Hybrid power transmission system parameter optimization method based on time-frequency analysis

Abstract

The invention provides a hybrid power transmission system parameter optimization method based on time-frequency analysis. The hybrid power transmission system parameter optimization method comprises the following steps: step 1, establishing a dynamic model of a hybrid power transmission system; step 2, establishing a total transient torque model Te of the engine; step 3, establishing a transmission torque model Td of the torsion damper; 4, calculating the total rotational inertia J3 of a gear on an input shaft of the gearbox and the total rotational inertia J4 of a gear on an output shaft of the gearbox; 5, the rotational inertia Jbody of the vehicle body, the output torque Tmg2 of the driving motor MG2 and the resistance torque T1 are calculated; step 6, establishing a motion differential equation set in a hybrid power mode; step 7, performing time-frequency analysis based on the motion differential equation set; 8, selecting a plurality of key parameters as optimization variables; step 9, setting a constraint condition and a target function; and step 10, executing a genetic algorithm to obtain an optimal solution. According to the method, a hybrid power transmission system model is established, and time-frequency analysis and a multi-target genetic algorithm are combined, so that optimal design of the transmission system is realized.

Description

Technical Field

[0001] The invention belongs to the technical field of vibration control of new energy vehicles and relates to a hybrid power transmission system parameter optimization method based on time-frequency analysis. Background Art

[0002] The transmission system of hybrid vehicles is prone to torsional vibration problems due to the complex power source coupling characteristics. These problems not only affect the NVH performance of the vehicle, but may also have an adverse effect on the reliability and life of the transmission system. Therefore, torsional vibration analysis and parameter optimization of the transmission system have become key links in improving the performance of hybrid vehicles.

[0003] Traditional research focuses on the analysis of the inherent characteristics of the transmission system, and analyzes the system's natural frequency and vibration mode through simplified modeling methods. However, the hybrid powertrain is affected by a variety of nonlinear excitations in actual operation, and its torsional vibration response has obvious time-varying and frequency-varying characteristics. Research based solely on static analysis is difficult to fully reflect the dynamic behavior of the system.

[0004] As a tool that can describe signal characteristics in both the time domain and the frequency domain, time-frequency analysis has important application value in the dynamic response analysis of transmission systems. Through the time-frequency analysis method, the vibration characteristics of the system under different working conditions can be captured, and the influence of key parameters on torsional vibration response can be deeply revealed. However, in existing research, the exploration of transmission system parameter optimization is mostly limited to single-objective optimization, and the time-frequency analysis technology is not fully combined, resulting in limited applicability and effect of the optimization results under actual working conditions.

[0005] Therefore, combining time-frequency analysis technology to study the dynamic response characteristics of the hybrid powertrain system and conducting multi-objective parameter optimization based on this will provide an effective technical approach to improve the torsional vibration resistance of the powertrain system and the operational adaptability of the vehicle. Summary of the invention

[0006] The purpose of the present invention is to provide a hybrid powertrain system parameter optimization method based on time-frequency analysis, by establishing a hybrid powertrain system model covering components such as an engine, a motor, a gearbox, a power coupling device, etc., combining time-frequency analysis and a multi-objective genetic algorithm to achieve optimal design of the transmission system.

[0007] A hybrid powertrain system parameter optimization method based on time-frequency analysis comprises the following steps:

[0008] Step 1: Establish a dynamic model of the hybrid powertrain system, which includes:

[0009] An engine and a flywheel, wherein the flywheel is connected to an output shaft of the engine;

[0010] A torsional vibration damper, wherein the primary mass is fixedly connected to the flywheel and the secondary mass is connected to the input shaft of the gearbox; a vibration damping spring is arranged between the secondary mass and the primary mass;

[0011] The gearbox comprises an output shaft and an input shaft;

[0012] The main reducer-differential includes two output shafts, each of which is connected to a half shaft, and the half shaft is equipped with a driving wheel; the main reducer-differential is connected to the gearbox and the driving motor MG2 through a gear set;

[0013] Step 2: Establish the total transient torque model T of the engine e ;

[0014] Step 3: Establish the transmission torque model T of the torsional vibration damper d ;

