Dragging identification method and system for disc brake of commercial vehicle under dynamic working condition

Abstract

The invention discloses a commercial vehicle disc brake dragging identification method and system under a dynamic working condition, and the method comprises the steps: S1, collecting the fault feedback information of a user disc brake, and classifying the causes of braking dragging; s2, on the basis of the classification result of the brake dragging causes, constructing a brake dragging mechanism dynamic model; s3, based on the user operation data, typical braking working conditions are extracted; and S4, based on the brake dragging mechanism kinetic model and the typical brake working condition, completing identification of the working condition where brake dragging easily occurs. According to the method, the fault feedback information of the user disc brake is collected, and the brake dragging causes are classified, so that the diagnosis efficiency and accuracy of the brake dragging problem are effectively improved. And secondly, the established dynamic model of the brake dragging mechanism comprehensively considers multiple factors, so that the actual occurrence process of the brake dragging can be simulated more accurately, and a powerful tool is provided for deeply analyzing the brake dragging mechanism.

Description

Technical Field

[0001] The present invention relates to the field of brake drag identification, and in particular to a method and system for identifying commercial vehicle disc brake drag under dynamic conditions. Background Art

[0002] With the rapid development of urbanization and logistics, the market demand for light commercial vehicles continues to grow. As market competition intensifies, commercial vehicle companies pay more and more attention to product quality and reliability. Disc brakes are the mainstream brakes used in light commercial vehicles, and their reliability is of great significance to vehicle driving safety. Brake drag is one of the main types of disc brake failures, which will cause the vehicle to continue braking during driving, increase the vehicle's driving resistance, fuel consumption rate and driving fatigue, cause leakage in the brake caliper chamber, abnormal wear and howling of the brake disc, and then reduce the life of the brake, seriously threatening driving safety. Summary of the invention

[0003] In order to solve the technical problems in the above background, the present invention provides a method for identifying the drag of a commercial vehicle disc brake under dynamic conditions, the steps comprising:

[0004] S1. Collect user disc brake fault feedback information and classify the causes of brake drag;

[0005] S2. Based on the classification results of the causes of brake drag, a dynamic model of the brake drag mechanism is constructed;

[0006] S3. Extract typical braking conditions based on user operation data;

[0007] S4. Based on the brake drag mechanism dynamic model and the typical braking conditions, identify the conditions where brake drag is likely to occur.

[0008] Preferably, step S1 includes: collecting and analyzing user fault feedback data, and counting the main fault modes of the brake; determining the main cause of brake drag based on the fault description in the feedback information and the fault component analysis, and clarifying the main influencing factors in combination with the drag torque mathematical model.

[0009] Preferably, the constructed dynamic model of the brake drag mechanism includes: an elastohydrodynamic lubrication model considering the roughness of the contact surface; a rough surface contact pressure model; a sealing contact surface microelastic deformation model; a sealing contact surface oil film thickness model; and a sealing contact surface friction calculation model.

[0010] Preferably, the method for solving the dynamic model of the brake drag mechanism includes: establishing t-V and t-pcyl interpolation data according to the braking condition parameters, and performing numerical simulation on the contact surface; secondly, calculating the static contact pressure of the sealing contact surface and the initial oil film thickness, and substituting the calculation results into the Reynolds equation to obtain the oil film disturbance pressure; at the same time, using the GW contact model combined with the nominal oil film thickness to calculate the roughness peak contact pressure; then, calculating the fluid viscosity shear stress and the roughness peak contact shear stress between the sealing contact surfaces, substituting the shear stress into the microelastic deformation equation of the sealing contact surface, calculating the microdeformation of the sealing contact surface, and substituting the roughness contact pressure into the wear model to calculate the dynamic wear profile; finally, calculating the final actual oil film thickness based on all the calculation results.

[0011] Preferably, step S3 includes: extracting braking conditions from user feedback data, using K-means clustering method to obtain three types of typical braking conditions, and characteristic parameters include: braking speed, braking deceleration, brake pedal travel and brake pedal change rate.

[0012] Preferably, step S4 includes: the brake drag mechanism dynamic model takes the piston movement speed and the sealing pressure as input, and converts the typical braking condition into the model input parameters by establishing the AMESim model of the braking system, and uses the brake drag mechanism dynamic model to analyze the piston reset force and the sliding resistance. When the reset force is less than the sliding resistance, it is determined that brake drag is likely to occur under this condition.

[0013] The present invention also provides a commercial vehicle disc brake drag identification system under dynamic conditions, the system is used to implement the above method, including: a collection module, a construction module, an extraction module and an identification module;

[0014] The collection module is used to collect user disc brake fault feedback information and classify the causes of brake drag;

[0015] The construction module is used to construct a brake drag mechanism dynamics model based on the classification results of the brake drag causes;

[0016] The extraction module is used to extract typical braking conditions based on user operation data;

[0017] The identification module is used to identify the brake dragging condition that is prone to occur based on the brake dragging mechanism dynamics model and the typical brake condition.

