Gear shift control method for reducing gear shift shock

Abstract

The present invention relates to a shift control method for reducing the impact of gear shifting, comprising the steps of: setting a vehicle to enter a main gear shifting state; obtaining original information; the original information including the current gear position, the requested gear position, the intermediate shaft speed, the output shaft speed, the dog tooth speed difference, the transmission oil temperature, the air source pressure, the vehicle mass and the road slope; calculating the current maximum shifting force; searching and calling to obtain the gear-in speed difference window; the transmission speed control operation; determining that the vehicle enters the gear-in state; estimating the instantaneous dog tooth speed difference of the gear-in top gear, and looking up the table to obtain the optimal shifting force; adjusting the shift cylinder pressure; then executing the gear-in action; and ending the main gear-in process. The present invention achieves a breakthrough in the optimization design of the speed difference window from theory to practical calculation application; achieves a good match between the gear-in speed difference and the shifting force; and further reduces the gear-in shock and improves the adaptability of the shifting control on the basis of maintaining the rapidity of the gear shifting.

Description

Technical Field

[0001] The present invention relates to the technical field of AMT shift control, and in particular to a shift control method for reducing gear shifting impact. Background Art

[0002] In the field of commercial vehicle AMT, for the sake of cost reduction, efficiency improvement, and convenient assembly and maintenance, the synchronizer-free design of the transmission main box has been recognized and promoted by many professional manufacturers and OEMs; compared with synchronizer shifting, the technical advantages of synchronizer-free shifting are simple structure and fast shifting, and the disadvantage is that it is heavily dependent on the performance of the control system. In engineering practice, problems such as shifting stagnation, shifting shock, and even shifting failure frequently occur; strictly speaking, from the perspective of the whole vehicle, the synchronizer-free shifting shock includes: the shock caused by engine unloading / retorsion, the shock caused by clutch separation / engagement, and the shock of gear shifting. The method proposed in the present invention is only used to reduce the shock of gear shifting;

[0003] The execution process of the main gearbox without synchronizer shift is as follows: first, the neutral gear is disengaged, then the intermediate shaft is braked for speed regulation / free deceleration or the clutch is engaged for secondary clutching to speed up, and after the dog gear speed difference is adjusted to the gear window, the sliding gear sleeve is forced to mesh with the target gear engagement ring under the action of the shift force to complete the gear shift; there is a qualitative relationship between the speed difference window and the shift force: the larger the speed difference window, the greater the required shift force; when designing the shift control strategy, the optimal speed difference window under different working conditions and the shift force requirements during the shift execution process must be considered;

[0004] The existing technology generally needs to test on a test bench to obtain the gear speed difference window of different gears, and design the gear shift execution strategy based on this; the current shift force limit is based on comprehensive considerations such as the gear shift resistance, self-locking force, and the upper and lower limits of the dog gear speed difference. Compared with the shift force required for the gear shift action, there is a certain redundancy, but in order to shorten the gear shift time, the shift force is usually kept at the limit;

[0005] The defects of the prior art are analyzed as follows:

[0006] like Figure 1 As shown, the geometric structure of the dog teeth determines that when the top teeth are engaged, circumferential torque excitation will be generated under the action of the shift force, resulting in circumferential collision of the engaging teeth during the forced engagement process.

[0007] Obviously, if the dog teeth speed difference is too large, the top teeth will inevitably cause excessive circumferential torque excitation, aggravating the collision between teeth and the torsional vibration of the transmission chain, and in severe cases, it will cause abnormal noise of teeth rattling; in addition, it will also cause axial reaction force F′ gs Too large, if F′ gs >F s , the gear sleeve will be pushed back, causing axial impact.

[0008] On the other hand, if the speed of the sliding gear sleeve and the engaging gear ring is controlled to be synchronous or with a very small speed difference, the circumferential meshing impact of the tooth entry moment will be greatly reduced, but there are the following disadvantages:

[0009] (1) The speed regulation time and power interruption time are extended.

[0010] (2) The difficulty of speed control is increased. The requirements for speed control are too stringent and it is difficult to adapt to complex road conditions, especially heavy-load slope conditions. When the engine is unloaded, the output shaft speed fluctuates, and the input end is difficult to respond to the synchronization request quickly and accurately, resulting in severe gear shift shock.

[0011] (3) Too small speed difference causes tooth alignment. The correct meshing condition of the meshing teeth is that the phase difference between the gear sleeve and the gear ring is π. At present, the phase difference cannot be corrected in engineering practice, and the phase can only be adjusted by friction or collision while maintaining a certain speed difference. When the speed difference is too small, the circumferential torque excitation generated by the top teeth is too low, and the input end cannot rotate at least one compensation angle relative to the output end, resulting in tooth alignment.

[0012] In order to solve the above technical problems, a very small number of researchers have conducted research on them, but have not been able to provide effective solutions; typical research is as follows:

[0013] (1) Cui Wenxia et al. studied the meshing performance of the meshing sleeve and meshing gear ring of a hybrid bus equipped with an AMT without synchronizer in the “Study on the Meshing Performance of the Meshing Sleeve and Meshing Gear Ring of a Synchronizer-Free Transmission”. Through ADAMS simulation, they believed that the optimal shift speed difference when the motor is actively synchronized is 20 rpm to 30 rpm. However, they did not discuss the optimal shift speed difference under different gears and different loads, nor did they give a theoretical calculation method for the optimal speed difference.

[0014] (2) Song Qiang et al. established a dynamic model of the shifting process of an AMT without synchronizer in electric vehicles in the article “Study on the nonlinear dynamic characteristics of the shifting process of an AMT without synchronizer based on the sleeve position angle” and analyzed in detail the effect of the shifting force on the shifting time and the shifting shock. However, since the AMT of electric vehicles has no clutch, the movement of the engaging teeth in the speed regulation process and the shifting process is different from that of the AMT of commercial vehicles. Therefore, the technical solution proposed in this article has limited guiding significance for the shifting force control of the AMT of commercial vehicles.