[0015] Step 4, calculate the total rotational inertia J3 of the gears on the input shaft of the gearbox and the total rotational inertia J4 of the gears on the output shaft;

[0016] Step 5: Calculate the moment of inertia J of the vehicle body body , output torque T of drive motor MG2 mg2 , resistance torque T1;

[0017] Step 6: Create a file containing T e , J3, J4, J body , T mg2 , the differential equations of motion of the powertrain in hybrid mode of T1;

[0018] Step 7: Perform time-frequency analysis based on the differential equations of motion;

[0019] Step 8: Select several key parameters as optimization variables;

[0020] Step 9: Set constraints and objective function;

[0021] Step 10: Execute the genetic algorithm to obtain the optimal solution.

[0022] Preferably, the expressions of J3 and J4 in step 4 are specifically:

[0023]

[0024] J a1 is the moment of inertia of the ML gear driving gear (1);

[0025] J a2 is the moment of inertia of the ML gear driven gear (2);

[0026] J a3is the moment of inertia of the MH gear driving gear (3);

[0027] J a4 is the moment of inertia of the MH gear driven gear (4);

[0028] J a5 is the moment of inertia of the MG1 motor driving gear (5);

[0029] J a6 is the moment of inertia of the driven gear (6) of the MG1 motor;

[0030] i ml It is the transmission ratio of ML gear;

[0031] i mh It is the transmission ratio of MH gear;

[0032] i mg1 is the gear ratio of the generator MG1.

[0033] Preferably, in step 5, J body , T mg2 The specific expressions of and T1 are:

[0034]

[0035] T l =(F f +F w )·r w

[0036]

[0037] F f =fmg

[0038] m is the vehicle mass, which is the set value;

[0039] r w is the radius of the driving wheel, is the set value;

[0040] i g and i0 are the transmission ratios of the gearbox and the final reducer respectively;

[0041] T tar is the motor target torque, is the set value;

[0042] τ is the motor torque response time;

[0043] T max The maximum torque output by the driving motor MG2;

[0044] t is the simulation time, is the set value;

[0045] F fis the rolling resistance;

[0046] F w is air resistance;

[0047] f is the rolling resistance coefficient between the driving wheel and the road surface;

[0048] m is the vehicle mass;

[0049] C D is the air resistance coefficient;

[0050] A is the windward area of ​​the vehicle;

[0051] u is the vehicle speed.

[0052] Preferably, the set of motion differential equations in step 6 is:

[0053]

[0054] J i is the equivalent moment of inertia of the ith component;

[0055] θ i is the rotation angle of the i-th component;

[0056] k ij is the equivalent torsional stiffness between the i-th component and the j-th component; c ij is the equivalent viscous damping between the i-th component and the j-th component;

[0057] Among them, the value set of i and j includes the engine torsion reducer, the gearbox input shaft, the gearbox output shaft, the main reducer, the left driving force, the right driving wheel, the generator MG1 rotor, and the drive motor MG2 rotor.

[0058] Preferably, the optimization variables in step 8 are:

[0059] x=[J2,k 12 ,c 12 ,k 23 ,k 34 ,k 56 ,k 57 ,J 10 ]

[0060] J2-equivalent moment of inertia of the engine;

[0061] k 12 -Equivalent torsional stiffness between the engine and the torsional vibrator;

[0062] c 12 -Equivalent viscous stiffness between the engine and the torsional vibrator;

[0063] k 23-Equivalent torsional stiffness between the torsional vibrator and the gearbox input shaft;

[0064] k 34 -Equivalent torsional stiffness between the gearbox input shaft and the gearbox output shaft;

[0065] k 56 -Equivalent torsional stiffness between the final reducer and the left drive wheel;

[0066] k 57 -Equivalent torsional stiffness between the final reducer and the right drive wheel;

[0067] J 10 -Equivalent moment of inertia of the rotor of the generator MG1.

[0068] Preferably, the objective function in step 9 is: minimizing the root mean square value of the second-order angular acceleration of the gearbox input shaft, the left drive wheel, and the right drive wheel.