[0018] Compared with the prior art, the present invention has the following beneficial effects:

[0019] The present invention can accurately locate the main causes of brake drag by collecting user disc brake fault feedback information and classifying the causes of brake drag, providing a reliable basis for subsequent model construction and working condition identification, thereby effectively improving the diagnostic efficiency and accuracy of brake drag problems. Secondly, the constructed brake drag mechanism dynamic model comprehensively considers multiple factors such as contact surface roughness, elastohydrodynamic lubrication, rough surface contact pressure, microelastic deformation, oil film thickness and friction calculation, and can more accurately simulate the actual occurrence process of brake drag, providing a powerful tool for in-depth analysis of the brake drag mechanism. Thirdly, based on user operation data, typical braking conditions are extracted, and the K-means clustering method is used to classify the braking conditions, which can clearly identify the types of conditions prone to brake drag, providing scientific guidance for the design optimization and fault prevention of commercial vehicle disc brakes. Finally, the commercial vehicle disc brake drag identification system under dynamic conditions provided by the present invention realizes the full process automation processing from fault information collection to drag condition identification, which not only improves work efficiency, but also reduces the errors caused by manual intervention, and has high practicality and promotion value. BRIEF DESCRIPTION OF THE DRAWINGS

[0020] In order to more clearly illustrate the technical solution of the present invention, the following briefly introduces the drawings required for use in the embodiments. Obviously, the drawings described below are only some embodiments of the present invention. For ordinary technicians in this field, other drawings can be obtained based on these drawings without paying creative labor.

[0021] Figure 1 A schematic diagram of a method flow of an embodiment of the present invention;

[0022] Figure 2 Schematic diagram of classification results of an embodiment of the present invention; wherein (a) represents the statistical result of the frequency of disc brake failure modes; (b) represents the statistical result of the classification of the causes of disc brake drag;

[0023] Figure 3 A mechanical model for brake drag formed in an embodiment of the present invention;

[0024] Figure 4 A schematic diagram of the model coupling relationship of an embodiment of the present invention;

[0025] Figure 5 A schematic diagram of a three-dimensional coordinate system of a rectangular sealing ring sealing structure according to an embodiment of the present invention;

[0026] Figure 6 A schematic diagram of the oil film thickness on the rough contact surface of an embodiment of the present invention;

[0027] Figure 7This is a schematic diagram of the contact between the rough surface of the sealing ring and the piston surface according to an embodiment of the present invention;

[0028] Figure 8 A schematic diagram of the force acting on the sealing ring body according to an embodiment of the present invention;

[0029] Fig. 9 A schematic diagram of the force on a rough surface according to an embodiment of the present invention;

[0030] Fig.10 A schematic diagram of a model solving process according to an embodiment of the present invention;

[0031] Fig.11 Schematic diagram of the time history curve of the friction force in the piston seal contact area under typical braking conditions of an embodiment of the present invention; wherein (a) is condition 1; (b) is condition 2; (c) is condition 3;

[0032] Fig.12 This is the typical working condition brake clearance time history of the embodiment of the present invention. DETAILED DESCRIPTION

[0033] The following will be combined with the drawings in the embodiments of the present invention to clearly and completely describe the technical solutions in the embodiments of the present invention. Obviously, the described embodiments are only part of the embodiments of the present invention, not all of the embodiments. Based on the embodiments of the present invention, all other embodiments obtained by ordinary technicians in this field without creative work are within the scope of protection of the present invention.

[0034] In order to make the above-mentioned objects, features and advantages of the present invention more obvious and easy to understand, the present invention is further described in detail below with reference to the accompanying drawings and specific embodiments.

[0035] Embodiment 1

[0036] like Figure 1 FIG. 1 is a schematic diagram of the method flow of this embodiment, and the steps include:

[0037] S1. Collect user disc brake fault feedback information and classify the causes of brake drag.

[0038] For a certain type of commercial vehicle disc brake, the user fault feedback data is analyzed and the main failure modes of the brake are counted. According to the fault description in the feedback information and the faulty parts analysis, the main cause of brake drag is determined, and the main influencing factors are clarified in combination with the mathematical model of the drag torque.

[0039] Based on the fault information reported by users, the causes of brake drag are classified, such as Figure 2 As shown. Figure 2(a) It can be seen that brake drag is the main failure mode of the brake, accounting for about 58%; among the six reasons that lead to brake drag, floating guide failure and brake piston return failure account for 20.24% and 73.81% respectively. Figure 2 (b) As shown. Obviously, insufficient return ability of the brake piston is the main cause of brake drag. Further, according to the fault description information, it is confirmed that the secondary factors causing poor piston return are mainly the piston return amount and piston sliding resistance.

[0040] Then the factors affecting brake drag are analyzed. The return ability of the brake piston is related to the static and dynamic sealing mechanism of the rectangular seal ring in the brake caliper. After the seal ring is installed on the piston, it will be in a pre-compressed state and bear the brake fluid pressure during operation. When the brake fluid pressure is always less than the static friction force generated by the pressure on the pre-compressed seal contact surface of the seal ring, the static seal is completed. In the actual working process, there is a brake fluid film with a thickness of microns between the contact surfaces of the seal ring and the piston. When the piston is stationary, the brake fluid film is kept in a certain position under the action of surface tension, thereby achieving dynamic sealing.

