the constraint total. This ensures that at any moment during the slip-overlap phase, the sum of the residual transmitted torque and the newly transmitted torque remains equal to the transient total torque. When the target gear clutch oil pressure rises to the target engagement pressure reference value, the current gear clutch pressure is forced to zero, completing the gear shift.

[0017] Compared with the prior art, the beneficial effects of the present invention are as follows: (1) When the shift command is triggered, the present invention collects real-time working parameters such as the pressure of the bulldozer blade lifting cylinder, the speed of the hydraulic torque converter turbine, and the speed of the track drive wheel, and substitutes them into the longitudinal dynamic mechanism model to solve the transient load gradient. This design overcomes the shortcomings of the traditional empirical calibration lookup method in adaptability under complex load conditions, and gives the shift control strategy the ability to adapt to the dynamic changes of external transient working load and the inertial force of the whole machine, thereby improving the smoothness of the shift process under heavy load conditions.

[0018] (2) This invention collects the temperature of the transmission hydraulic oil, performs temperature compensation of the dynamic friction coefficient of the engagement oil pressure of the target gear clutch according to the hydrostatic pressure balance equation, and calculates the actual volumetric flow rate based on the empirical relationship of viscosity-temperature characteristics and the orifice throttling model. The above mechanism provides feedforward compensation at the physical mechanism level for the temperature drift characteristics of the hydraulic transmission system, ensuring the calculation accuracy of the clutch target engagement pressure reference value and fast charging oil sequence under wide temperature range and different oil supply pressure conditions, and preventing power handover delay caused by changes in the physical properties of the oil.

[0019] (3) In the slip-overlap stage, this invention calculates the transmission torque of the clutch between the current gear and the target gear in real time based on Coulomb's law of friction, and uses the transient total torque output by the hydraulic torque converter turbine as a global constraint variable, driving the proportional-integral-derivative controller to output a compensation slope according to the deviation value. This torque-closed-loop-based collaborative adjustment mechanism forces the sum of the torques of the disengaged and engaged clutches to always match the real-time power output of the transmission system, preventing power interruption and torque surge in the transmission system during gear shifting.

[0020] (4) This invention precisely customizes the fast-charging time based on the clutch clearance oil volume parameters and fluid volumetric flow rate, and seamlessly connects to closed-loop monitoring of pressure characteristic values ​​after the fast-charging time is reached. This composite control strategy ensures that the clutch piston can quickly pass the separation gap to shorten the shift delay, while switching to slip control state in a timely manner through real-time pressure inflection point determination, effectively avoiding rigid mechanical impact of the piston on the friction plate, and taking into account both the vehicle's power response speed and the service life of the clutch friction components. Attached Figure Description

[0021] To more clearly illustrate the technical solutions of the embodiments of the present invention, the accompanying drawings used in the description of the embodiments will be briefly introduced below. Obviously, the drawings described below are only some embodiments of the present invention. For those skilled in the art, other drawings can be obtained based on these drawings without creative effort.

[0022] Figure 1 This is a schematic diagram of the implementation steps of the method of the present invention.

[0023] Figure 2 This is a schematic diagram of the data waveforms used to solve the transient load gradient using the longitudinal dynamics model of this invention.

[0024] Figure 3 This is a schematic diagram of the closed-loop adjustment and forced zeroing control of clutch pressure during the slip-friction overlap stage of the present invention.

[0025] Figure 4 This is a schematic diagram of the system module connections of the present invention. Detailed Implementation

[0026] The technical solutions of the embodiments of the present invention will be clearly and completely described below with reference to the accompanying drawings. Obviously, the described embodiments are only some embodiments of the present invention, and not all embodiments. Based on the embodiments of the present invention, all other embodiments obtained by those skilled in the art without creative effort are within the scope of protection of the present invention.

[0027] Please see Figure 1As shown, the intelligent smooth shifting method for bulldozers based on an industrial mechanism model proposed in this invention includes: S1, when the shifting command is triggered, collecting the rodless chamber pressure of the bulldozer blade lifting cylinder, the turbine speed of the hydraulic torque converter, and the speed of the track drive wheel, and substituting the collected pressure and speed into the longitudinal dynamic mechanism model to solve for the transient load gradient.

[0028] In a preferred embodiment, the step of solving the transient load gradient includes: converting the collected rodless chamber pressure into blade cutting resistance based on the static force balance relationship between the rodless chamber pressure of the bulldozer blade lifting cylinder and the blade cutting resistance; converting the driving wheel speed into the overall machine travel inertial resistance; inputting the hydraulic torque converter turbine speed into the transmission system dynamic equation that includes the torque converter characteristics, and combining the blade cutting resistance and the overall machine travel inertial resistance to calculate the transient load gradient in real time.

[0029] In a specific embodiment, the signal acquisition process is triggered the instant the transmission control unit receives the shift command. The acquired data includes three physical quantities: the rodless chamber pressure of the bulldozer blade lifting cylinder, the torque converter turbine speed, and the track drive wheel speed. The rodless chamber pressure of the bulldozer blade lifting cylinder is measured by a pressure sensor installed in the rodless chamber oil circuit, reflecting the magnitude of the hydraulic support force during blade operation. The torque converter turbine speed is obtained through a speed sensor at the turbine shaft end, indicating the rotational speed of the torque converter output. The track drive wheel speed is acquired through a speed sensor at the drive wheel hub, representing the actual rotational speed of the machine's driving end.

[0030] The collected pressure and rotational speed data are input in real time into the longitudinal dynamics mechanism model. The longitudinal dynamics mechanism model is a set of mathematical equations describing the force and motion relationship of the bulldozer along its travel direction. It correlates the external operating load, the overall machine inertial force, and the internal dynamic characteristics of the transmission system, ultimately solving for the transient load gradient. The transient load gradient characterizes the torque value of the overall machine's external load and rotational inertia acting on the turbine shaft end of the torque converter at the transient moment of gear shifting, before it is buffered by the transmission mechanism, and is measured in Newton-meters.

[0031] The solution process is achieved as follows: Step 1, based on the static force balance relationship between the pressure in the rodless chamber of the bulldozer blade lifting cylinder and the cutting resistance of the blade, the collected pressure in the rodless chamber is converted into the cutting resistance of the blade. The bulldozer blade lifting mechanism is a defined linkage system, which includes a lifting cylinder, a push arm, a brace, and the blade body. The cutting resistance of the blade is the working resistance exerted by the soil on the blade edge when the blade pushes the soil. This resistance is transmitted through the linkage and forms a torque balance with the hydraulic pressure in the rodless chamber of the lifting cylinder. Ignoring transient inertial effects and friction, this static force balance relationship can be expressed as:

[0032]

[0033] in This represents the cutting resistance of the scraper, measured in Newtons. The pressure in the rodless chamber of the bulldozer blade lifting cylinder is expressed in Pascals. This represents the effective working area of ​​the rodless cavity, in square meters. This area is determined by the cylinder diameter parameter, and the specific value is determined based on the cylinder assembly drawing. The mechanical gain proportionality coefficient is dimensionless and reflects the lever multiplication relationship from the cylinder thrust to the tangential resistance of the blade. This coefficient is derived by performing spatial kinematic analysis on the three-dimensional digital model of the blade and linkage mechanism, and extracting the lever arm ratio of the cylinder thrust vector to the blade cutting force vector.

[0034] Step two: Convert the drive wheel speed into the overall machine's inertial resistance. Changes in the drive wheel speed directly reflect the overall machine's speed trend. Calculating the time derivative of the drive wheel speed yields the drive wheel angular acceleration. Then, combining this with the overall machine mass and drive wheel radius, the inertial force resisting the machine's speed change can be determined. The conversion relationship for the overall machine's inertial resistance is as follows:

[0035]

[0036] in Represents the overall inertial drag force of the machine, measured in Newtons; This represents the overall weight of the machine, expressed in kilograms. This parameter is taken from the bulldozer's factory technical specifications. The pitch circle radius represents the drive wheel of the track, in meters, and is determined based on the number of teeth on the drive wheel and the track pitch design value. This represents the rotational speed of the track drive wheel, measured in radians per second. Represents the angular acceleration of the drive wheel, which is calculated by the difference in drive wheel speeds within adjacent sampling periods and smoothed by low-pass filtering. The unit is radians per second squared.