[0069] Compared with the prior art, the advantages of the present invention are:

[0070] Through multi-objective genetic algorithm optimization based on time-frequency analysis, the anti-torsional vibration performance and vibration control effect of the hybrid vehicle transmission system are significantly improved. This method effectively reduces the second-order angular acceleration of the gearbox input shaft and the left and right drive wheels by accurately modeling the dynamic characteristics of key components, reduces the vibration amplitude and noise, and improves the NVH performance of the vehicle. Compared with traditional optimization methods, this invention ensures that the optimization parameters are within a reasonable range, has strong versatility and application value, and can provide strong support for the design of hybrid power transmission systems. BRIEF DESCRIPTION OF THE DRAWINGS

[0071] Figure 1 is a schematic diagram of the structure of a hybrid powertrain system;

[0072] Figure 2 Kinematic and dynamic analysis diagrams of the engine crank-connecting rod mechanism;

[0073] Figure 3 It is the torsional characteristic curve diagram;

[0074] Figure 4 It is the torsional vibration model of the whole vehicle in hybrid mode;

[0075] Figure 5(a) to Figure 5(c) It is the system time-frequency analysis process;

[0076] Figure 6 Flowchart of the multi-objective optimization algorithm.

[0077] Among them, 1-ML gear driving gear, 2-ML gear driven gear, 3-MH gear driving gear, 4-MH gear driven gear, 5-MG1 motor driving gear, 6-MG1 motor driven gear. DETAILED DESCRIPTION

[0078] The hybrid powertrain parameter optimization method based on time-frequency analysis of the present invention will be described in more detail below in conjunction with a schematic diagram, wherein a preferred embodiment of the present invention is shown, and it should be understood that those skilled in the art can modify the present invention described herein while still achieving the advantageous effects of the present invention. Therefore, the following description should be understood as being widely known to those skilled in the art, and not as a limitation of the present invention.

[0079] 1. Establish a dynamic model of the hybrid powertrain system.

[0080] The hybrid powertrain includes:

[0081] an engine and a flywheel, the flywheel being connected to an output shaft of the engine;

[0082] The primary mass of the torsional vibration damper is fixedly connected to the flywheel, and together they serve as the power input end to transfer the engine torque to the secondary mass.

[0083] The secondary mass of the torsional vibration damper is connected to the primary mass through the damping spring and directly connected to the gearbox input shaft to complete the power output.

[0084] The gearbox has an input shaft connected to the output end of the torsion damper, and a plurality of gears are arranged on the output shaft and the input shaft.

[0085] The gearbox is a two-shaft, two-speed gearbox, which achieves gear switching by adjusting the inner and outer gear hub engagement state of the MH (high gear) and ML (low gear) driven gears.

[0086] When the gearbox is in ML gear, the shift actuator pushes the outer gear hub of the ML driven gear 2 to engage with the inner gear hub. At this time, the power transmission path is that the ML gear driving gear 1 on the input shaft is meshed with the ML gear driven gear 2 on the output shaft to complete the power transmission. Although the MH gear driving gear 3 and the MH gear driven gear 4 are long meshed, the MH gear driven gear 4, MG1 motor driving gear 5 and MG1 motor driven gear 6 on its branch road are all in idling state.

[0087] When the gearbox switches to MH gear, the shift actuator pushes the outer gear hub of the MH driven gear 4 to combine with the inner gear hub. At this time, the power is transmitted from the MH gear driving gear 3 on the input shaft to the output shaft through the MH gear driven gear 4. The MH gear driven gear 4 is fixedly connected to the MG1 motor driving gear 5. The MG1 motor driving gear 5 drives the MG1 motor driven gear 6 through a fixed transmission ratio, and the branch where the ML gear driving gear 1 and the ML gear driven gear 2 are located is in idling in this state.

[0088] The main reducer-differential includes two output shafts, each of which is connected to a half-shaft on which a drive wheel is mounted; wherein the main reducer-differential is meshed with the ML gear driven gear 2 and the drive motor MG2 respectively through a gear set.

[0089] Among them, the connection relationship between the main reducer-differential and the ML gear driven gear 2 and the drive motor MG2 is the existing technology.

[0090] The drive motor MG2 and the engine transmit torque to the final reducer-differential respectively.