[0041] Brake drag occurs during the piston stroke after the brake is released. The piston returns to its original position by relying on the energy of the seal ring's compression deformation, and the reset force is equal to the maximum static friction force of the static seal; the brake fluid film in the dynamic seal will affect the lubrication and friction wear state of the seal contact area, thereby affecting the piston sliding resistance. If the reset force is less than the sliding resistance, residual resistance will be generated between the friction plate and the brake disc, thus forming a drag torque.

[0042] The drag torque can be expressed as the residual braking torque of the friction plate and the brake disc, as shown in the following formula:

[0043] T 拖滞 =F μ ×R e =μ·F s ×R e (1)

[0044] Where, T 拖滞 Indicates the drag torque; F μ Represents the residual friction of the friction plate; R e represents the effective braking radius; μ represents the friction coefficient of the friction plate; F s Indicates the positive pressure between the friction pad and the brake disc.

[0045] Generally, the effective radius of the brake is a fixed value, so the drag torque is mainly related to the positive pressure F between the friction pad and the brake disc. s Therefore, a mechanical model of brake drag formation is established to analyze Fs, such as Figure 3 shown.

[0046] In this model, the spring stiffness is used to represent the stiffness of each component, where K t K is the clamp body stiffness; c K is the sealing ring stiffness; p is the friction plate stiffness; f s is the sliding resistance of the friction plate in the slideway; f x is the sliding resistance of the guide pin. During the braking process, L' 钳体 is the elastic deformation of the brake caliper body; L 活塞 L is the elastic deformation of the rectangular sealing ring caused by the movement of the piston; 外(内)摩擦片 is the elastic deformation of the friction plate. Then, the residual positive pressure between the friction plate and the brake disc when the brake is released is as follows:

[0047] F s =F Δ +f s +f x

[0048] =K(ΔL 活塞 -ΔL 钳体 -2ΔL 摩擦片 )+f(2)

[0049] Where, ΔL 活塞 is the piston return amount; ΔL 钳体 is the deformation of the clamp body; ΔL 摩擦片 is the deformation of the friction plate; K is the brake elastic coefficient; f is the sliding resistance. It can be deduced that the formation of brake drag is mainly related to the piston return amount, the deformation of the caliper body and the deformation of the friction plate.

[0050] S2. Based on the classification results of the causes of brake drag, a dynamic model of the brake drag mechanism is constructed.

[0051] In the actual piston sealing system, the sealing contact area is in a mixed lubrication state, and there are various wear forms affected by the surface roughness of the sealing ring, relative motion speed, oil film thickness and brake hydraulic pressure. Therefore, it is necessary to establish an accurate brake drag dynamics model from the perspective of fluid-solid coupling in the sealing contact area, accurately solve the friction force of the piston movement, and analyze the return ability of the piston.

[0052] In order to analyze the force characteristics of the piston seal contact area, this embodiment establishes a brake drag dynamic analysis model. It mainly includes: an elastohydrodynamic lubrication model considering the roughness of the contact surface; a rough surface contact pressure model; a seal contact surface microelastic deformation model; a seal contact surface oil film thickness model; and a seal contact surface friction calculation model. The coupling relationship between the models is as follows: Figure 4 shown.

[0053] (1) Elastic-hydrodynamic lubrication model considering surface roughness morphology

[0054] In the seal contact area, the brake fluid changes dynamically. Based on the law of conservation of mass, the two-dimensional Reynolds equation can be obtained by analyzing the fluid dynamics of the brake fluid. The Reynolds equation is used to solve the oil film pressure, providing a basis for the calculation of the fluid viscous shear stress and the oil film thickness. Figure 5 The coordinate system shown is used to transform the Reynolds equation and normalize it.

[0055]

[0056] A=c a / {(1+c b p)[1+(c a +c b )p]}-α (5)

[0057] In the formula, p cyl is the brake pressure; p is the local static pressure of the sealing fluid; h is the nominal oil film thickness; h0 is the average oil film thickness; W is the contact width of the seal ring; L is the inner diameter circumference of the seal ring; V is the relative motion speed; η is the brake fluid viscosity; c a 、c b is the fluid constant; α is the pressure viscosity coefficient. A coordinate system is established with the lower right vertex of the seal ring as the origin, inward as the positive direction of the x-axis, upward as the positive direction of the z-axis, and leftward as the positive direction of the y-axis; P represents the dimensionless pressure; H represents the dimensionless oil film thickness; Y represents the dimensionless longitudinal coordinate; X represents the dimensionless transverse coordinate; W represents the seal ring contact width; T represents the dimensionless time. The relative motion speed is positive when the brake piston is in the inner stroke and negative when it is in the outer stroke.

[0058] Since the thickness of the brake fluid film is affected by the roughness of the contact surface, and the rough peaks of the rough surface will cause contact deformation when the surface slides, the thickness distribution of the brake fluid film on the rough surface in the sealing contact area is as follows: Figure 6 shown.