[0015] In terms of engineering, the existing technology generally explores the gear speed differential window based on bench tests first, and then adaptively corrects the gear speed differential window according to different working conditions. The design of the gear speed differential limit value relies heavily on bench tests and empirical calibration values, lacking the support of dynamic theory. When downshifting, in order to shorten the secondary clutch speed regulation time, the lower limit of the window is designed to be low, which is easy to cause gear shift shock; when upshifting, in order to reduce gear shift noise and increase the success rate of one-time gear shifting, the upper limit of the speed differential window is designed to be small. The instability and inaccuracy of brake speed regulation can easily cause insufficient gear speed differential and prolong the gear matching time.

[0016] In terms of shift force control, in order to shift gears quickly, the shift solenoid valve is usually kept fully open, and the dog tooth speed difference changes each time the gear is engaged, and the shift force requirements during the gear engagement process also change: before the top tooth collision, the shift force should be reduced appropriately to reduce the collision force, and the shift force should be increased appropriately during the top tooth collision process to speed up the tooth misalignment and shorten the gear engagement time. It is difficult to meet the actual needs of gear engagement by adopting a fixed value control of the shift force, and the shift shock phenomenon caused by excessive shift force often occurs. However, due to the complexity of cylinder dynamics and the hysteresis characteristics of the air pressure regulation in the cylinder, it is difficult for the pneumatic actuator to quickly and accurately adjust the shift force during the gear engagement process. The existing technology only controls the cylinder pressure based on the duty cycle, which often has limited effect in engineering. Summary of the invention

[0017] In view of the above problems, the present invention provides a gear shift control method for reducing the gear shift shock, which aims to achieve a breakthrough in the optimization design of the speed difference window from theory to practical calculation application; to achieve a good match between the gear shift speed difference and the gear shift force; and to further reduce the gear shift shock and improve the adaptability of the gear shift control while maintaining the rapidity of the gear shift.

[0018] To solve the above problems, the technical solution provided by the present invention is:

[0019] A gear shift control method for reducing gear shifting impact comprises the following steps:

[0020] S100. Set the vehicle to enter the main gear shift state; then obtain the original information; the original information includes the current gear, the requested gear, the intermediate shaft speed, the output shaft speed, the dog gear speed difference, the transmission oil temperature, the air source pressure, the vehicle mass and the road slope;

[0021] S200. Calculate the current maximum shift force according to the gas source pressure;

[0022] S300. According to the maximum shift force and the original information, the gear speed difference window is searched and called; the gear speed difference window is obtained by manual calculation and preset in the controller;

[0023] S400. Perform transmission speed control operation according to the gear speed difference window;

[0024] S500. Determine whether the vehicle has entered a gear-engaging state, and then perform the following operations based on the determination result:

[0025] If the vehicle enters the gear-engaging state, execute S600;

[0026] If the vehicle is not in gear, return to and execute S400 again;

[0027] S600. Estimate the instantaneous speed difference of the dog teeth when the top gear is engaged, and obtain the optimal shifting force by looking up the table;

[0028] S700. Adjust the shift cylinder pressure according to the optimal shift force; then execute the shift action;

[0029] S800. End the main gear shifting process.

[0030] Preferably, the gear speed difference window is obtained by manual calculation according to the following steps:

[0031] S310. Establishing a collision force model according to the collision test; then calculating the forces on the sleeve and the ring during the collision of the engaging teeth according to the collision force model;

[0032] S320. Establishing a transmission torsion model; then calculating the response of the forced meshing process system according to the transmission torsion model;

[0033] S330. Based on the first constraint condition and the second constraint condition, respectively, an upper limit of the dog tooth speed difference under the first constraint condition and an upper limit of the dog tooth speed difference under the second constraint condition are calculated; then, between the upper limit of the dog tooth speed difference under the first constraint condition and the upper limit of the dog tooth speed difference under the second constraint condition, the smaller value of the upper limit of the dog tooth speed difference is taken as the upper limit of the gear shift speed difference window;

[0034] S340. Calculate the lower limit of the dog tooth speed difference based on the calibrated longest gear time; then use the lower limit of the dog tooth speed difference as the lower limit of the gear speed difference window;

[0035] S350. Combine the upper limit of the gear-in-speed difference window and the lower limit of the gear-in-speed difference window to obtain the gear-in-speed difference window.

[0036] Preferably, the upper limit of the dog tooth speed difference under the first constraint condition in S330 is expressed as follows:

[0037]

[0038]

[0039] Where: α is the equivalent chamfer of the engaging teeth; m slv is the mass of the gear sleeve; μ g is the sliding friction coefficient of the collision contact area; R g is the radius of the meshing circle of the engaging teeth; t fe is the idle travel time of the gear; v n is the normal relative velocity of the collision zone; and v n =g(F gn );Δt is the duration of the tooth collision process;Φ is the normal pressure of the contact area under the second constraint condition;nhigh1 is the upper limit of the speed difference under the first constraint; n high2 is the upper limit of the speed difference under the second constraint condition;

[0040] Among them: the first constraint condition is that the gear sleeve does not rebound axially when the gear is topped; the second constraint condition is that the impact force when the gear is topped is not greater than 10m / s 3 ;

[0041] The upper limit of the dog tooth speed difference under the second constraint condition is expressed as follows:

[0042]

[0043] Where: i0 is the rear axle speed ratio; R tire is the tire radius, in m;

[0044] The upper limit of the gear speed difference window in S340 is expressed as follows:

[0045]

[0046] Where: n high is the upper limit of the forward gear speed difference window.