[0091] In summary, an overall model of the hybrid powertrain system covering components such as the engine, motor, gearbox, power coupling device, etc. is constructed. The model should include the engine, generator, drive motor, gearbox, final reducer, and power coupling device, and the concentrated mass method and simplified model are used to describe the dynamic characteristics of each major component.

[0092] 2. Establish the total transient torque model of the engine.

[0093] The engine transient torque model is based on the kinematics and dynamics of the crank-connecting rod mechanism. It comprehensively considers the gas pressure in the cylinder and the force of the equivalent reciprocating mass. It can accurately describe the periodic fluctuation characteristics of the engine output torque and provide an important excitation source for the torsional vibration analysis of the vehicle powertrain system.

[0094] like Figure 2 As shown, first, according to the kinematic relationship of the crank-connecting rod mechanism, the crankshaft angle α, the connecting rod swing angle β, and the piston displacement s are defined. k The relationship between them is:

[0095] lsinβ=rsinα (1)

[0096]

[0097] Among them, r is the crank radius, which is the set value;

[0098] l is the connecting rod length, which is the set value;

[0099] λ p is the ratio of crank radius to connecting rod length, which is the set value;

[0100] The piston force F includes the equivalent reciprocating mass inertia force F of the piston, piston ring, piston pin and part of the connecting rod. m and the in-cylinder gas pressure F g Two parts:

[0101]

[0102] in:

[0103]

[0104] Among them, P g (ωt) is the in-cylinder fuel gas pressure that changes with the crankshaft angle and is a table value determined based on the instantaneous crankshaft angle;

[0105] J1 is the equivalent moment of inertia of the engine, which is the set value;

[0106] ω is the crankshaft angular velocity, which is obtained by formula (3);

[0107] d p is the piston diameter, is the set value;

[0108] m p is the equivalent reciprocating mass, is the set value.

[0109] The transient torque of a single cylinder is obtained from the piston force, which is composed of the reciprocating inertia torque and the gas pressure torque:

[0110] T e =T m +T g (6)

[0111]

[0112]

[0113] 3. Establish the torque transmission model between flywheel and torsional vibration damper.

[0114] The torsional vibration damper is a key component for attenuating the torsional vibration transmitted from the engine to the transmission system. Its basic structure includes a flywheel, a primary mass, a secondary mass, and a vibration damping spring and damping assembly.

[0115] Specifically, the primary and secondary masses refer to the mass of the flywheel and the damper fixed part and the mass connected to the gearbox input shaft, respectively. The primary mass of the flywheel and the torsional vibration damper are fixed together, which is simplified as the first inertia of a system. The secondary mass is the mass represented by the other part of the torsional vibration damper (that is, the part connected to the gearbox input shaft), which is simplified as the second inertia. Through this model, the total inertia of the flywheel and the torsional vibration damper can be regarded as the sum of the inertia of these two parts.

[0116] In addition, the shock absorber adopts a two-level stiffness design. The first and second stiffness of the torsional vibration damper correspond to the stiffness provided by the two levels of damping springs. The first level torsional stiffness is provided by the first level damping spring, which mainly affects the torsional vibration suppression on the driving side; the second level torsional stiffness is provided by the second level damping spring, which mainly affects the torsional vibration suppression on the dragging side.

[0117] Among them, "primary spring" refers to the vibration-damping spring between the secondary mass and the primary mass, when it is in the first torsional compression state; "first-stage torsional stiffness" refers to the stiffness of the primary spring.

[0118] "Secondary spring" refers to the damping spring between the secondary mass and the primary mass when it is in the second torsional compression state; "secondary torsional stiffness" refers to the stiffness of the secondary spring.

[0119] like Figure 3 As shown, the transmitted torque T of the torsional vibration damper d Divided into elastic torque T s and hysteresis damping torque T h There are two parts. The elastic torque is determined by the stiffness characteristics of the shock absorber, while the hysteresis damping torque is related to the dynamic response of the shock absorber. The specific expression is as follows:

[0120] T d =T s +T h (10)

[0121] in:

[0122]

[0123] Among them, θ r is the relative rotation angle between the primary mass and the secondary mass of the torsional vibration damper, and is the set value;

[0124] T d is the transmitted torque of the torsional vibration damper, is the simulation output value;