[0037] Step 3: Input the hydraulic torque converter turbine speed into the transmission system dynamics equation, which includes the torque converter characteristics. Combine this with the previously obtained cutting resistance of the scraper and the overall machine's inertial resistance, and calculate the transient load gradient in real time. The hydraulic torque converter turbine speed not only determines the turbine's own inertial torque but also implicitly reflects the current speed ratio of the transmission chain. The transmission system dynamics equation represents the balance after deducting the torque required to accelerate the turbine shaft inertia from the external load torque at the turbine shaft end. The specific formula is:

[0038]

[0039] In the formula, This is the transient load gradient we are looking for, expressed in Newton-meters. Represents the gear ratio of the transmission in the current gear. It is dimensionless and its value is determined by the gear ratio of the currently engaged gear. It represents the total final transmission ratio formed by the central drive and the final drive. It is dimensionless, comes from the design parameters of the transmission system, and is equal to the product of the transmission ratio of the central drive bevel gear pair and the reduction ratio of the final drive planetary gear. The mechanical efficiency of the transmission chain from the gearbox to the drive wheel is dimensionless. The efficiency value under typical working conditions determined by bench tests is set as a constant value, taking into account gear meshing friction loss, bearing friction loss and oil churning loss. The moment of inertia representing the turbine shaft and its synchronously rotating inertial elements is equivalent to the turbine axis. It includes the sum of the moments of inertia of the turbine body, the gearbox input shaft, and the drive gear. The unit is kilogram-square meter. This value is calculated based on a three-dimensional solid model. This represents the turbine speed of the hydraulic torque converter, measured in radians per second. Representing the turbine angular acceleration, it is also obtained from the turbine speed difference and filtered to eliminate signal jitter, with units of radians per second squared. The numerator of the first term in the formula is the product of the sum of the blade cutting resistance and the overall machine's inertial resistance, multiplied by the drive wheel radius. Its physical meaning is to convert the external linear longitudinal force into torque on the drive wheel, and then divide by the overall transmission ratio and efficiency to obtain the external equivalent load torque directly acting on the turbine shaft end. The second term... This is to accelerate the inertial torque required for the rotating components of the turbine shaft system. After addition, the transient load gradient... It accurately provides the total load torque requirement value at the moment of gear shift, after the external load of the whole machine is completely transferred to the turbine shaft and the effect of rotational inertia is superimposed. This total requirement value will be used for precise pressure matching of the gear clutch.

[0040] An exemplary implementation scenario is as follows: When a certain type of hydraulic bulldozer is performing gear shifting control, at the moment the gear shifting command is triggered, the pressure in the rodless chamber of the bulldozer blade lifting cylinder is collected. The effective working area of ​​the rodless chamber of this type of bulldozer blade cylinder is 350,000 Pascals. Based on an aperture of 160 mm, the calculated value is 0.0201 square meters, and the mechanical gain proportionality coefficient is... The mechanism model is calibrated to 2.08. Substituting the static force equilibrium relationship, the cutting resistance of the shovel is... Newton. The track drive wheel speed collected at the same moment. The angular acceleration of the drive wheel is obtained by differential calculation with a control cycle of 1.05 radians per second and digital filtering with a cutoff frequency of 10 Hz after a 5-millisecond control cycle. The speed is 0.48 radians per second squared, and the overall machine weight is... Based on the nameplate parameters, the weight is 23,500 kg, and the drive wheel pitch circle radius is... Based on the design data of 0.42 meters, the overall inertial drag of the machine is... Newton. The gearbox is currently in second forward gear. The final transmission ratio is 3.2. The mechanical efficiency is 4.8. Taking 0.88, and converting the above force value to the turbine shaft end, the external equivalent load torque is: Newton-meters. The turbine speed of the hydraulic torque converter at the same instant. The turbine angular acceleration is 162.3 radians per second, obtained through differential calculation. The equivalent rotational inertia of the turbine shaft is 18.2 radians per second squared. The CAD software measurement shows a value of 0.62 kg / m², therefore the turbine acceleration inertial torque is... Newton-meters. The transient load gradient is finally calculated. Newton-meters. This value is used in subsequent steps as a reference to determine the load torque that the target gear clutch should deliver. Please refer to [link / reference]. Figure 2 As shown, Figure 2 This is a schematic diagram of the data waveforms used in the longitudinal dynamics model of this invention to calculate the transient load gradient. The diagram visually demonstrates the high-frequency dynamic response trend of the total transient load gradient during gear shift, which is formed by superimposing the equivalent cutting resistance, equivalent driving inertia, and turbine acceleration inertia calculated to the turbine shaft end.

[0041] S2. Based on the transient load gradient and the structural parameters of the target gear clutch, calculate the minimum engagement oil pressure of the target gear clutch according to the hydrostatic pressure balance relationship, and use it as the target engagement pressure reference value.

[0042] In a preferred embodiment, the step of calculating the minimum engagement oil pressure of the target gear clutch based on the hydrostatic pressure balance relationship as the target engagement pressure reference value includes: obtaining the effective piston working area, number of friction pair groups, average friction radius, and dynamic friction coefficient under the current operating conditions of the target gear clutch; converting the transient load gradient into the load torque that the clutch needs to transmit; and calculating the target engagement pressure reference value based on the effective piston working area, number of friction pair groups, average friction radius, dynamic friction coefficient, and load torque, according to the balance equation of hydrostatic thrust and friction torque.

[0043] In a further preferred embodiment, after calculating the target engagement pressure reference value, the method further includes: performing temperature correction on the dynamic friction coefficient based on the hydraulic oil temperature to obtain a corrected dynamic friction coefficient; and updating the target engagement pressure reference value based on the corrected dynamic friction coefficient.

[0044] In a specific embodiment, after the transient load gradient is calculated in step S1, the transmission control unit then calculates the target engagement pressure reference value based on this gradient and the internal structural parameters of the target gear clutch through the hydrostatic pressure balance relationship. The target engagement pressure reference value refers to the minimum hydraulic oil pressure required for the target gear clutch to completely eliminate slippage and fully transmit the load; it serves as the final target value for subsequent slip-friction overlap and closed-loop adjustment. The entire calculation process is divided into four stages: parameter acquisition, load torque conversion, initial calculation, and temperature correction, with each stage closely linked.

[0045] The transmission control unit first reads the effective piston contact area, number of friction pair groups, and average friction radius of the target gear clutch from its non-volatile memory. The effective piston contact area is the annular area of ​​the clutch piston under oil pressure after being filled with oil, measured in square meters. This value is determined by the piston's outer and inner diameters and is precisely measured and written into the memory before clutch assembly. The number of friction pair groups refers to the number of friction surfaces formed by the steel plates and friction discs inside the clutch; it is dimensionless and determined by the clutch structural design. The average friction radius is the average effective lever arm of the friction disc contact surface from the inner to the outer diameter, measured in meters. It is the arithmetic mean of the outer and inner radii of the friction discs to ensure equivalence in torque calculation.

[0046] In sync with the aforementioned geometric parameters, the transmission control unit acquires the hydraulic oil temperature in the transmission oil pan in real time. This temperature is obtained through an NTC thermistor sensor built into the transmission oil pan, and the initial dynamic friction coefficient under the current operating condition is retrieved based on a pre-stored dynamic friction coefficient and oil temperature characteristic curve. The dynamic friction coefficient is the ratio of frictional force to clamping force under relative sliding conditions, and is dimensionless. The characteristic curve is obtained by fitting multiple sets of measured data at different oil temperature points on a dedicated friction test bench for the clutch assembly, reflecting the influence of oil viscosity changes and the temperature sensitivity of friction materials on frictional force.

[0047] Next, the transient load gradient obtained in step S1 is converted into the load torque that the target gear clutch needs to transmit. For the target gear clutch preparing to engage during gear shifting, as a key node for power transmission, the relationship between the torque it needs to withstand after full engagement and the transient load gradient is determined by its topological position in the transmission system. In this embodiment, since the driving end of the clutch is directly rigidly connected to the hydraulic torque converter turbine shaft (i.e., the gear ratio between them is 1), the driven end transmits power to the rear axle through gear meshing. Therefore, the load torque that the clutch needs to transmit... Directly equal to transient load gradient The value is expressed in Newton-meters. If the target gear clutch is located on another shaft system, it is multiplied by the corresponding gear ratio for conversion.