[0059] The distance between the rough peak of the sealing ring surface and the reference line is δ2, and the distance between the rough peak of the piston surface and the reference line is δ1. The distance between the two reference lines is the nominal film thickness h, and the distance between the two surfaces is the actual oil film thickness h. T , the piston surface movement speed is U1, and its value is V.

[0060] The actual oil film thickness on the rough surface is:

[0061] h T =h+δ1+δ2. (6)

[0062] (2) Rough surface contact pressure model

[0063] The seal contact area is in a mixed lubrication state. According to the mixed lubrication theory, when the film thickness ratio is less than 3, the rough peaks of the two rough surfaces come into contact and generate contact pressure. When the contact pressure is calculated using the GW contact model, the piston is regarded as an ideal surface. The rough peak contact pressure is calculated by the contact model, and then the rough peak contact shear stress is obtained, which provides a basis for the calculation of the micro-deformation of the sealing ring surface in the next step.

[0064] The schematic diagram of the contact between the rough surface of the sealing ring and the piston surface is as follows Figure 7 As shown, the dotted line is the theoretical sealing ring surface, the solid line is the piston surface, the broken line is the actual sealing ring surface, z0 is the distance from the roughness peak to the sealing ring baseline, and d0 is the deformation of the roughness peak.

[0065] According to the Hertzian theory, the area A of a single rough peak after contact deformation can be obtained: i and contact force F i The calculation expression is:

[0066] A i =πRδ0 (7)

[0067]

[0068] Then the roughness peak contact pressure p of the sealing ring is obtained. c and asperity peak contact shear stress τ c :

[0069]

[0070] Where R is the radius of curvature of a single roughness peak; δ0 is the deformation of the roughness peak; E′ is the comprehensive elastic modulus; E R 、E S is the elastic modulus of the piston and sealing ring; μ R , μ S is the Poisson's ratio of the piston and the sealing ring; σ is the comprehensive surface roughness; Among them, σ1 is the surface roughness of the sealing ring; σ2 is the surface roughness of the piston; f c is the dry friction coefficient; the negative sign indicates that the direction of shear stress is opposite to the movement speed.

[0071] (3) Microelastic deformation model of sealing contact surface

[0072] In the sealing contact area, the sealing ring and the piston surface are affected by the disturbance pressure caused by the oil film pressure and the roughness of the sealing ring, which causes radial micro-deformation of the sealing ring and the rough surface of the piston. The micro-deformation of the sealing contact surface is calculated by the deformation equation, which provides the initial value for the next step of calculating the oil film thickness.

[0073] The entire sealing contact area is discretized using a rectangular grid, and each node has a unique number (i, j).

[0074] Normal pressure p at the node (i, j) in the sealing area ij Elastic micro-deformation w is generated at (k, l) p It is expressed as:

[0075]

[0076] in, Represents the elastic influence coefficient matrix; m and n are both natural numbers.

[0077] The expression is:

[0078]

[0079] Among them, x′, y′ are the coordinates of the force application point; x, y are the coordinates of the deformation point.

[0080] To avoid calculation When the function is discontinuous at the point (x, y), the sealing contact surface is transformed into a rectangular discrete area as shown below:

[0081]

[0082] In the formula, a and b are half of the width and length of the rectangular unit in the calculation area, respectively.

[0083] Subject to shear stress τ zx , τ zv Elastic micro-deformation The expressions of are shown in equations (15) to (16).

[0084]

[0085] Using the principle of small deformation and superposition, the micro deformation is added to obtain the total micro elastic deformation w of the piston rough surface. rod , whose expression is shown in formula (17).

[0086]

[0087] Since the elastic half space is not suitable for soft rectangular sealing rings, the main body of the sealing ring is divided into rectangular blocks perpendicular to the y-axis with dimensions of dx×dy×b1. The force diagram is shown in Figure 8 As shown. b1 is the thickness of the sealing ring; p sc is the static contact pressure; P0 is the pressure at the contact point between the seal ring and the piston, Q is the pressure at the contact point between the seal ring and the brake caliper cavity; Fa 、F b is the shear force.

[0088]

[0089] In formula (18), G seal is the shear modulus of the sealing ring; u z is the slight deformation of the rectangular block surface. The equilibrium equation of the normal force along the z-axis is as follows:

[0090]

[0091] In formula (19), u seal S is the slight deformation of the sealing ring surface; rod is the piston surface roughness height; is the radial elastic deformation of the piston surface caused by the braking pressure and temperature, and its expression is shown in formula (20).

[0092]

[0093] Where D rod is the outer diameter of the piston; d rod is the inner diameter of the piston; α rod is the thermal expansion coefficient of the piston; q is the operating temperature; q0 is the initial temperature. Introduce the dimensionless variable U = u seal / h0, and discretize equation (19) to obtain the micro deformation of the sealing ring surface.