[0047] Preferably, in S340, the lower limit of the dog tooth speed difference is calculated based on the calibrated longest gear-matching time, which is specifically expressed as follows:

[0048]

[0049] Where: s is the tooth thickness of the engaging teeth; t m is the longest tooth alignment time.

[0050] Preferably, S600 specifically includes the following steps:

[0051] S610. Establish the motion equation of the system during the gear-locking idle travel phase;

[0052] S620. According to the motion equation of the gear-engaging idle travel stage system established in S610, establish an estimation equation for the instantaneous speed difference of the top teeth and an estimation equation for the idle travel time;

[0053] S630. Obtain a two-dimensional table of collision axial reaction force, shift force, and dog gear speed difference; the two-dimensional table is obtained by theoretical calculation combined with engagement tooth collision test;

[0054] S640. Determine the optimal shifting force using the half-division method.

[0055] Preferably, S640 specifically includes the following steps:

[0056] S641. Create an estimation zone for the best shift force; the lower limit of the initial estimation zone is 1 / 2 of the upper limit of the initial estimation zone; the upper limit of the initial estimation zone is the current maximum shift force;

[0057] S642. Set the target shift force equal to the lower limit of the estimation area;

[0058] Then, according to the estimation equation for the instantaneous dog tooth speed difference of the top tooth and the estimation equation for the idle travel time established in S620, the instantaneous dog tooth speed difference and the idle travel time of the top tooth are calculated respectively;

[0059] Then, the reaction force of the axial component of the positive pressure in the contact area is calculated based on the instantaneous dog tooth speed difference and the idle travel time of the top tooth; then, the magnitude of the reaction force of the axial component of the positive pressure in the contact area is compared with the target shift force, and the following operations are performed based on the comparison result:

[0060] If the reaction force of the axial component of the contact zone normal pressure is less than or equal to the target shift force, assigning the value of the target shift force to the lower limit of the estimation zone;

[0061] If the reaction force of the axial component of the normal pressure in the contact area is greater than the target shift force, calculating an average value of a lower limit of the estimation area and an upper limit of the estimation area, and assigning the lower limit to the estimation area;

[0062] S643. Execute S642 in a loop; the number of loops is 5; then the loop ends, and the value of the target shift force is assigned to the optimal shift force.

[0063] Preferably, adjusting the shift cylinder pressure according to the optimal shift force in S700 specifically includes the following steps:

[0064] S710. Determine whether the shift force adjustment condition is met; then perform the following operations according to the determination result:

[0065] If the shift force adjustment condition is met, execute S720;

[0066] If the shift force adjustment condition is not met, execute S730 to S750;

[0067] S720. Execute the gear-engaging action; during the gear-engaging action, the forward solenoid valve of the shift cylinder is fully opened, and the reverse solenoid valve is closed;

[0068] S730. Calculate the opening time of the reverse solenoid valve of the shift cylinder;

[0069] S740. Compare the opening time of the shift cylinder reverse solenoid valve with the opening time threshold, and compare the shift position with the manually preset calibrated shift position; then perform the following operations based on the two comparison results:

[0070] If the opening time of the reverse solenoid valve of the shift cylinder is greater than the opening time threshold, and the shift position is greater than the calibrated shift position, the reverse solenoid valve is opened and timing is performed until the opening time of the reverse solenoid valve is greater than or equal to the opening time of the reverse solenoid valve of the shift cylinder, and then the reverse solenoid valve is closed;

[0071] S750. Determine whether the first gear shift is successful, and then perform the following operations based on the determination result:

[0072] If the one-time shift is successful, the shift solenoid valve will be switched to exhaust;

[0073] Otherwise, the opening time of the reverse electromagnetic valve of the shift cylinder is reduced by 10%, and the supplementary shifting action is performed.

[0074] Preferably, the opening time threshold is expressed as follows:

[0075]

[0076] Among them, t ve is the opening time threshold; λ is the gear shifting action at t fe The ratio of the shift force at the moment to the current maximum shift force;

[0077] The opening time of the reverse solenoid valve of the shift cylinder is expressed as follows:

[0078]

[0079] Among them, t pu It is the opening time of the reverse solenoid valve of the shift cylinder.

[0080] Preferably, the shift force adjustment condition in S710 is expressed as follows:

[0081] λF m -50≤F s ^

[0082] Preferably, the collision force model in S310 is based on the Hunt-Crossley nonlinear spring-damper model of Hertz contact theory, and is expressed as follows:

[0083]

[0084] Among them, F gn is the collision force of the engaging teeth; δ is the deformation of the contact area, in m; K nis the equivalent contact stiffness, in N / m 1.5 ; C n is the damping coefficient, in N / (m 2.5 / s), and C n The following conditions are met:

[0085] C n ∝(v n ,e)

[0086] Where: e is the coefficient of restitution; v n is the normal relative velocity of the collision zone, in m / s, and when the geometric structure and material properties of the collision body are determined, F gn =f(v n ).

[0087] Compared with the prior art, the present invention has the following advantages:

[0088] 1. Since the present invention adopts a complete set of calculation methods for the limit value of the gear-in speed difference window and the adaptive control method of the shift force, a breakthrough in the optimization design of the speed difference window from theory to practical calculation application is achieved;

[0089] 2. Since the present invention provides an adaptive optimization control method for the shift force control during the gear shifting process, the gear shifting speed difference and the shift force are well matched;

[0090] 3. Since the present invention provides an adaptive optimization control method for the shift force control during the gear shifting process, the gear shifting shock is further reduced and the adaptability of the gear shifting control is improved while maintaining the rapidity of the gear shifting. BRIEF DESCRIPTION OF THE DRAWINGS

[0091] Figure 1 It is a schematic diagram of the collision force analysis of the engaging teeth in the prior art;

[0092] Figure 2 It is a control flow diagram of a specific embodiment of the present invention;

[0093] Figure 3 The relative coordinate system of the gear sleeve and the engaging gear ring of the specific embodiment of the present invention;

[0094] Figure 4a A schematic diagram of a lower chamfer collision zone according to a specific embodiment of the present invention;

[0095] Figure 4b It is a schematic diagram of the upper chamfer collision interval of a specific embodiment of the present invention;

[0096] Figure 5 A schematic diagram of collision force analysis of a simplified geometric structure of a coupling tooth according to a specific embodiment of the present invention;

[0097] Figure 6This is a transmission system torsional model of a specific embodiment of the present invention.