[0048] After obtaining the load torque, the initial target engagement pressure reference value is calculated based on the balance equation of hydrostatic thrust and frictional torque. Hydrostatic thrust is the axial clamping force generated by hydraulic pressure acting on the piston, and frictional torque is the limiting torque that a multi-plate clutch can transmit under clamping force. When the two are balanced, keeping the clutch from slipping, the relationship is:

[0049]

[0050] In the formula This represents the initial calculated target engagement pressure baseline value, in Pascals. The number of friction pairs in the target gear clutch is dimensionless. The coefficient of dynamic friction of the clutch in the target gear under the current operating conditions is dimensionless. The initial value obtained here is from the oil temperature table. The average friction radius of the clutch in the target gear position, in meters; This represents the effective working area of ​​the piston in the target gear clutch, measured in square meters. The denominator of this formula is the product of the number of friction pairs, the coefficient of dynamic friction, the average friction radius, and the piston area, in cubic meters. The numerator, the load torque, is in Newton-meters. Dividing by this gives Pascals, consistent with hydraulic pressure units. Gear pump oil supply pressure is often measured in megapascals (MPA), which can be further calculated by dividing by... Conversion.

[0051] To compensate for the real-time impact of temperature changes on the dynamic friction coefficient, a temperature correction is performed immediately after the initial calculation. The transmission control unit recalls the hydraulic oil temperature again and reads the temperature correction factor function from memory. This function stores measured data in the form of a linear interpolation table, with the hydraulic oil temperature as the input. The unit is degrees Celsius, and the output is a dimensionless correction factor. Using the friction coefficient at 80 degrees Celsius as a baseline, the correction factor increases or decreases for every 1 degree Celsius deviation in temperature. Multiplying the correction factor by the initial kinetic friction coefficient yields the corrected kinetic friction coefficient. :

[0052]

[0053] The target engagement pressure reference value is updated based on the corrected dynamic friction coefficient and finally stored in the control variable:

[0054]

[0055] In the formula This is the target engagement pressure reference value after temperature correction, which is used as the switching threshold for pressure relief engagement in step S6.

[0056] An exemplary implementation scenario is as follows: Continuing from the example in step S1, due to the correct compensation for the inertial torque of the rotating component, the transient load gradient... Solved as Newton-meters. The target gear is third gear. The effective piston contact area of ​​the target gear clutch is read from the transmission control unit's memory. The area is 0.0085 square meters, and the number of friction pairs is... The value is 6, and the average friction radius is... The depth is 0.073 meters. The temperature of the hydraulic oil in the transmission oil pan at this time is... The initial dynamic friction coefficient at 82 degrees Celsius was obtained from the friction plate temperature characteristic curve. The value is 0.127. The initial target engagement pressure reference value is calculated according to the formula: Then, temperature correction is applied, and the temperature correction factor function outputs at 82 degrees Celsius. The corrected coefficient of kinetic friction is obtained. Substitute back into the formula: This pressure value is set as the target engagement pressure reference value for subsequent gear shifting processes. When the clutch pressure of the target gear reaches this value, it means that the clutch has the ability to transmit the current full load torque, and then the clutch of the current gear is triggered to return to zero. The entire calculation process strictly follows the physical balance between the clutch hydrostatic thrust and friction torque, and the friction coefficient is corrected for real-time oil temperature to ensure that the reference value adapts to changes in operating conditions.

[0057] S3. Collect the hydraulic oil temperature in the transmission oil pan, determine the real-time dynamic viscosity of the hydraulic oil based on the mapping relationship between oil temperature and dynamic viscosity, and calculate the actual volumetric flow rate of the shift ratio solenoid valve under unit control quantity by combining the main oil circuit pressure collection value of the transmission and the orifice throttling flow relationship.

[0058] In a preferred embodiment, calculating the actual volumetric flow rate of the shift ratio solenoid valve under a unit control quantity includes: substituting the hydraulic oil temperature into the empirical formula for viscosity-temperature characteristics to obtain the real-time dynamic viscosity; collecting the main oil circuit pressure of the transmission and using it as the pressure difference across the valve orifice of the shift ratio solenoid valve for the target gear clutch; and calculating the actual volumetric flow rate based on the equivalent throttling area, flow coefficient, real-time dynamic viscosity, and pressure difference across the valve orifice of the shift ratio solenoid valve, according to the orifice throttling flow relationship.

[0059] In a specific embodiment, after the temperature correction calculation of the target engagement pressure reference value is completed in step S2, the transmission control unit immediately initiates the process of step S3 to provide real-time and accurate volumetric flow rate parameters for subsequently determining the fast-charge oil time. Volumetric flow rate refers to the volume of hydraulic oil that the shift ratio solenoid valve can inject into the target gear clutch per unit time under full duty cycle drive. It directly determines how quickly the clutch piston eliminates the separation gap; therefore, the output accuracy of step S3 affects the smoothness and response speed of the shifting process.

[0060] The transmission control unit first acquires the hydraulic oil temperature from the transmission oil pan. This temperature is obtained by an NTC thermistor sensor installed inside the oil pan. The sensor converts the resistance value into a voltage signal, which is then sent to the transmission control unit via an analog-to-digital converter. The temperature value, expressed in degrees Celsius, is obtained through table lookup and interpolation. The transmission oil pan, located at the bottom of the transmission housing, stores the hydraulic oil, and its temperature represents the current thermal state of the hydraulic fluid in the entire hydraulic system.

[0061] After obtaining the hydraulic oil temperature, the transmission control unit calls upon the empirical viscosity-temperature characteristic formula pre-stored in the non-volatile memory area, substitutes the hydraulic oil temperature into the calculation, and obtains the real-time dynamic viscosity. Dynamic viscosity is a physical quantity that measures the magnitude of internal friction in hydraulic oil, characterizing the oil's resistance to flow deformation; its unit is Pascal-second. Viscosity-temperature characteristics refer to the physical law that hydraulic oil viscosity decreases as temperature increases. The empirical viscosity-temperature characteristic formula stored in the transmission control unit is derived from viscosity-temperature curve data provided by the oil supplier, and is written into the controller after parameter fitting in the form of a double logarithmic viscosity-temperature equation. This empirical formula is written as:

[0062]

[0063] In the formula 0.7 represents the kinematic viscosity of hydraulic oil at the current temperature, measured in square millimeters per second; 0.7 is an inherent empirical constant in the double logarithmic viscosity-temperature international standard equation, dimensionless, used to correct the nonlinear deviation of the oil in the extremely low viscosity region and ensure the fitting accuracy of the equation across the entire temperature range. The value represents the hydraulic oil temperature, which is the temperature of the hydraulic oil collected from the transmission oil pan, in degrees Celsius; 273.15 is a conversion constant from degrees Celsius to Kelvin temperature units, used to convert relative temperature into absolute thermodynamic temperature to satisfy the physical meaning of logarithmic operations in the equation. and These are dimensionless constant parameters obtained by fitting two viscosity test points for a specific grade of hydraulic oil. These two parameters are determined by the oil viscosity grade and are written into the transmission controller during factory calibration. The kinematic viscosity obtained by solving the equation... Multiply by hydraulic oil density That is, the real-time dynamic viscosity is obtained. :

[0064]

[0065] In the formula Represents real-time dynamic viscosity, measured in Pascal-seconds; This represents the density of the hydraulic oil, expressed in kilograms per cubic meter, and is calibrated to a constant value of 860 in the controller; multiplied by... It is square millimeters per second With kilograms per cubic meter After multiplication, the dimensions are standardized to make the result conform to Pascal seconds.

[0066] Meanwhile, the transmission control unit collects the transmission main oil circuit pressure. The transmission main oil circuit pressure refers to the common supply pressure from the main oil pump, regulated by the main pressure regulating valve, and supplied to each shift proportional solenoid valve. It is measured by a pressure sensor installed on the common main oil circuit line. The transmission controller directly uses the transmission main oil circuit pressure as the inlet pressure of the shift proportional solenoid valve, while the outlet pressure of the shift proportional solenoid valve is the pressure required to engage the clutch in the target gear. At the beginning of the fast-charging phase, the pressure inside the clutch chamber is approximately zero, thus allowing a pressure difference across the shift proportional solenoid valve. equal to the main oil circuit pressure of the transmission All units are Pascals.

[0067] After acquiring the dynamic parameters of dynamic viscosity and valve orifice pressure difference, the transmission control unit also reads two fixed characteristic parameters of the shift ratio solenoid valve from the storage area: equivalent throttling area and flow coefficient. The equivalent throttling area is the equivalent flow cross-sectional area of ​​the flow channel when the shift ratio solenoid valve orifice is fully open under full duty cycle command, measured in square meters. This parameter is calculated using the valve core geometry and calibrated through bench steady-state flow tests. The flow coefficient is a dimensionless correction factor characterizing the flow energy loss at the valve orifice. Also derived from calibration tests, it compensates for the contraction and friction effects caused by the actual flow channel shape deviating from the ideal orifice.