[0094] (4) Oil film thickness model of sealing contact surface

[0095] Due to the dynamic coupling of lubrication / wear in the seal contact area, the oil film thickness and the seal surface wear affect each other and tend to dynamic equilibrium. After the friction and wear reach dynamic equilibrium, the oil film thickness meets the convergence condition, and the next step of friction force calculation can be carried out.

[0096] The force diagram of the rough surface is given by Fig. 9 Display, actual oil film thickness h T is the initial oil film thickness h s and the roughness peak deformation h after stress def The sum of the oil film pressure p in the sealing contact area f and asperity peak contact pressure p c With static contact pressure p sc dynamic equilibrium.

[0097] The expression of actual oil film thickness is:

[0098] h T =H s +h def (twenty one)

[0099]

[0100] Where I represents the lubrication characteristic parameter.

[0101] The new surface roughness profile is obtained by subtracting the wear depth calculated in each iteration from the original surface roughness profile. According to the Archard wear model, the wear depth h of each iteration is i for:

[0102]

[0103] In the formula, k w is the dynamic wear coefficient; H m is the surface hardness of the sealing ring; Δt represents the iteration time step.

[0104] The dynamic wear coefficient is related to the film thickness ratio and can be described as a function of the wear coefficient k0 under dry friction conditions, as shown in formula (25).

[0105]

[0106] Here, λ represents the film thickness ratio.

[0107] The expression of wear depth in each time interval is obtained:

[0108] h i.new =h i，old +h i (26)

[0109] Among them, h i.new Indicates the updated wear depth; h i，old Indicates the wear depth before renewal.

[0110] The oil film thickness including the wear correction is:

[0111] h T =h s +h def +h i，new (27)

[0112] (5) Calculation model of friction force on sealing contact surface

[0113] The friction force in the piston reciprocating seal contact area comes from the viscous shear stress of the brake fluid and the contact shear stress between the surface roughness peaks. The following boundary conditions are met when calculating the friction force:

[0114]

[0115] Where p is the local static pressure of the sealing fluid; y is the ordinate of the contact area; Wc is the actual contact width of the sealing ring; is the angular variable in the circumferential direction; z, u, v correspond to the vertical z, circumferential x and axial y of the sealing contact area respectively; h is the nominal oil film thickness; V is the relative motion speed between the piston surface and the sealing ring surface.

[0116] The oil pressure on the left side of the seal contact area is the brake hydraulic pressure p cyl , the right air side pressure is set to 0.

[0117] The viscous shear stress of the fluid is (h>0):

[0118]

[0119] Surface roughness peak contact shear stress (h = 0):

[0120]

[0121] in, represents the shear stress in the x and y directions of the piston, Indicates the shear stress of the sealing ring in the x and y directions.

[0122] When calculating the friction force, the actual contact width W of the seal ring should also be taken into account. c .

[0123] W c =(1+ε y )W-2y c (32)

[0124] Where η is the viscosity of the brake fluid; V is the relative speed; f c is the dry friction coefficient; ε y is the axial strain of the sealing ring; y c is the root of formula (32).

[0125]

[0126] In the formula, ε z represents the vertical strain of the sealing ring; ΔR represents the radial interference; r represents the chamfer radius of the sealing ring; b represents the axial width of the sealing ring; θ represents the operating temperature; θ0 represents the initial temperature; Represents the angle variable in the circumferential direction.

[0127] Contact area friction:

[0128]

[0129] After the model is built, it is solved. The process is as follows Fig.10 shown.

[0130] First, according to the braking condition parameters, the interpolation data of t-V (time-speed) and t-pcyl (time-pressure) are established, and the contact surface is numerically simulated; secondly, the static contact pressure of the sealing contact surface is calculated according to formula (35), the initial oil film thickness is calculated according to formula (21), and the calculation results are substituted into the Reynolds equation to obtain the oil film disturbance pressure; at the same time, the roughness peak contact pressure is calculated by combining the nominal oil film thickness with the GW contact model; then, the fluid viscosity shear stress and the roughness peak contact shear stress between the sealing contact surfaces are calculated according to formulas (30) and (31), the shear stress is substituted into the microelastic deformation equation of the sealing contact surface, the micro deformation of the sealing contact surface is calculated, and the roughness contact pressure is substituted into the wear model to calculate the dynamic wear profile; finally, all the calculation results are substituted into formula (36) to calculate the final actual oil film thickness. If converged, the oil film thickness is updated and the friction force of the contact area is calculated using the shear stress; otherwise, it is necessary to recycle until the oil film thickness converges.

[0131]

[0132] Among them, E seal Indicates the elastic modulus of the sealing ring; ε x Indicates the circumferential strain of the sealing ring; ε z Represents the vertical strain of the sealing ring.

[0133]

[0134] In the formula, Indicates the deformation of the piston surface under the installation pressure of the sealing contact surface.

[0135] S3. Extract typical braking conditions based on user operation data.

[0136] In order to clarify the correlation between brake drag and braking conditions, typical braking conditions were identified and extracted based on user operation data. This embodiment extracted braking conditions from user feedback data collected from a total of 446 vehicles, and used the K-means clustering method to obtain three types of typical braking conditions. The main characteristic parameters include braking speed, braking deceleration, brake pedal travel, brake pedal change rate, etc.