[0098] in:

[0099] Figure 1 Middle: F s is the shift force, in N; f s is the friction force on the sleeve, in N; F gn is the positive pressure in the contact area, in N; F gτ With F gs F gn The tangential and axial components of the force are in N; f g is the sliding friction force in the contact area, in N; ω gear With ω slv are the speed of the engagement ring gear and the speed of the gear sleeve, in rad / s; T d is the drag torque equivalent to the target gear, in Nm;

[0100] Figure 3 to Figure 5 In the equation, s is the tooth thickness of the engaging teeth, in mm; e k is the meshing clearance, in mm; N is the number of engaging teeth, in pieces; α is the equivalent chamfer of the engaging teeth, in rad; v slv is the axial speed of the sleeve, in m / s; F' gn 、F' gτ 、F' gs and f' g F gn 、F gτ 、F gs and f g The reaction force acts on the gear sleeve;

[0101] Figure 6 In the equation, J1 is the comprehensive moment of inertia of the rotating components at the input end equivalent to the target gear, in kg*m 2 ; J2 is the comprehensive moment of inertia of all parts from the drive axle half shaft to the transmission main box sliding gear sleeve equivalent to the output shaft, the unit is kg*m 2 ; J3 is the equivalent moment of inertia of the tire, in kg*m 2 ; J4 is the equivalent moment of inertia of the vehicle body, in kg*m 2 θ gear ,θ slv ,θ tire ,θ vhclare the rotation angles of the coupling ring gear, sliding gear sleeve, tire and vehicle body, in rad; k1 and c1 are the comprehensive torsional stiffness of the input shaft and intermediate shaft equivalent to the coupling ring gear, in Nm / rad, and the damping coefficient, in Nms / rad; k2 and c2 are the comprehensive torsional stiffness of the half shaft and transmission shaft equivalent to the output shaft, in Nm / rad, and the damping coefficient, in Nms / rad; k3 and c3 are the equivalent torsional stiffness of the tire, in Nm / rad, and the damping coefficient, in Nms / rad; T f is the driving resistance torque, in Nm. DETAILED DESCRIPTION

[0102] The present invention is further explained below in conjunction with specific embodiments. It should be understood that these embodiments are only used to illustrate the present invention and are not used to limit the scope of the present invention. After reading the present invention, various equivalent forms of modifications to the present invention by those skilled in the art all fall within the scope defined by the claims attached to this application.

[0103] like Figure 2 As shown, a shift control method for reducing gear shifting impact comprises the following steps:

[0104] S100. Set the vehicle to enter the main gear shift state; then obtain the original information; the original information includes the current gear, the requested gear, the intermediate shaft speed, the output shaft speed, the dog gear speed difference, the transmission oil temperature, the air source pressure, the vehicle mass and the road slope.

[0105] S200. Calculate the current maximum shift force F according to the air source pressure m .

[0106] S300. According to the maximum shift force and the original information, the gear-in speed difference window is searched and called; the gear-in speed difference window is obtained by manual calculation and preset in the controller.

[0107] It should be noted that the gear-input speed difference window includes an upper limit and a lower limit; the upper and lower limits of the gear-input speed difference window are determined by theoretical calculation and combined with the engagement tooth collision test, and finally written into the controller.

[0108] In this specific embodiment, the gear-in speed difference window is obtained by manual calculation according to the following steps:

[0109] S310. Establish a collision force model based on the collision test; the collision force model is a mechanical model; then, based on the collision force model, calculate the forces on the gear sleeve and the gear ring during the collision process of the engaging teeth.

[0110] In this specific embodiment, the collision force model in S310 is based on the Hunt-Crossley nonlinear spring-damper model of the Hertz contact theory, which is expressed as follows:

[0111]

[0112] Among them, F gn is the collision force of the engaging teeth; δ is the deformation of the contact area, in m, which is measured by the collision test and then preset in the system; K n is the equivalent contact stiffness, in N / m 1.5 ; C n is the damping coefficient, in N / (m 2.5 / s), and C n The following conditions are met and expressed as formula (2):

[0113] C n ∝(v n ,e) (2)

[0114] Where: e is the coefficient of restitution; v n is the normal relative velocity of the collision zone, in m / s, and when the geometric structure and material properties of the collision body are determined, F gn =f(v n ).

[0115] It should be noted that the theoretical basis of S310 is that the top tooth phenomenon of the sliding gear sleeve and the engaging gear ring can be abstracted as a multi-body system collision model.

[0116] It should be noted that if Figure 3 As shown, since the ratio of the tooth thickness of the engaging tooth to the arc radius of the dog tooth end face is small, the arc is simplified to a secant segment for the convenience of calculation.

[0117] Before calculating the collision circumferential torque, considering that the circumferential torque excitation directions of the engaging teeth when colliding at the upper chamfer and the lower chamfer are opposite, the type of collision and the constraint conditions must be determined.