[0068] After all the above parameters are complete, the transmission control unit calculates the actual volumetric flow rate of the shift ratio solenoid valve under a unit control quantity based on the orifice throttling flow rate relationship. The orifice throttling flow rate relationship is derived based on Bernoulli's equation and describes the relationship between the volumetric flow rate of incompressible fluid flowing through a narrow valve orifice and the pressure difference, flow area, and fluid density. A dynamic viscosity correction term is also introduced to compensate for the laminar flow effect caused by the increased viscous force at low Reynolds numbers, ensuring the formula maintains accuracy over a wide temperature range. The formula for calculating the actual volumetric flow rate is expressed as follows:

[0069]

[0070] in the formula This refers to the actual volumetric flow rate of the shift proportional solenoid valve under a unit control quantity, expressed in cubic meters per second. The flow coefficient of the shift ratio solenoid valve is dimensionless. The equivalent throttling area of ​​the shift ratio solenoid valve is expressed in square meters. The pressure difference across the valve port of the shift ratio solenoid valve is expressed in Pascals. This represents the density of hydraulic oil, expressed in kilograms per cubic meter. Represents real-time dynamic viscosity, measured in Pascal-seconds; The viscosity correction characteristic length, in meters, characterizes the equivalent hydraulic characteristic size of the valve orifice. It is taken from the empirically fitted value in bench flow tests and is used to adjust the decay rate of viscosity effects. To reference the nominal flow rate, in cubic meters per second, the measured flow rate of the solenoid valve at full opening at the nominal oil temperature is taken as the normalization benchmark for the viscosity correction term. This is the viscosity correction factor. It is an exponential function with the natural constant e as the base. When the dynamic viscosity is very low or the pressure difference is large, the factor approaches 1. At this time, the flow rate is equivalent to the result of the ideal turbulent orifice formula. When the oil temperature is extremely low and the dynamic viscosity is very high, the factor is less than 1, which makes the actual flow rate reasonably reduced.

[0071] An exemplary implementation scenario is as follows: After calculating the temperature-corrected target engagement pressure reference value in step S2, the transmission controller performs the following specific calculations for step S3: The hydraulic oil temperature in the transmission oil pan is collected. The temperature was 82 degrees Celsius, and the main oil circuit pressure of the transmission was also collected. The pressure difference is 1.42 MPa, or 1,420,000 Pascals. Therefore, at the start of fast charging, the pressure difference across the valve of the shift ratio solenoid valve is... Take 1,420,000 Pascals. The oil viscosity-temperature fitting constant is stored in the controller. Take 22.49, Taking 3.68, substituting it into the empirical relationship for viscosity-temperature characteristics, we get: , Taking 5.8725, the right side of the equation is... Solve this problem. ,and then , to obtain kinematic viscosity Square millimeters per second. Hydraulic oil density. Set to 860 kg / m³, multiply by kinematic viscosity and convert dimensions to obtain real-time dynamic viscosity. Pascal second.

[0072] The equivalent throttling area of ​​the shift ratio solenoid valve is then read from the storage area. It is 3.14 square millimeters, that is square meters; flow coefficient It is 0.68; viscosity correction characteristic length 0.002 meters; reference nominal flow rate It is 2.4 liters per minute, which translates to cubic meters per second. Substituting into the orifice throttling flow rate formula, we first calculate the ideal turbulence term: , .

[0073] Next, calculate the exponential term of the viscosity correction factor: The viscosity correction factor is... This indicates that under 82°C hot oil conditions, the flow regime has fully developed into turbulence, and the viscous effect is negligible. The final actual volumetric flow rate... The flow rate is cubic meters per second, which translates to approximately 7.36 liters per minute. This actual volumetric flow rate is used as the denominator in the calculation of the fast-filling time in step S4, ensuring that the time parameter accurately reflects the filling rate under the current oil temperature and main oil circuit pressure.

[0074] S4. Obtain the clearance oil filling volume parameter of the target gear clutch, divide it by the actual volumetric flow rate to obtain the fast oil filling time under the current working condition, and output a full duty cycle control signal to the target gear clutch solenoid valve based on the fast oil filling time.

[0075] In a preferred embodiment, obtaining the clearance oil-filled volume parameter of the target gear clutch includes: identifying the target gear indicated by the shift command, and reading the pre-calibrated and stored clearance oil-filled volume parameter corresponding to the target gear clutch from the non-volatile storage area of ​​the transmission controller. The clearance oil-filled volume parameter is determined by the product of the clutch piston area and the designed separation gap of the friction plate.

[0076] In a further preferred embodiment, after step S4 is completed, the method further includes: real-time monitoring of the pressure of the target gear clutch; when the pressure exceeds the pressure characteristic value that represents the end of the piston stroke, it is determined that the clutch clearance has been eliminated, and then the fast charging stage ends and the slip-friction overlap control of step S5 is triggered.

[0077] In a specific embodiment, after calculating the actual volumetric flow rate of the shift ratio solenoid valve under unit control quantity in step S3, the transmission control unit immediately begins to determine the fast-charging oil time of the target gear clutch, providing precise timing control for quickly eliminating the internal clearance of the clutch. The fast-charging oil stage is the first hydraulic action stage of the shifting process, requiring the target gear clutch piston to move rapidly under the push of the engagement oil pressure, crossing the separation gap between the friction plates, so as to achieve mutual contact between the friction plates and the steel plates, and avoid mechanical impact between the piston and the friction plates.

[0078] The transmission control unit first identifies the target gear indicated by the shift command. The target gear is determined by the shift lever electrical signal or the automatic shift strategy. After parsing this gear code, the transmission control unit uses it as an address index to read the pre-calibrated and stored clearance oil-filling volume parameters corresponding to the target gear clutch from the transmission controller's non-volatile memory. The non-volatile memory refers to a memory that retains data even after power failure, such as EEPROM or ferroelectric memory. This area has been pre-calibrated with the clearance oil-filling volume parameters for each gear clutch during the transmission's factory calibration phase. The clearance oil-filling volume parameter refers to the volume of space within the clutch chamber that hydraulic oil needs to fill during the stroke from when the clutch piston of the target gear is fully reset to when the friction plates just make contact, measured in cubic meters. The clearance oil-filling volume parameter is determined by the product of the clutch piston area and the designed separation clearance of the friction plates. (Clutch piston area...) It refers to the complete circular area of ​​the side of the clutch piston that bears the hydraulic thrust, and this area is determined by the piston's outer diameter. The friction plate is designed with a separation clearance. This refers to the total clearance value calculated along the entire piston stroke direction when the clutch is in the disengaged state, multiplied by the number of friction pairs and the clearance on one side between adjacent steel plates and friction plates. The unit is meters, and this clearance is taken from the technical requirements dimensions of the clutch assembly drawings.

[0079] Subsequently, the transmission control unit uses the actual volumetric flow rate of the shift ratio solenoid valve under unit control quantity obtained in step S3 as the divisor. This determines the fast-charging oil time. Equal to the gap filling oil volume parameter Divide by actual volumetric flow rate The unit is seconds. Fast charging time refers to the time required for the solenoid valve to fully open and fill the space corresponding to the clearance charging volume parameter with hydraulic oil when charging the clutch.

[0080] After calculating the fast-charge oil time, the transmission control unit uses this fast-charge oil time as the timer load value and outputs a full duty cycle control signal to the target gear clutch solenoid valve. The full duty cycle control signal means that the transmission control unit outputs a continuous high-level command to the coil drive circuit of the shift ratio solenoid valve, with a 100% duty cycle. This controls the solenoid valve spool to remain fully open at its maximum stroke, allowing hydraulic oil to fill the target gear clutch chamber at full speed according to the actual volumetric flow rate determined in step S3. During this process, the clutch piston moves smoothly, eliminating the design separation gap of the friction plates. When the piston movement terminates, it contacts the friction plates without impact. This is achieved by precisely controlling the timing of the piston to end the full-open oil supply when it has completed its gap stroke.

[0081] After step S4 is completed, the transmission control unit does not directly end the fast-charging phase by relying on the timer. Instead, it immediately monitors the pressure of the target gear clutch in real time after the fast-charging oil time is reached. Pressure monitoring is performed by a pressure sensor installed in the target gear clutch oil circuit. The sensor output signal is continuously sampled and compared with the pressure characteristic value representing the end of the piston stroke. The pressure characteristic value is the pressure inflection point value in the oil circuit when the clutch piston completes clearance elimination and the friction plates just make contact, measured in Pascals. This characteristic value is derived from multiple oil-charging tests of the clutch assembly on a hydraulic test bench, recording the pressure reading corresponding to the moment the piston displacement sensor signal returns to zero, and storing multiple average values ​​in the transmission control unit. When the real-time pressure exceeds the pressure characteristic value, the transmission control unit determines that the clutch clearance has been eliminated, immediately ends the fast-charging phase, and switches the target gear clutch solenoid valve from full duty cycle control to the slip-overlap control in step S5. This pressure feedback determination mechanism compensates for the accumulated error of pure time-based open-loop control, ensuring accurate capture of the clearance elimination moment even when factors such as oil temperature and wear cause changes in the actual clearance.