[0137] The average values ​​of characteristic parameters of three typical braking conditions are statistically analyzed, and the specific values ​​are shown in Tables 1 and 2. According to the characteristic parameters, condition 1 is defined as a low-speed short braking condition; condition 2 is defined as a medium-low speed long braking condition; and condition 3 is defined as a medium-high speed short braking condition.

[0138] Table 1

[0139]

[0140] Table 2

[0141]

[0142] S4. Based on the dynamic model of brake drag mechanism and typical braking conditions, complete the identification of conditions prone to brake drag.

[0143] The dynamic model of brake drag mechanism takes piston movement speed and sealing pressure as input, so it is necessary to establish the AMESim model of the brake system to convert typical braking conditions into model input parameters, and the AMESim model simulation results can be used as an auxiliary basis for determining brake drag.

[0144] The brake drag mechanism dynamics model is used to analyze the piston reset force and sliding resistance. When the reset force is less than the sliding resistance, it is determined that brake drag is likely to occur under this operating condition. This embodiment stipulates that brake drag can be determined when the piston displacement exceeds the single-sided brake clearance or the drag torque is greater than or equal to 5Nm.

[0145] By inputting the piston movement speed and brake fluid seal pressure time history data into the brake drag mechanism model and performing dynamic simulation of the reciprocating stroke of the brake caliper piston, the friction force time history curve of the piston seal contact area under typical braking conditions can be obtained, such as Fig.11 shown.

[0146] Fig.11 (a) and (b) show that there is a period of alternating positive and negative friction at the beginning of the outer stroke. This is because after the hydraulic thrust is greater than the maximum static friction, the piston begins to move, and the sealing ring is affected by the shear force to produce tangential deformation; the hydraulic pressure continues to increase during the outer stroke of the piston, and the thrust will once again exceed the maximum static friction, and the sealing ring will begin to slide relative to the piston, so there will be a stage of frequent switching between dynamic and static friction. In the entire reciprocating stroke of the brake piston, the maximum static friction value at the beginning of the outer stroke is the piston reset force provided by the sealing ring in the inner stroke, and the maximum friction in the inner stroke of the piston is the sliding resistance. The reset force of condition 1 is 63.08N, which is the same as the sliding resistance; the reset force of condition 2 is 63.0623N, and the sliding resistance is 66.9276N; the reset force of condition 3 is 58.2432N, and the sliding resistance is 58.2418N. Comparing the magnitude of the reset force and the piston sliding resistance, it can be obtained that under the medium and low speed long braking condition (condition 2), the reset force provided by the sealing ring is less than the piston sliding resistance, and the piston cannot return to its position.

[0147] In order to further determine the working conditions that are prone to brake drag, the time history of the brake piston displacement is first solved, and then the brake system model established above is used to simulate three typical braking conditions to obtain the time history of the brake clearance under different conditions. The result is that in working condition 2, in the first reciprocating stroke (0-7.64s), the piston finally moves to 0.0064mm from the initial position, and the piston can return. In the second reciprocating stroke (15.64-20)s, the piston displacement stops at 0.1554mm from the initial position, which is larger than the unilateral brake clearance (0.12mm), so the piston cannot be reset.

[0148] Depend on Fig.12 It can be seen that the brake clearance fluctuation of working condition 2 is larger than that of the other two working conditions, and the brake clearance value is the smallest at the end of the working condition braking time, and it is a negative value, indicating the compression of the friction plate, indicating that the friction plate is in contact with the brake disc. Combining the brake system drag index of the three typical braking conditions, as shown in Table 3, it can be clearly seen that the required fluid volume is inversely correlated with the drag torque, the residual brake pressure and the relative displacement of the piston are positively correlated with the drag torque, and the brake clearance is inversely correlated with the drag torque.

[0149] Table 3

[0150]

[0151] In summary, it can be determined that the medium-low speed long braking condition (condition 2) is a condition where brake drag is prone to occur.

[0152] Embodiment 2

[0153] The present embodiment also provides a commercial vehicle disc brake drag identification system under dynamic conditions, including: an acquisition module, a construction module, an extraction module and an identification module; the acquisition module is used to collect user disc brake fault feedback information and classify the causes of brake drag; the construction module is used to construct a brake drag mechanism dynamic model based on the classification results of the brake drag causes; the extraction module is used to extract typical braking conditions based on user operation data; the identification module is used to complete the identification of brake drag-prone conditions based on the brake drag mechanism dynamic model and typical braking conditions.

[0154] The following will describe in detail how the present invention solves the technical problems in practical work in conjunction with this embodiment.

[0155] The acquisition module is used to collect user disc brake fault feedback information and classify the causes of brake drag.

[0156] For a certain type of commercial vehicle disc brake, the user fault feedback data is analyzed and the main failure modes of the brake are counted. According to the fault description in the feedback information and the faulty parts analysis, the main cause of brake drag is determined, and the main influencing factors are clarified in combination with the mathematical model of the drag torque.