[0118] As shown in Figure 4, the analysis of the lower chamfer collision during the gear shifting process must satisfy Condition 1; Condition 1 is expressed as formula (3):

[0119]

[0120] The occurrence of upper chamfer collision must satisfy Condition 2; Condition 2 is expressed by formula (4):

[0121]

[0122] No top tooth collision must satisfy Condition 3; Condition 3 is expressed by formula (5):

[0123]

[0124] In equations (3), (4) and (5): p is the shift position, in mm; p a is the end face position of the dog tooth of the engagement ring gear, in mm; Δθ is the relative rotation angle between the engagement ring gear and the gear sleeve, in rad; t fe is the top tooth moment, in seconds; R g is the radius of the meshing circle of the coupling teeth, in mm; θ0 is the initial phase, in rad; the value range of m is m∈[0,1,2,3,...).

[0125] The relationship between p and Δθ is expressed as follows:

[0126]

[0127] Normal relative velocity v before collision n According to formula (7):

[0128]

[0129] Where: m slv is the mass of the gear sleeve, in kg; Drt(Δθ) is the sign function, defined as follows:

[0130]

[0131] Normal relative velocity v during collision n According to formula (9):

[0132]

[0133] Where: Δt is the duration of the collision process, in seconds, measured by bench test.

[0134] like Figure 5 As shown, Figure 5 Analysis shows that when the tooth is top-toothed, the engaging tooth is excited by the circumferential torque T gτ According to formula (10):

[0135] T gτ = Drt(Δθ)R g (f g sinα-F gn cosα) (10)

[0136] The gear sleeve is subjected to axial reaction force F′ gs According to formula (11):

[0137] F′ gs =F gn sinα (11)

[0138] Coupling tooth collision friction fg According to formula (12):

[0139] f g =μ g F gn (12)

[0140] Where: μ g is the sliding friction coefficient in the collision contact area.

[0141] S320. Establish a transmission system torsion model; then calculate the response of the forced meshing process system based on the transmission system torsion model.

[0142] In this specific embodiment, S320 specifically includes the following contents:

[0143] like Figure 6 As shown, in order to quantitatively analyze the influence of meshing tooth collision on the transmission system, the concentrated mass method is used to establish the linear torsion model of the transmission system during the forced meshing process of the meshing teeth.

[0144] The motion equation of the meshing tooth system is expressed as follows:

[0145]

[0146] Get state variables Then the state space expression of the system is expressed as follows:

[0147]

[0148] in:

[0149]

[0150]

[0151]

[0152] S330. Calculate the upper limit n of the dog tooth speed difference under the first constraint condition based on the first constraint condition and the second constraint condition respectively. high1 , and the upper limit of the dog tooth speed difference under the second constraint condition; then between the upper limit of the dog tooth speed difference under the first constraint condition and the upper limit of the dog tooth speed difference under the second constraint condition, take the smaller upper limit value of the dog tooth speed difference as the upper limit of the gear speed difference window.

[0153] In this specific embodiment: According to formula (13), the first constraint condition is that the axial motion gear sleeve does not rebound, that is, For a given axial force F s , expressed as formula (15):

[0154]

[0155] Take F gn =f(v n ) is the inverse function of v n =g(F gn ), and the associated equations (7) and (9) can be used to obtain the upper limit n of the dog tooth speed difference under the first constraint condition of S330. high1 According to formulas (16), (17) and (18):

[0156]

[0157]

[0158]

[0159] Where: α is the equivalent chamfer of the engaging teeth; m slv is the mass of the gear sleeve; μ g is the sliding friction coefficient of the collision contact area; R g is the radius of the meshing circle of the engaging teeth; t fe is the idle travel time of the gear; v n is the normal relative velocity of the collision zone; and v n =g(F gn );Δt is the duration of the tooth collision process;Φ is the normal pressure of the contact area under the second constraint condition;n high1 is the upper limit of the speed difference under the first constraint; n high2 is the upper limit of the speed difference under the second constraint condition.

[0160] Further simplification can be done by taking the upper limit n of the dog tooth speed difference under the first constraint as high1 According to formula (19):

[0161]

[0162] Among them: the first constraint condition is that the gear sleeve does not rebound axially when the gear is topped; the second constraint condition is that the impact force is not greater than 10m / s when the gear is topped 3 .

[0163] Then, the upper limit of the gear speed difference n is calculated based on the second constraint condition: high2 :

[0164] During the forced meshing process of the meshing teeth, in addition to the impact caused by the axial action, the impact caused by the circumferential action cannot be ignored. The impact degree is used as the second constraint condition for the upper limit of the dog tooth speed difference, that is, the impact degree j≤10m / s 3 ,, expressed according to formula (20):

[0165]

[0166] In the above formula, i0 is the rear axle speed ratio; R tire is the tire radius, in meters.

[0167] F gτ Assuming that the solution of the state equation (14) is According to formula (21):

[0168] F gτ ≤q -1 (∫F(t)dt) (21)

[0169] Relating equations (7), (9), (10) and (12), we can express it as follows:

[0170]

[0171] Further simplification can be made by taking the upper limit of the gear speed difference n based on the second constraint as high2 According to formula (23):

[0172]

[0173] The upper limit of the speed difference of the forward gear is the smaller value of the upper limit of the speed difference under the two constraints, that is: n high =min(n high1 ,n high2 ).

[0174] S340. Calculate the lower limit of the dog tooth speed difference based on the calibrated longest gear-matching time; and then use the lower limit of the dog tooth speed difference as the lower limit of the gear-advancing speed difference window.

[0175] In this specific embodiment: In S340, the lower limit of the dog tooth speed difference is calculated based on the calibrated maximum tooth alignment time; specifically, consider an extreme top tooth working condition: the gear sleeve and the engaging gear ring teeth are aligned. If the speed difference is extremely small at this time, the collision cannot generate sufficient circumferential torque excitation, then the tooth alignment time will be extended until the two ends are meshed under system disturbance. To avoid the above working condition, the calibrated maximum tooth alignment time is t m , then the lower limit of the speed difference n low According to formula (24):

[0176]

[0177] Where: s is the tooth thickness of the engaging teeth; t m The longest gear matching time.