[0082] An exemplary implementation scenario is as follows: the actual volumetric flow rate is calculated in step S3. for Based on the cubic meters per second, the transmission control unit executes the various calculations and controls in step S4. The target gear for the shift command is third gear. The control unit uses this gear code as an index to read the corresponding clutch clearance oil volume parameter from the non-volatile memory area. The piston area of ​​the forward third gear clutch recorded in the storage area. The area is 0.0085 square meters, and the piston diameter is 104 millimeters; the friction plate is designed with a separation gap. The value is 0.0006 meters, which is accumulated from the 0.1 mm clearance on each side of the 6 friction pairs. Multiplying the two values ​​yields the clearance oil-filled volume parameter: cubic meters. The volume of step S3... Substituting cubic meters per second into the fast-charging fuel time calculation formula: The timer load is 42 milliseconds. The transmission control unit sets the timer load value to 42 milliseconds and outputs a control signal with 100% duty cycle to the target gear clutch solenoid valve. The solenoid valve is fully open, and hydraulic oil is charged into the clutch at a rate of approximately 7.36 liters per minute. After the 42-millisecond countdown ends, the control unit continuously monitors the pressure of the target gear clutch. The pressure characteristic value of this type of clutch calibrated in bench testing is 0.18 MPa, or 180,000 Pascals. When the pressure sensor reading rises from zero and exceeds 180,000 Pascals for the first time, the control unit determines that the clutch clearance has been eliminated and the piston is in contact with the friction plate without impact. It then ends the fast charging phase and enters step S5, the slip-overlap control. If the actual clearance deviates slightly from the nominal value due to oil temperature deviation or mechanical wear, the pressure detection mechanism ensures accurate determination of the clearance elimination time and prevents piston impact or insufficient oil filling due to open-loop timing errors.

[0083] S5. After the fast oil charging time ends, the slip friction overlap stage is entered. The real-time discharge pressure of the current gear clutch solenoid valve and the real-time pressure build-up pressure of the target gear clutch solenoid valve are obtained. Based on Coulomb's law of friction and the characteristic parameters of the clutch friction pair, the residual transmission torque of the current gear clutch attenuation and the new transmission torque generated by the target gear clutch are calculated respectively.

[0084] In a preferred embodiment, calculating the residual transmission torque of the current gear clutch and the newly generated transmission torque of the target gear clutch includes: obtaining the number of friction pair groups, average friction radius, and friction coefficient of the current gear clutch; calculating the residual transmission torque according to the Coulomb friction torque formula based on the real-time release pressure of the current gear clutch solenoid valve, the piston working area of ​​the current gear clutch, the number of friction pair groups, average friction radius, and friction coefficient; specifically, calculating the newly generated transmission torque includes: obtaining the number of friction pair groups, average friction radius, and friction coefficient of the target gear clutch, and calculating the newly generated transmission torque based on the real-time pressure build-up pressure of the target gear clutch solenoid valve, the piston working area of ​​the target gear clutch, and the obtained number of friction pair groups, average friction radius, and friction coefficient.

[0085] In a specific embodiment, after determining in step S4 that the clutch clearance has been eliminated and the fast-charging phase has ended, the transmission control unit immediately switches the shift control from the fast-charging phase to the slip-overlap phase. The slip-overlap phase is the core transition phase of the entire shifting process. During this phase, the current gear clutch and the target gear clutch are in a synchronous separation and engagement dynamic process, and there is a speed difference between the friction plates of the two clutches, resulting in a slip friction state. During this phase, the transmission control unit needs to continuously acquire the pressure signals from the two clutch solenoid valves, calculate the residual transmitted torque and the newly added transmitted torque respectively, and provide real-time feedback for the closed-loop torque coordination in step S6.

[0086] The transmission control unit first acquires the real-time release pressure of the current gear clutch solenoid valve and the real-time build-up pressure of the target gear clutch solenoid valve. The real-time release pressure refers to the instantaneous pressure of the hydraulic oil in the clutch circuit of the gear about to disengage, reflecting the remaining clamping strength between the clutch friction plates. It is continuously collected by a pressure sensor installed in the clutch circuit, and the unit is Pascal. The real-time build-up pressure refers to the instantaneous pressure of the hydraulic oil in the clutch circuit of the target gear about to engage, reflecting the increasing clamping strength between the friction plates. It is also collected by a pressure sensor in the corresponding circuit, and the unit is Pascal. Both are updated synchronously at the analog-to-digital converter sampling period to ensure the time consistency of slip overlap control.

[0087] Subsequently, the transmission control unit reads three sets of fixed friction pair characteristic parameters for the current gear clutch and the target gear clutch from the non-volatile storage area: the number of friction pair groups, the average friction radius, and the friction coefficient. The number of friction pair groups refers to the number of friction surfaces formed by the steel plates and friction plates inside the clutch; it is dimensionless and has the same concept as the number of friction pair groups for the target gear clutch in step S2. The average friction radius refers to the effective lever arm that equates the friction torque of multiple friction pairs to a single point of concentrated force application; its unit is meters. It is the arithmetic mean of the outer and inner radii of the friction plates to ensure the equivalence of torque calculations on variable radius friction surfaces. This value is determined through integral calculation during the clutch design phase and written into the controller. The friction coefficient refers to the dynamic friction coefficient of the friction material under sliding friction conditions; it is dimensionless and is also affected by oil temperature. The transmission control unit obtains this coefficient in real-time from the stored temperature-friction coefficient compensation table based on the hydraulic oil temperature. Here, the hydraulic oil temperature can be reused from the transmission oil pan temperature collected in step S3.

[0088] To calculate the residual transmitted torque, the transmission control unit obtains the number of friction pairs, average friction radius, and friction coefficient of the clutch in the current gear position. Then, based on the real-time release pressure of the solenoid valve of the current gear clutch and the piston operating area of ​​the current gear clutch, it uses Coulomb's friction torque formula to calculate the residual torque that the clutch can still transmit during disengagement. Coulomb's law of friction states that sliding friction is proportional to the normal force, and the clamping force of the friction plates is determined by the product of the oil pressure and the piston operating area. To simplify the model, the slight deviation caused by the preload of the return spring is ignored. The multi-plate structure leads to a multiplication of torque. The formula for calculating the residual transmitted torque is:

[0089]

[0090] In the formula Represents the residual transmitted torque, measured in Newton-meters; The number of friction pairs in the clutch of the current gear is dimensionless. The friction coefficient of the clutch in the current gear is dimensionless. This represents the average friction radius of the clutch in the current gear, in meters. The real-time relief pressure of the clutch solenoid valve for the current gear is represented by Pascals. The piston working area of ​​the clutch in the current gear position refers to the effective annular area of ​​the clutch piston pressed against the friction plate by oil pressure, and the unit is square meters. This is consistent with the meaning of the effective piston working area in step S2.

[0091] For calculating the newly transmitted torque, the transmission control unit obtains the number of friction pairs, average friction radius, and friction coefficient of the target gear clutch. Based on the real-time pressure build-up of the target gear clutch solenoid valve and the piston action area of ​​the target gear clutch, it also uses the Coulomb friction torque formula to calculate the newly generated transmitted torque of the target gear clutch during engagement. The calculation formula is as follows:

[0092]

[0093] In the formula This represents the newly added transmitted torque, measured in Newton-meters. The number of friction pairs in the target gear clutch is dimensionless. The friction coefficient of the target gear clutch is dimensionless. This value has been temperature corrected and stored in step S2. The corrected dynamic friction coefficient can be reused here to maintain the consistency of control parameters. The average friction radius of the clutch in the target gear position, in meters; The real-time pressure build-up of the target gear clutch solenoid valve, in Pascals; The piston working area of ​​the clutch in the target gear is represented by square meters.

[0094] The real-time calculation of these two torque values ​​provides the torque contribution of the separation side and the engagement side for the closed-loop adjustment in step S6, and the drive control system can accurately sense the continuous transfer state of power transmission during the slip-friction overlap process.