[0157] Based on the fault information reported by users, the causes of brake drag are classified, such as Figure 2 As shown. Figure 2 (a) It can be seen that brake drag is the main failure mode of the brake, accounting for about 58%; among the six reasons that lead to brake drag, floating guide failure and brake piston return failure account for 20.24% and 73.81% respectively. Figure 2 (b) As shown. Obviously, insufficient return ability of the brake piston is the main cause of brake drag. Further, according to the fault description information, it is confirmed that the secondary factors causing poor piston return are mainly the piston return amount and piston sliding resistance.

[0158] Then the factors affecting brake drag are analyzed. The return ability of the brake piston is related to the static and dynamic sealing mechanism of the rectangular seal ring in the brake caliper. After the seal ring is installed on the piston, it will be in a pre-compressed state and bear the brake fluid pressure during operation. When the brake fluid pressure is always less than the static friction force generated by the pressure on the pre-compressed seal contact surface of the seal ring, the static seal is completed. In the actual working process, there is a brake fluid film with a thickness of microns between the contact surfaces of the seal ring and the piston. When the piston is stationary, the brake fluid film is kept in a certain position under the action of surface tension, thereby achieving dynamic sealing.

[0159] Brake drag occurs during the piston stroke after the brake is released. The piston returns to its original position by relying on the energy of the seal ring's compression deformation, and the reset force is equal to the maximum static friction force of the static seal; the brake fluid film in the dynamic seal will affect the lubrication and friction wear state of the seal contact area, thereby affecting the piston sliding resistance. If the reset force is less than the sliding resistance, residual resistance will be generated between the friction plate and the brake disc, thus forming a drag torque.

[0160] The drag torque can be expressed as the residual braking torque of the friction plate and the brake disc, as shown in the following formula:

[0161] T 拖滞 =F μ ×R e =μ·F s ×R e (37)

[0162] Where, T 拖滞 Indicates the drag torque; F μ Represents the residual friction of the friction plate; R erepresents the effective braking radius; μ represents the friction coefficient of the friction plate; F s Indicates the positive pressure between the friction pad and the brake disc.

[0163] Generally, the effective radius of the brake is a fixed value, so the drag torque is mainly related to the positive pressure F between the friction pad and the brake disc. s Therefore, a mechanical model of brake drag formation is established to analyze Fs, such as Figure 3 shown.

[0164] In this model, the spring stiffness is used to represent the stiffness of each component, where K t K is the clamp body stiffness; c K is the sealing ring stiffness; p is the friction plate stiffness; f s is the sliding resistance of the friction plate in the slideway; f x is the sliding resistance of the guide pin. During the braking process, L' 钳体 is the elastic deformation of the brake caliper body; L 活塞 L is the elastic deformation of the rectangular sealing ring caused by the movement of the piston; 外(内)摩擦片 is the elastic deformation of the friction plate. Then, the residual positive pressure between the friction plate and the brake disc when the brake is released is as follows:

[0165] F s =F Δ +f s +f x

[0166] =K(ΔL 活塞 -ΔL 钳体 -2ΔL 摩擦片 )+f(38)

[0167] Where, ΔL 活塞 is the piston return amount; ΔL 钳体 is the deformation of the clamp body; ΔL 摩擦片 is the deformation of the friction plate; K is the brake elastic coefficient; f is the sliding resistance. It can be deduced that the formation of brake drag is mainly related to the piston return amount, the deformation of the caliper body and the deformation of the friction plate.

[0168] The construction module constructs a dynamic model of the brake drag mechanism based on the classification results of the causes of brake drag.

[0169] In the actual piston sealing system, the sealing contact area is in a mixed lubrication state, and there are various wear forms affected by the surface roughness of the sealing ring, relative motion speed, oil film thickness and brake hydraulic pressure. Therefore, it is necessary to establish an accurate brake drag dynamics model from the perspective of fluid-solid coupling in the sealing contact area, accurately solve the friction force of the piston movement, and analyze the return ability of the piston.

[0170] In order to analyze the force characteristics of the piston seal contact area, this embodiment establishes a brake drag dynamic analysis model. It mainly includes: an elastohydrodynamic lubrication model considering the roughness of the contact surface; a rough surface contact pressure model; a seal contact surface microelastic deformation model; a seal contact surface oil film thickness model; and a seal contact surface friction calculation model. The coupling relationship between the models is as follows: Figure 4 shown.

[0171] The extraction module extracts typical braking conditions based on user operation data. In order to clarify the correlation between brake drag and braking conditions, typical braking conditions are identified and extracted based on user operation data. This embodiment extracts braking conditions from user feedback data collected from a total of 446 vehicles, and uses the K-means clustering method to obtain three types of typical braking conditions. The main characteristic parameters include braking speed, braking deceleration, brake pedal travel, brake pedal change rate, etc.

[0172] The average values ​​of characteristic parameters of three typical braking conditions are statistically analyzed, and the specific values ​​are shown in Tables 1 and 2. According to the characteristic parameters, condition 1 is defined as a low-speed short braking condition; condition 2 is defined as a medium-low speed long braking condition; and condition 3 is defined as a medium-high speed short braking condition.