[0178] It should be noted that, based on the above calculations, combined with information such as air pressure, mass, slope, etc., a table of gear speed difference limits for different gears, mass grades and slope grades of the transmission can be given; this table can be preset in the system and can be called up at any time.

[0179] S350. Combine the upper limit of the gear-in speed difference window and the lower limit of the gear-in speed difference window to obtain the gear-in speed difference window.

[0180] S400. Perform transmission speed control operation according to the gear-in speed difference window.

[0181] S500. Determine whether the vehicle has entered a gear-engaging state, and then perform the following operations based on the determination result:

[0182] If the vehicle enters the gear-engaging state, S600 is executed.

[0183] If the vehicle is not in gear, the process returns to and executes S400 again.

[0184] S600. Estimate the instantaneous dog tooth speed difference of the top gear when engaging the gear, and obtain the optimal shifting force by looking up the table.

[0185] It should be noted that the principle of S600 is briefly summarized as follows: the instantaneous dog tooth speed difference of the top gear when the gear is engaged is estimated based on the system motion equation, and then the optimal shift force F is selected by looking up the table. s ^. The following is a detailed description of the process and principle of S600:

[0186] In this specific embodiment: S600 specifically includes the following steps:

[0187] S610. Establish the motion equation of the system in the gear-engaging idle travel phase, expressed as formula (25):

[0188]

[0189] Among them: J in The moment of inertia of the transmission input equivalent to the target gear, in kg*m 2 ; J out The moment of inertia of the vehicle end equivalent to the sliding gear sleeve, in kg*m 2 ; T v T is the resistance torque equivalent to the output shaft from the driving resistance, in Nm; re It is the residual braking torque generated by the exhaust of the intermediate shaft brake, in Nm, and is 0 when downshifting.

[0190] It should be noted that the drag torque T d Related to the transmission oil temperature, the residual braking torque T re It is related to the braking capacity coefficient, braking time and brake release time of the intermediate shaft brake, both of which can be obtained through bench tests and empirical formulas between the influencing factors.

[0191] S620. According to the motion equation of the gear-engaging idle travel stage system established in S610, establish an estimation equation for the instantaneous speed difference of the top gear and an estimation equation for the idle travel time; specifically:

[0192] In the idle stroke stage, the output shaft speed change and system disturbance are ignored. The dog tooth speed difference at the end of speed regulation is assumed to be n1, and the idle stroke time is t fe , then the instantaneous dog tooth speed difference n2 of the top tooth is expressed by formula (26):

[0193]

[0194] Idle travel time t fe According to formula (27):

[0195]

[0196] S630. Obtain collision axial reaction force F′ gs and shift force F s 、Dog teeth speed difference n d The two-dimensional table F′ gs (F s ,n d ); the two-dimensional table F′ gs (F s ,n d ) is given based on theoretical calculations combined with collision tests.

[0197] S640. Determine the optimal shifting force using the half-point method.

[0198] In this specific embodiment: S640 specifically includes the following steps:

[0199] S641. Create an estimation zone for optimal shifting force [F bestL ,F bestH ]; the lower limit of the initial estimation zone is 1 / 2 of the upper limit of the initial estimation zone; the upper limit of the initial estimation zone is the current maximum shift force.

[0200] In this specific embodiment, the initial estimation area is expressed as: [0.5F m ,F m ].

[0201] S642. Set the target shift force equal to the lower limit of the estimation zone, that is, F s ^=F bestL .

[0202] Then, according to the estimation equation of the instantaneous dog tooth speed difference of the top tooth and the estimation equation of the idle travel time established in S620, the instantaneous dog tooth speed difference and the idle travel time of the top tooth are calculated respectively, that is, t fe and n2.

[0203] Then, according to the calculated instantaneous dog tooth speed difference and idle travel time of the top tooth, the reaction force of the axial component of the positive pressure in the contact area is calculated; then, the magnitude of the reaction force of the axial component of the positive pressure in the contact area is compared with the target shift force, and the following operations are performed according to the comparison result:

[0204] If the reaction force of the axial component of the contact zone normal pressure is less than or equal to the target shift force, the value of the target shift force is assigned to the lower limit of the estimation zone.

[0205] If the reaction force of the axial component of the contact area normal pressure is greater than the target shift force, the average of the lower limit of the estimation area and the upper limit of the estimation area is calculated and assigned to the lower limit of the estimation area.

[0206] S643 loop execution S642; the number of loops is 5 times; then the loop ends, and the value of the target shift force is assigned to the optimal shift force F s ^.

[0207] It should be noted that S640 is expressed in mathematical form as follows:

[0208] Calculate t according to S620 fe and n2, if the current loop F s ^、The dog tooth speed difference n2 satisfies the condition F′ gs (F s ^,n2)≤F s ^, then the target shift force F s ^=F bestL ;otherwise S640 is repeated, and the upper limit of the number of cycles is 5 times.

[0209] S700. Adjust the shift cylinder pressure according to the optimal shift force; then execute the gear shifting action.

[0210] In this specific embodiment, adjusting the shift cylinder pressure according to the optimal shift force in S700 specifically includes the following steps:

[0211] S710. Determine whether the shift force adjustment condition is met; then perform the following operations according to the determination result:

[0212] If the shift force adjustment condition is met, it means that the shift force does not need to be adjusted, and S720 is executed.

[0213] If the shift force adjustment condition is not met, it means that the shift force needs to be adjusted, and then S730 to S750 are executed.

[0214] In this specific embodiment, the shift force adjustment condition in S710 is expressed as follows:

[0215] λF m-50≤F s ^ (28)

[0216] S720. Execute the gear shifting action; during the gear shifting action, the forward solenoid valve of the shift cylinder is fully opened, and the reverse solenoid valve is closed.