[0095] An exemplary implementation scenario is as follows: After the 42-millisecond fast charging phase in step S4, pressure monitoring determines that the clutch pressure of the target gear exceeds 180,000 Pascals, the gap is eliminated, and the transmission control unit then enters the slip-overlap phase in step S5. At the 10-millisecond mark of entering this phase, the sensor collects the real-time release pressure of the current gear clutch solenoid valve. The real-time pressure build-up of the target gear clutch solenoid valve is 1,380,000 Pascals. The rated value is 230,000 Pascals. The transmission control unit reads the number of friction pairs for the current gear clutch (forward second gear) from the storage area. The average friction radius is 5. The value is 0.068 meters, and the coefficient of friction is... Based on the hydraulic oil temperature of 82 degrees Celsius, the value obtained from the table is 0.121, representing the piston working area. It is 0.0081 square meters. Substituting into the formula for residual transmitted torque: .

[0096] First, calculate the product of area and pressure: Newton. Then multiply by the friction radius and coefficient: , Finally, multiply by the number of friction pairs (5) to get... Newton's rice.

[0097] At the same time, read the number of friction pairs of the target gear clutch (forward third gear). The value is 6, and the average friction radius is... The value is 0.073 meters, and the coefficient of friction is... Directly use the temperature-corrected value of 0.125095 from step S2, the piston working area. The area is 0.0085 square meters. Substituting this into the newly added torque transmission formula: .

[0098] Product of pressure and area: Newton. Coefficient-radius product: , Multiply by 6 to get The residual transmitted torque is approximately 459.85 Nm, and the newly transmitted torque is approximately 107.1 Nm, totaling 566.95 Nm. This value is compared with the transient total torque output by the hydraulic torque converter turbine in step S6 to generate a deviation signal, driving the coordinated adjustment of the solenoid valve's duty cycle. During the continuous slip-friction overlap, as the real-time discharge pressure gradually decreases and the real-time build-up pressure gradually increases, the above torque value is continuously updated. Finally, when the clutch pressure of the target gear reaches the target engagement pressure reference value, the complete transfer of torque from the current gear to the target gear is completed.

[0099] S6. Using the transient total torque output by the hydraulic torque converter turbine as the constraint total, the current gear clutch solenoid valve and the target gear clutch solenoid valve are closed-loop regulated to ensure that at any time during the slip-overlap stage, the sum of the residual transmitted torque and the newly transmitted torque remains equal to the transient total torque. When the target gear clutch oil pressure rises to the target engagement pressure reference value, the current gear clutch pressure is forced to zero to complete the gear shift.

[0100] In a preferred embodiment, the closed-loop regulation of the current gear clutch solenoid valve and the target gear clutch solenoid valve includes: subtracting the sum of the residual transmitted torque and the newly transmitted torque from the transient total torque output by the torque converter turbine to obtain a torque deviation value; using the torque deviation value as the input of the PID controller to output the compensation slope of the solenoid valve control duty cycle; correcting the reciprocal of the pressure relief of the current gear clutch solenoid valve and the pressure build-up slope of the target gear clutch solenoid valve according to the compensation slope to ensure that the sum of the torques during the disengagement and engagement processes remains the transient total torque; and forcibly setting the control signal of the current gear clutch solenoid valve to zero when the real-time detected target gear clutch pressure reaches the target engagement pressure reference value.

[0101] In a specific embodiment, based on the continuous calculation of residual transmission torque and new transmission torque in step S5, step S6 performs closed-loop regulation of the entire process of slip-friction overlap stage. The transient total torque output by the hydraulic torque converter turbine is used as a unified constraint target to coordinate and control the duty cycle of the two clutch solenoid valves, ensuring seamless power transfer between disengagement and engagement, and forcibly cutting off the residual pressure of the old gear clutch after engagement.

[0102] The transmission control unit first needs to determine the transient total torque output by the torque converter turbine. The transient total torque output by the torque converter turbine refers to the torque value actually transmitted from the turbine shaft to the transmission input shaft at the current moment. It includes both the effective driving torque to overcome external loads and the inertial torque consumed by accelerating the turbine itself and the mass of the transmission components, and is measured in Newton-meters (Nm). This torque value can be obtained by combining the torque converter turbine speed and the characteristics of the torque converter pump impeller obtained in step S1. The transmission control unit calculates the torque converter speed ratio based on the ratio of the torque converter turbine speed to the pump impeller speed, and then retrieves the torque ratio coefficient and pump impeller torque coefficient from the original torque converter characteristic curve stored in the non-volatile memory area. The calculation method is based on the engine flywheel torque. With torque ratio coefficient The product of, where The data is broadcast in real time by the engine controller via the controller area network bus. During this process, the transient total torque output by the hydraulic torque converter turbine is recorded as... This value is recalculated and updated by the transmission control unit in each control cycle.

[0103] Next, the transmission control unit executes the core step of closed-loop adjustment. This torque deviation value... Equal to the transient total torque output of the hydraulic torque converter turbine Subtract residual transmitted torque With the addition of torque transmission The sum, measured in Newton-meters. When When positive, it indicates that the sum of the torques of the two clutches is lower than the turbocharger output torque. The remaining torque will be released as a sudden increase in engine speed or increased clutch slippage, requiring accelerated pressure build-up or reduced pressure release. When the value is negative, it indicates that the clutch torque has exceeded the turbine output, which will cause the transmission system to be over-constrained and generate shock, requiring accelerated depressurization or slowed pressure build-up.

[0104] The torque deviation value serves as the input to the PID controller. The PID controller is a closed-loop feedback algorithm that incorporates proportional, integral, and derivative operations. Based on the current amplitude, cumulative history, and trend of the deviation value, it comprehensively outputs the compensation slope of the solenoid valve control duty cycle. The compensation slope of the solenoid valve control duty cycle is an additional adjustment rate superimposed on the reciprocal of the pressure relief slope of the current gear clutch solenoid valve and the pressure build-up slope of the target gear clutch solenoid valve. It is dimensionless and changes the rate of clutch pressure rise and fall by correcting the duty cycle change rate of the PWM drive signal. The discrete output equation of the PID controller is:

[0105]

[0106] In the formula The compensation slope represents the duty cycle of the solenoid valve control, with the unit being the change in duty cycle per second. Its physical meaning is the percentage increase or decrease of the PWM duty cycle per second. The torque deviation value represents the current control cycle, expressed in Newton-meters. This represents the torque deviation value from the previous control cycle, expressed in Newton-meters. This represents the operation cycle of the PID controller, in seconds. It is set to 10 milliseconds, or 0.01 seconds, and is synchronized with the sampling update cycle of the pressure sensor. The proportional gain coefficient is expressed in units of duty cycle change per second per Newton-meter. This coefficient determines the strength of the compensation slope response to the instantaneous value of the deviation. Based on the calibration results of the step load response in the bench test, it is determined to be 0.00008. Represents the integral gain coefficient, with units of duty cycle change per second per Newton-meter-second. It is used to eliminate long-term small steady-state errors, and the limit value to avoid integral saturation was determined to be 0.0004 through repeated testing. The differential gain coefficient, measured in units of duty cycle change per second per Newton-meter per second, is used to predict deviation trends and suppress overshoot. Based on the system damping characteristics, it is set to 0.00001. All three coefficients were obtained through vehicle shift tests and are stored in the non-volatile memory of the transmission control unit.

[0107] After obtaining the compensation slope, the transmission control unit corrects the reciprocal of the pressure relief of the current gear clutch solenoid valve and the pressure build-up slope of the target gear clutch solenoid valve based on the compensation slope. The reciprocal of the pressure relief is the decrease in pressure per unit time during disengagement of the current gear clutch, measured in Pascals per second; a positive value indicates a pressure drop. The pressure build-up slope is the increase in pressure per unit time during engagement of the target gear clutch, measured in Pascals per second; a positive value indicates a pressure rise. The relationship before and after correction is as follows:

[0108]

[0109]

[0110] In the formula This represents the reciprocal of the pressure relief of the current gear clutch solenoid valve after correction, in Pascals per second. This represents the reciprocal of the basic pressure relief, measured in Pascals per second, and is determined by the slip-fist time planning, which requires the clutch to disengage within a preset overlap time. The pressure build-up slope of the target gear clutch solenoid valve after correction is expressed in Pascals per second. This represents the foundation slope, measured in Pascals per second, and is also determined by the slip-friction time planning. This is the pressure relief adjustment conversion factor, with the unit being Pascals per duty cycle change. It is used to convert the duty cycle compensation slope into the adjustment amount of the actual pressure relief rate. It is determined to be a constant value of 4200000 based on the slope of the pressure-duty cycle characteristic curve of the solenoid valve. This is the conversion factor for pressure build-up regulation, with the same unit of Pascals per duty cycle change, and the determination method is the same as... Consistent, take 3950000. When the torque deviation is positive, When the output is positive, the reciprocal of the pressure relief of the clutch in the current gear decreases, which temporarily slows down the pressure relief. The pressure build-up slope of the clutch in the target gear increases, which accelerates the pressure build-up. These two measures work together to increase the total torque and bring it close to the target gear's torque. When the torque deviation is negative, the effect is reversed.