[0173] The final identification module completes the identification of conditions prone to brake drag based on the dynamic model of the brake drag mechanism and typical braking conditions.

[0174] The dynamic model of brake drag mechanism takes piston movement speed and sealing pressure as input, so it is necessary to establish the AMESim model of the brake system to convert typical braking conditions into model input parameters, and the AMESim model simulation results can be used as an auxiliary basis for determining brake drag.

[0175] The brake drag mechanism dynamics model is used to analyze the piston reset force and sliding resistance. When the reset force is less than the sliding resistance, it is determined that brake drag is likely to occur under this operating condition. This embodiment stipulates that brake drag can be determined when the piston displacement exceeds the single-sided brake clearance or the drag torque is greater than or equal to 5Nm.

[0176] By inputting the piston movement speed and brake fluid seal pressure time history data into the brake drag mechanism model and performing dynamic simulation of the reciprocating stroke of the brake caliper piston, the friction force time history curve of the piston seal contact area under typical braking conditions can be obtained, such as Fig.11 shown.

[0177] The embodiments described above are only descriptions of the preferred embodiments of the present invention and are not intended to limit the scope of the present invention. Without departing from the design spirit of the present invention, various modifications and improvements made to the technical solutions of the present invention by ordinary technicians in this field should all fall within the protection scope determined by the claims of the present invention.

Claims

1. A method for identifying the drag of a commercial vehicle disc brake under dynamic conditions, characterized in that the steps include: S1. Collect user disc brake fault feedback information and classify the causes of brake drag; S2. Based on the classification results of the causes of brake drag, a dynamic model of the brake drag mechanism is constructed; S3. Extract typical braking conditions based on user operation data; S4. Based on the brake drag mechanism dynamic model and the typical braking conditions, identify the conditions where brake drag is likely to occur.

2. The method for identifying the drag of a commercial vehicle disc brake under dynamic conditions according to claim 1, characterized in that: Step S1 includes: collecting and analyzing user fault feedback data, and counting the main fault modes of the brake; determining the main cause of brake drag based on the fault description and fault component analysis in the feedback information, and clarifying the main influencing factors in combination with the drag torque mathematical model.

3. The method for identifying the drag of a commercial vehicle disc brake under dynamic conditions according to claim 1, characterized in that: The constructed dynamic model of the brake drag mechanism includes: an elastohydrodynamic lubrication model considering the roughness of the contact surface; a rough surface contact pressure model; a seal contact surface microelastic deformation model; a seal contact surface oil film thickness model; and a seal contact surface friction calculation model.

4. The method for identifying the drag of a commercial vehicle disc brake under dynamic conditions according to claim 3, characterized in that: The method for solving the dynamic model of the brake drag mechanism includes: establishing t-V and t-pcyl interpolation data according to the braking condition parameters, and performing numerical simulation on the contact surface; secondly, calculating the static contact pressure of the sealing contact surface and the initial oil film thickness, and substituting the calculation results into the Reynolds equation to obtain the oil film disturbance pressure; at the same time, using the GW contact model combined with the nominal oil film thickness to calculate the roughness peak contact pressure; then, calculating the fluid viscosity shear stress and the roughness peak contact shear stress between the sealing contact surfaces, substituting the shear stress into the microelastic deformation equation of the sealing contact surface, calculating the microdeformation of the sealing contact surface, and substituting the roughness contact pressure into the wear model to calculate the dynamic wear profile; finally, calculating the final actual oil film thickness based on all the calculation results, and when the oil film thickness meets the convergence condition, using the friction force calculation model to calculate the friction force of the sealing contact area.

5. The method for identifying the drag of a commercial vehicle disc brake under dynamic conditions according to claim 1, characterized in that: Step S3 includes: extracting braking conditions from user feedback data, using K-means clustering method to obtain three types of typical braking conditions, and characteristic parameters include: braking speed, braking deceleration, brake pedal travel and brake pedal change rate.

6. The method for identifying the drag of a commercial vehicle disc brake under dynamic conditions according to claim 1, characterized in that: Step S4 includes: the brake drag mechanism dynamic model takes the piston movement speed and the sealing pressure as input, and converts the typical braking conditions into model input parameters by establishing the AMESim model of the braking system. The brake drag mechanism dynamic model is used to analyze the piston reset force and the sliding resistance. When the reset force is less than the sliding resistance, it is determined that brake drag is likely to occur under this condition.

7. A commercial vehicle disc brake drag identification system under dynamic conditions, the system being used to implement the method according to any one of claims 1 to 6, characterized in that: include: Acquisition module, construction module, extraction module and recognition module; The collection module is used to collect user disc brake fault feedback information and classify the causes of brake drag; The construction module is used to construct a brake drag mechanism dynamics model based on the classification results of the brake drag causes; The extraction module is used to extract typical braking conditions based on user operation data; The identification module is used to identify the brake dragging condition that is prone to occur based on the brake dragging mechanism dynamics model and the typical brake condition.

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