[0217] S730. Calculate the opening time of the reverse solenoid valve of the shift cylinder.

[0218] S740. Compare the opening time of the reverse solenoid valve of the shift cylinder with the opening time threshold, and compare the shift position with the manually preset calibrated shift position; then perform the following operations based on the two comparison results:

[0219] If the opening time of the reverse solenoid valve of the shift cylinder is greater than the opening time threshold, and the shift position is greater than the calibrated shift position, the reverse solenoid valve is opened and timing is started until the opening time of the reverse solenoid valve is greater than or equal to the opening time of the reverse solenoid valve of the shift cylinder, and then the reverse solenoid valve is closed.

[0220] In this specific embodiment, the opening time threshold is expressed as follows:

[0221]

[0222] Among them, t ve is the opening time threshold; λ is the gear shifting action at t fe The ratio of the shift force at a certain moment to the current maximum shift force.

[0223] The opening time of the reverse solenoid valve of the shift cylinder is expressed as follows:

[0224]

[0225] Among them, t pu It is the opening time of the reverse solenoid valve of the shift cylinder.

[0226] S750. Determine whether the first gear shift is successful, and then perform the following operations based on the determination result:

[0227] If the gear is entered successfully at one time, the shift solenoid valve will be switched to exhaust.

[0228] Otherwise, the opening time of the reverse solenoid valve of the shift cylinder is reduced by 10%, and the compensatory action is performed.

[0229] S800. End the main gear shifting process.

[0230] In the above detailed description, various features are grouped together in a single embodiment to simplify the disclosure. This method of disclosure should not be interpreted as reflecting an intention that the embodiments of the claimed subject matter require more features than are clearly stated in each claim. On the contrary, as reflected in the appended claims, the invention is in a state of having less than all the features of the disclosed individual embodiments. Therefore, the appended claims are hereby expressly incorporated into the detailed description, with each claim standing on its own as a separate preferred embodiment of the invention.

[0231] The disclosed embodiments are described above to enable any person skilled in the art to implement or use the present invention. Various modifications of these embodiments are obvious to those skilled in the art, and the general principles defined herein may also be applied to other embodiments without departing from the spirit and scope of the present disclosure. Therefore, the present disclosure is not limited to the embodiments given herein, but is consistent with the broadest scope of the principles and novel features disclosed in this application.

[0232] The above description includes examples of one or more embodiments. Of course, it is not possible to describe all possible combinations of components or methods for the purpose of describing the above embodiments, but it should be recognized by those skilled in the art that the various embodiments may be further combined and arranged. Therefore, the embodiments described herein are intended to cover all such changes, modifications and variations that fall within the scope of protection of the appended claims. In addition, with respect to the term "comprising" used in the specification or claims, the word is covered in a manner similar to the term "including", just as "including," is explained as a transitional word in the claims. In addition, any term "or" used in the specification of the claims is intended to mean "non-exclusive or".

[0233] The specific implementation methods described above further illustrate the objectives, technical solutions and beneficial effects of the present invention in detail. It should be understood that the above description is only a specific implementation method of the present invention and is not intended to limit the scope of protection of the present invention. Any modifications, equivalent substitutions, improvements, etc. made within the spirit and principles of the present invention should be included in the scope of protection of the present invention.

Claims

1. A gear shift control method for reducing gear shifting impact, characterized in that: The following steps are involved: S100. Set the vehicle to enter the main gear shift state; then obtain the original information; the original information includes the current gear, the requested gear, the intermediate shaft speed, the output shaft speed, the dog gear speed difference, the transmission oil temperature, the air source pressure, the vehicle mass and the road slope; S200. Calculate the current maximum shift force according to the gas source pressure; S300. According to the maximum shift force and the original information, the gear speed difference window is searched and called; the gear speed difference window is obtained by manual calculation and preset in the controller; S400. Perform transmission speed control operation according to the gear speed difference window; S500. Determine whether the vehicle has entered a gear-engaging state, and then perform the following operations based on the determination result: If the vehicle enters the gear-engaging state, execute S600; If the vehicle is not in gear, return to and execute S400 again; S600. Estimate the instantaneous speed difference of the dog teeth when the top gear is engaged, and obtain the optimal shifting force by looking up the table; S700. Adjust the shift cylinder pressure according to the optimal shift force; then execute the shift action; S800. End the main box gear shifting process; The gear speed difference window is obtained by manual calculation according to the following steps: S310. Establishing a collision force model according to the collision test; then calculating the forces on the sleeve and the ring during the collision of the engaging teeth according to the collision force model; S320. Establishing a transmission torsion model; then calculating the response of the forced meshing process system according to the transmission torsion model; S330. Based on the first constraint condition and the second constraint condition, respectively, an upper limit of the dog tooth speed difference under the first constraint condition and an upper limit of the dog tooth speed difference under the second constraint condition are calculated; then, between the upper limit of the dog tooth speed difference under the first constraint condition and the upper limit of the dog tooth speed difference under the second constraint condition, the smaller value of the upper limit of the dog tooth speed difference is taken as the upper limit of the gear shift speed difference window; S340. Calculate the lower limit of the dog tooth speed difference based on the calibrated longest gear time; then use the lower limit of the dog tooth speed difference as the lower limit of the gear speed difference window; S350. Combine the upper limit of the gear-in-speed difference window and the lower limit of the gear-in-speed difference window to obtain the gear-in-speed difference window.