[0111] Throughout the slip-overlap phase, the transmission control unit cyclically executes the aforementioned deviation calculation, PID solution, and slope correction closed-loop regulation every 10 milliseconds control cycle to ensure that the sum of the residual transmitted torque and the newly transmitted torque remains equal to the transient total torque output by the torque converter turbine at any given time. Simultaneously, before the end of each cycle, the transmission control unit checks whether the real-time detected target gear clutch pressure has reached the target engagement pressure reference value calculated and stored in step S2. The target gear clutch pressure is the value collected by the target gear clutch oil circuit pressure sensor, and the target engagement pressure reference value has been corrected in step S2 based on the current operating conditions and temperature. When this pressure condition is met, it indicates that the target gear clutch has the ability to independently transmit the full load torque. The transmission control unit forcibly sets the control signal of the current gear clutch solenoid valve to zero, that is, jumps the PWM duty cycle from the current value to zero, driving the current gear clutch to completely depressurize and disengage, completing the gear shift. This forced zeroing action is irreversible, marking the formal completion of power transfer; thereafter, only the target gear clutch maintains engagement by holding pressure.

[0112] An exemplary implementation scenario is as follows: at the 10th millisecond of the slip overlap phase in step S5, the transmission control unit calculates the residual transmitted torque within the current control cycle. The torque transmission capacity is 459.85 Nm, with an additional torque capacity. It is 107.1 Newton-meters. At the same moment, the transient total torque output of the hydraulic torque converter turbine is... The torque converter characteristic model yielded a value of 585 Nm after real-time calculation. The torque deviation was calculated using the formula: Newton-meters. The torque deviation value of 18.05 Newton-meters is input to the PID controller. The deviation value of the PID controller in the previous cycle. Assuming a value of 14.2 Newton-meters, the control period is... The proportional gain coefficient is 0.01 seconds. Take 0.00008 as the integral gain coefficient. Take 0.0004 as the differential gain coefficient. Take 0.00001. The result of the proportional term calculation is: The integral term calculation assumes the cumulative sum has reached 1.82 Nm / s, therefore... The differential term is calculated as follows: The sum of the three terms yields the compensation slope for the solenoid valve control duty cycle. The unit is the change in duty cycle per second.

[0113] Basic pressure relief reciprocal At this point in time, the planned speed is 85,000,000 Pascals per second, and the foundation slope is... The planned pressure relief rate is 62,000,000 Pascals per second. (Pressure relief adjustment conversion factor) The pressure build-up adjustment conversion factor is 4,200,000 Pascals per duty cycle change. This is a change of 3,950,000 Pascals per duty cycle. After correction, we get: Pascals per second, meaning the pressure relief rate decreased slightly by approximately 25292 Pascals per second, and the separation slowed down slightly; after correction The pressure build-up rate is increased slightly by approximately 23,787 Pascals per second, resulting in a slightly faster engagement. These two fine-tuning adjustments are performed simultaneously to ensure that the sum of the residual and newly generated transmission torques in the next control cycle converges to 585 Newton-meters.

[0114] Closed-loop adjustment continues every 0.01 seconds until the target gear clutch pressure sensor feedback value reaches the target engagement pressure reference value calculated in step S2. This reference value is approximately 1,316,582 Pascals in this shift. When this condition is met, the transmission control unit immediately reduces the duty cycle of the current gear clutch solenoid valve to zero. The current gear clutch oil circuit is fully connected to the oil tank via the shift ratio solenoid valve to achieve rapid pressure relief. The residual transmitted torque drops to zero in a very short time, and power transmission is entirely undertaken by the target gear clutch, completing the shift process. The entire S6 closed-loop adjustment process ensures that the sum of the torques of the two clutches during the slip overlap phase always matches the turbocharger output, eliminating torque interruption and shock during the shift process, while ensuring that the handover ends at the set time. Please refer to [link to relevant documentation]. Figure 3 As shown, Figure 3 This diagram illustrates the closed-loop adjustment and forced zeroing control of clutch pressure during the slip-friction overlap phase of this invention. By comparing it with the traditional open-loop fixed-slope pressure build-up / release trajectory, the significant advantage of this invention's closed-loop dynamic correction strategy—adaptively adjusting the pressure slope under real-time torque deviation drive and precisely triggering forced zeroing of the disengagement-side clutch based on the target engagement pressure reference value—is highlighted, thus achieving seamless power transfer.

[0115] Please see Figure 4 As shown, the second aspect of the present invention provides a bulldozer intelligent smooth shifting system based on an industrial mechanism model, comprising: a transient load calculation module, a target pressure calculation module, a volumetric flow rate calculation module, a fast charging timing control module, a slip torque calculation module, and a torque closed-loop adjustment module.

[0116] The transient load calculation module is used to collect the rodless chamber pressure of the bulldozer blade lifting cylinder, the turbine speed of the hydraulic torque converter, and the speed of the track drive wheel when the shift command is triggered. The collected pressure and speed are then substituted into the longitudinal dynamic mechanism model to solve for the transient load gradient.

[0117] The target pressure calculation module is used to calculate the minimum engagement oil pressure of the target gear clutch based on the transient load gradient and the structural parameters of the target gear clutch, according to the hydrostatic pressure balance relationship, and use it as the target engagement pressure reference value.

[0118] The volumetric flow rate calculation module is used to collect the hydraulic oil temperature in the transmission oil pan, determine the real-time dynamic viscosity of the hydraulic oil based on the mapping relationship between oil temperature and dynamic viscosity, and calculate the actual volumetric flow rate of the shift ratio solenoid valve under a unit control quantity by combining the main oil circuit pressure collection value of the transmission and the orifice throttling flow relationship.

[0119] The fast charging timing control module is used to obtain the gap oil filling volume parameter of the target gear clutch, divide it by the actual volume flow rate to obtain the fast charging time under the current working condition, and output a full duty cycle control signal to the target gear clutch solenoid valve with the fast charging time.

[0120] The slip torque calculation module is used to enter the slip overlap stage after the fast oil charging time ends, obtain the real-time discharge pressure of the current gear clutch solenoid valve and the real-time pressure build-up pressure of the target gear clutch solenoid valve, and calculate the residual transmission torque of the current gear clutch decay and the new transmission torque generated by the target gear clutch based on Coulomb's friction law and the friction pair characteristic parameters of the clutch.

[0121] The torque closed-loop adjustment module is used to use the transient total torque output by the hydraulic torque converter turbine as the constraint total to perform closed-loop adjustment on the current gear clutch solenoid valve and the target gear clutch solenoid valve, so that at any time during the slip-overlap stage, the sum of the residual transmitted torque and the newly transmitted torque remains equal to the transient total torque, and when the target gear clutch oil pressure rises to the target engagement pressure reference value, the current gear clutch pressure is forced to zero to complete the gear shift.

[0122] Each of the modules can be implemented in whole or in part through software, hardware, or a combination thereof. It supports hardware embedded in or independent of the processor in the computer device, and also supports software stored in the memory of the computer device, so that the processor can call and execute the operations corresponding to each of the above modules.

[0123] The above embodiments are only used to illustrate the technical solutions of the present invention, and are not intended to limit it. Although the present invention has been described in detail with reference to the foregoing embodiments, those skilled in the art should understand that modifications can still be made to the technical solutions described in the foregoing embodiments, or equivalent substitutions can be made to some of the technical features. Such modifications or substitutions do not cause the essence of the corresponding technical solutions to deviate from the spirit and scope of the technical solutions of the embodiments of the present invention, and should all be included within the protection scope of the present invention.