2. The gear shift control method for reducing gear shifting shock according to claim 1, characterized in that: The upper limit of the dog tooth speed difference under the first constraint condition in S330 is expressed as follows: Where: α is the equivalent chamfer of the engaging teeth; m slv is the mass of the gear sleeve; μ g is the sliding friction coefficient of the collision contact area; R g is the radius of the meshing circle of the engaging teeth; t fe is the idle travel time of the gear shift; vn is the normal relative speed of the collision zone; and v n =g(F gn );Δt is the duration of the tooth collision process;Φ is the normal pressure of the contact area under the second constraint condition;n high1 is the upper limit of the speed difference under the first constraint; n high2 is the upper limit of the speed difference under the second constraint condition; Among them: the first constraint condition is that the gear sleeve does not rebound axially when the gear is topped; the second constraint condition is that the impact force when the gear is topped is not greater than 10m / s 3 ; The upper limit of the dog tooth speed difference under the second constraint condition is expressed as follows: Where: i0 is the rear axle speed ratio; R tire is the tire radius; The upper limit of the gear speed difference window in S340 is expressed as follows: Where: n high is the upper limit of the forward gear speed difference window.

3. The gear shift control method for reducing gear shifting shock according to claim 2, characterized in that: In S340, the lower limit of the dog tooth speed difference is calculated based on the calibrated longest gear matching time, which is specifically expressed as follows: Where: s is the tooth thickness of the engaging teeth; t m is the longest tooth-matching time.

4. The gear shift control method for reducing gear shifting shock according to claim 3, characterized in that: S600 specifically includes the following steps: S610. Establish the motion equation of the system during the gear-locking idle travel phase; S620. According to the motion equation of the gear-engaging idle travel stage system established in S610, establish an estimation equation for the instantaneous speed difference of the top teeth and an estimation equation for the idle travel time; S630. Obtain a two-dimensional table of collision axial reaction force, shift force, and dog gear speed difference; the two-dimensional table is obtained by theoretical calculation combined with engagement tooth collision test; S640. Determine the optimal shifting force using the half-division method.

5. The gear shift control method for reducing gear shifting shock according to claim 4, characterized in that: S640 specifically includes the following steps: S641. Create an estimation zone for the best shift force; the lower limit of the initial estimation zone is 1 / 2 of the upper limit of the initial estimation zone; the upper limit of the initial estimation zone is the current maximum shift force; S642. Set the target shift force equal to the lower limit of the estimation area; Then, according to the estimation equation for the instantaneous dog tooth speed difference of the top tooth and the estimation equation for the idle travel time established in S620, the instantaneous dog tooth speed difference and the idle travel time of the top tooth are calculated respectively; Then, the reaction force of the axial component of the positive pressure in the contact area is calculated based on the instantaneous dog tooth speed difference and the idle travel time of the top tooth; then, the magnitude of the reaction force of the axial component of the positive pressure in the contact area is compared with the target shift force, and the following operations are performed based on the comparison result: If the reaction force of the axial component of the contact zone normal pressure is less than or equal to the target shift force, assigning the value of the target shift force to the lower limit of the estimation zone; If the reaction force of the axial component of the normal pressure in the contact area is greater than the target shift force, calculating an average value of a lower limit of the estimation area and an upper limit of the estimation area, and assigning the lower limit to the estimation area; S643. Execute S642 in a loop; the number of loops is 5; then the loop ends, and the value of the target shift force is assigned to the optimal shift force.

6. The gear shift control method for reducing gear shifting shock according to claim 5, characterized in that: In S700, the shift cylinder pressure is adjusted according to the optimal shift force, which specifically includes the following steps: S710. Determine whether the shift force adjustment condition is met; then perform the following operations according to the determination result: If the shift force adjustment condition is met, execute S720; If the shift force adjustment condition is not met, execute S730 to S750; S720. Execute the gear-engaging action; during the gear-engaging action, the forward solenoid valve of the shift cylinder is fully opened, and the reverse solenoid valve is closed; S730. Calculate the opening time of the reverse solenoid valve of the shift cylinder; S740. Compare the opening time of the shift cylinder reverse solenoid valve with the opening time threshold, and compare the shift position with the manually preset calibrated shift position; then perform the following operations based on the two comparison results: If the opening time of the reverse solenoid valve of the shift cylinder is greater than the opening time threshold, and the shift position is greater than the calibrated shift position, the reverse solenoid valve is opened and timing is performed until the opening time of the reverse solenoid valve is greater than or equal to the opening time of the reverse solenoid valve of the shift cylinder, and then the reverse solenoid valve is closed; S750. Determine whether the first gear shift is successful, and then perform the following operations based on the determination result: If the one-time shift is successful, the shift solenoid valve will be switched to exhaust; Otherwise, the opening time of the reverse electromagnetic valve of the shift cylinder is reduced by 10%, and the supplementary shifting action is performed.

7. The gear shift control method for reducing gear shifting shock according to claim 6, characterized in that: The opening time threshold is expressed as follows: Among them, t ve is the opening time threshold; λ is the gear shifting action at t fe The ratio of the shift force at the moment to the current maximum shift force; The opening time of the reverse solenoid valve of the shift cylinder is expressed as follows: Among them, t pu It is the opening time of the reverse solenoid valve of the shift cylinder.

8. The gear shift control method for reducing gear shifting shock according to claim 7, characterized in that: The shift force adjustment condition in S710 is expressed as follows: λF m -50≤F s ^ 。 9. The gear shift control method for reducing gear shifting shock according to claim 8, characterized in that: The collision force model in S310 is based on the Hunt-Crossley nonlinear spring-damper model of Hertz contact theory and is expressed as follows: Among them, F gn is the collision force of the engaging teeth; δ is the deformation of the contact area; K n is the equivalent contact stiffness; C n is the damping coefficient, and C n The following conditions are met: C n ∞(v n ,e) Where: e is the coefficient of restitution; v n is the normal relative velocity of the collision zone, and when the geometric structure and material properties of the collision body are determined, F gn =f(v n ).

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