Claims

1. A method for intelligent and smooth gear shifting in bulldozers based on an industrial mechanism model, characterized in that, include: S1. When the shift command is triggered, the rodless chamber pressure of the bulldozer blade lifting cylinder, the turbine speed of the hydraulic torque converter, and the speed of the track drive wheel are collected. The collected pressure and speed are substituted into the longitudinal dynamic mechanism model to solve the transient load gradient. S2. Based on the transient load gradient and the structural parameters of the target gear clutch, calculate the minimum engagement oil pressure of the target gear clutch according to the hydrostatic pressure balance relationship, and use it as the target engagement pressure reference value. S3. Collect the hydraulic oil temperature in the transmission oil pan, determine the real-time dynamic viscosity of the hydraulic oil based on the mapping relationship between oil temperature and dynamic viscosity, and calculate the actual volumetric flow rate of the shift ratio solenoid valve under unit control quantity by combining the main oil circuit pressure collection value of the transmission and the orifice throttling flow relationship. S4. Obtain the clearance oil filling volume parameter of the target gear clutch and divide it by the actual volume flow rate to obtain the fast oil filling time under the current working condition. Use this fast oil filling time to output a full duty cycle control signal to the target gear clutch solenoid valve. S5. After the fast oil charging time ends, the slip friction overlap stage is entered. The real-time discharge pressure of the current gear clutch solenoid valve and the real-time pressure build-up pressure of the target gear clutch solenoid valve are obtained. Based on Coulomb's friction law and the friction pair characteristic parameters of the clutch, the residual transmission torque of the current gear clutch attenuation and the new transmission torque generated by the target gear clutch are calculated respectively. S6. Using the transient total torque output by the hydraulic torque converter turbine as the constraint total, the current gear clutch solenoid valve and the target gear clutch solenoid valve are closed-loop regulated to ensure that at any time during the slip-overlap stage, the sum of the residual transmitted torque and the newly transmitted torque remains equal to the transient total torque. When the target gear clutch oil pressure rises to the target engagement pressure reference value, the current gear clutch pressure is forced to zero to complete the gear shift.

2. The intelligent and smooth gear shifting method for bulldozers based on an industrial mechanism model according to claim 1, characterized in that, The process of solving for the transient load gradient includes: Based on the static force balance relationship between the rodless chamber pressure of the bulldozer blade lifting cylinder and the cutting resistance of the blade, the collected rodless chamber pressure is converted into the cutting resistance of the blade. The inertial resistance of the entire machine is calculated based on the speed of the drive wheels. The turbine speed of the hydraulic torque converter is input into the dynamic equation of the transmission system, which includes the characteristics of the torque converter. Combined with the cutting resistance of the shovel and the inertial resistance of the whole machine, the transient load gradient is calculated in real time.

3. The intelligent and smooth gear shifting method for bulldozers based on an industrial mechanism model according to claim 1, characterized in that, The calculation of the minimum engagement oil pressure of the target gear clutch based on the hydrostatic pressure balance relationship, as the target engagement pressure reference value, includes: Obtain the effective working area of ​​the piston of the target gear clutch, the number of friction pairs, the average friction radius, and the dynamic friction coefficient under the current operating conditions; Convert the transient load gradient into the load torque that the clutch needs to transmit; Based on the effective working area of ​​the piston, the number of friction pairs, the average friction radius, the dynamic friction coefficient, and the load torque, the target engagement pressure reference value is calculated according to the balance equation between static pressure thrust and friction torque.

4. The intelligent and smooth gear shifting method for bulldozers based on an industrial mechanism model according to claim 1, characterized in that, After calculating the target engagement pressure baseline value, the following is also included: The dynamic friction coefficient is corrected by temperature based on the hydraulic oil temperature to obtain the corrected dynamic friction coefficient. The target engagement pressure reference value is updated based on the corrected dynamic friction coefficient.

5. The intelligent and smooth gear shifting method for bulldozers based on an industrial mechanism model according to claim 1, characterized in that, The calculation of the actual volumetric flow rate of the shift ratio solenoid valve under a unit control quantity includes: Substitute the hydraulic oil temperature into the empirical relationship of viscosity-temperature characteristics to obtain the real-time dynamic viscosity. Collect the main oil circuit pressure of the transmission and use it as the pressure difference across the valve port of the shift ratio solenoid valve of the target gear clutch; Based on the equivalent throttling area, flow coefficient, real-time dynamic viscosity, and pressure difference across the valve orifice of the shift ratio solenoid valve, the actual volumetric flow rate is calculated according to the orifice throttling flow relationship.

6. The intelligent and smooth gear shifting method for bulldozers based on an industrial mechanism model according to claim 1, characterized in that, The step of obtaining the clearance oil-filled volume parameter of the target gear clutch includes: identifying the target gear indicated by the shift command, and reading the pre-calibrated and stored clearance oil-filled volume parameter corresponding to the target gear clutch from the non-volatile storage area of ​​the transmission controller. The clearance oil-filled volume parameter is determined by the product of the clutch piston area and the designed separation gap of the friction plate.

7. The intelligent and smooth gear shifting method for bulldozers based on an industrial mechanism model according to claim 1, characterized in that, After step S4 is completed, the process also includes: real-time monitoring of the pressure of the target gear clutch; when the pressure exceeds the pressure characteristic value that represents the end of the piston stroke, it is determined that the clutch clearance has been eliminated, and then the fast charging stage ends and the slip friction overlap control in step S5 is triggered.

8. The intelligent and smooth gear shifting method for bulldozers based on an industrial mechanism model according to claim 1, characterized in that, The calculation of the residual transmitted torque due to the decay of the clutch in the current gear and the additional transmitted torque generated by the clutch in the target gear includes: Obtain the number of friction pairs, average friction radius, and friction coefficient of the clutch in the current gear; The residual transmitted torque is calculated based on the real-time release pressure of the current gear clutch solenoid valve, the piston action area of ​​the current gear clutch, the number of friction pairs, the average friction radius, and the friction coefficient, according to the Coulomb friction torque formula. The calculation of the additional transmitted torque specifically includes: obtaining the number of friction pairs, average friction radius, and friction coefficient of the target gear clutch; and calculating the additional transmitted torque based on the real-time pressure build-up of the target gear clutch solenoid valve, the piston action area of ​​the target gear clutch, and the obtained number of friction pairs, average friction radius, and friction coefficient.

9. The intelligent and smooth gear shifting method for bulldozers based on an industrial mechanism model according to claim 1, characterized in that, The closed-loop adjustment of the current gear clutch solenoid valve and the target gear clutch solenoid valve includes: The torque deviation value is obtained by subtracting the sum of the residual transmitted torque and the newly added transmitted torque from the transient total torque output by the hydraulic torque converter turbine. The torque deviation value is used as the input of the PID controller, and the output is the compensation slope of the solenoid valve control duty cycle. The pressure relief reciprocal of the current gear clutch solenoid valve and the pressure build-up slope of the target gear clutch solenoid valve are corrected according to the compensation slope to ensure that the sum of the torques during the disengagement and engagement processes remains the transient total torque. When the real-time detected clutch pressure of the target gear reaches the target engagement pressure reference value, the control signal of the current gear clutch solenoid valve is forcibly set to zero.

10. A bulldozer intelligent smooth gear shifting system based on an industrial mechanism model, characterized in that, include: The transient load calculation module is used to collect the rodless chamber pressure of the bulldozer blade lifting cylinder, the turbine speed of the hydraulic torque converter, and the speed of the track drive wheel when the shift command is triggered. The collected pressure and speed are substituted into the longitudinal dynamic mechanism model to solve for the transient load gradient. The target pressure calculation module is used to calculate the minimum engagement oil pressure of the target gear clutch based on the transient load gradient and the structural parameters of the target gear clutch, according to the hydrostatic pressure balance relationship, and serve as the target engagement pressure reference value. The volumetric flow rate calculation module is used to collect the hydraulic oil temperature in the transmission oil pan, determine the real-time dynamic viscosity of the hydraulic oil based on the mapping relationship between oil temperature and dynamic viscosity, and calculate the actual volumetric flow rate of the shift ratio solenoid valve under a unit control quantity by combining the main oil circuit pressure collection value of the transmission and the orifice throttling flow relationship. The fast charging timing control module is used to obtain the gap oil filling volume parameter of the target gear clutch, divide it by the actual volume flow rate, obtain the fast charging time under the current working condition, and output a full duty cycle control signal to the target gear clutch solenoid valve with the fast charging time. The slip torque calculation module is used to enter the slip overlap stage after the fast oil charging time ends. It obtains the real-time discharge pressure of the current gear clutch solenoid valve and the real-time pressure build-up pressure of the target gear clutch solenoid valve. Based on Coulomb's friction law and the friction pair characteristic parameters of the clutch, it calculates the residual transmission torque of the current gear clutch attenuation and the new transmission torque generated by the target gear clutch, respectively. The torque closed-loop adjustment module is used to perform closed-loop adjustment of the current gear clutch solenoid valve and the target gear clutch solenoid valve, using the transient total torque output by the hydraulic torque converter turbine as the constraint total. This ensures that at any moment during the slip-overlap phase, the sum of the residual transmitted torque and the newly transmitted torque remains equal to the transient total torque. When the target gear clutch oil pressure rises to the target engagement pressure reference value, the current gear clutch pressure is forced to zero, completing the gear shift.

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