[0125] φ 23 ,φ 23 'The limit turning angles of the driving side and the dragging side are set values;

[0126] φ 12 is the switching angle between the first-stage torsional stiffness and the second-stage torsional stiffness on the drive side of the torsional vibration damper, which is the set value;

[0127] k1 is the first-stage torsional stiffness of the drive side of the torsional vibration damper, which is the set value;

[0128] k2 is the second-stage torsional stiffness of the drive side of the torsional vibration damper, which is the set value;

[0129] H1 is the hysteresis damping torque on the drive side of the torsional vibration damper, which is the set value;

[0130] φ 12 ′ is the switching angle between the first-stage torsional stiffness and the second-stage torsional stiffness on the dragging side of the torsional vibration damper, which is the set value;

[0131] k1′ is the first-level torsional stiffness of the torsional vibration damper on the dragging side, which is the set value;

[0132] k2′ is the second-level torsional stiffness of the torsional damper on the dragging side, which is the set value;

[0133] H1′ is the hysteresis damping torque on the dragging side of the torsional vibration damper, which is the set value.

[0134] Among them, "driving side" refers to the primary mass side; "drag side" refers to the secondary mass side.

[0135] The term "drive-side" refers to the section of the shock-absorbing spring that is connected to the primary mass.

[0136] The "trailing side" refers to the section of the shock absorber spring that connects to the secondary mass.

[0137] 4. Calculate the total rotational inertia J3 of the gears on the input shaft and the total rotational inertia J4 of the gears on the output shaft of the gearbox.

[0138] The gearbox model is mainly used to describe the power transmission and torque regulation characteristics of the gearbox in the hybrid powertrain system. This model uses a two-shaft two-speed gearbox with two working gears, ML and MH. The structure of the gearbox realizes the selection and switching of different gears through the internal shifting mechanism and the meshing between the external gear hub and the gear. The difference in power transmission structure between ML and MH also needs to be considered separately in the modeling process.

[0139] The gearbox can be simplified into two inertial elements, two elastic elements and two damping elements.

[0140] The inertial elements include gears on the input and output shafts of the gearbox;

[0141] The elastic element represents the torsional stiffness of the input shaft and output shaft of the gearbox.

[0142] The damping element represents the damping of the gearbox input shaft and output shaft.

[0143] Since the gear meshing relationship of the gearbox is different in different gears, it is necessary to calculate the transmission system in ML gear and MH gear separately.

[0144] In ML or MH gear, the total moment of inertia of the gears on the input and output shafts of the gearbox is calculated as follows:

[0145]

[0146] Among them, J a1 is the moment of inertia of the ML gear driving gear 1, which is the set value;

[0147] J a2 is the moment of inertia of the driven gear 2 of the ML gear, which is the set value;

[0148] J a3 is the moment of inertia of the MH gear driving gear 3, which is the set value;

[0149] J a4 is the moment of inertia of the driven gear 4 of MH gear, which is the set value;

[0150] J a5 is the moment of inertia of the driving gear of the MG1 motor, which is the set value;

[0151] J a6 is the moment of inertia of the driven gear 6 of the MG1 motor, which is the set value;

[0152] i ml It is the transmission ratio of ML gear;

[0153] i mh It is the transmission ratio of MH gear;

[0154] i mg1 is the gear ratio of the generator MG1.

[0155] i ml is the transmission ratio of ML gear, is the set value; i ml = speed of driving gear 1 / ML speed of driven gear 2;

[0156] i mh is the MH gear transmission ratio, is the set value; i mh = speed of MH gear driving gear 3 / speed of MH gear driven gear 4;

[0157] i mg1 is the gear ratio of the generator MG1, which is a set value.

[0158] 5. Calculate the moment of inertia J of the vehicle body body , output torque T of drive motor MG2 mg2 , resistance torque T1.

[0159] (1) Main reducer-differential model

[0160] The final drive-differential model integrates the final drive and differential.

[0161] The main reducer achieves speed reduction and torque increase through gear transmission, and the differential is used to distribute torque when the wheel speeds are inconsistent. When the working condition does not include turning, the differential evenly distributes the torque output by the gearbox to the left and right wheels. Since the gear stiffness of the main reducer and differential is large and the rotational inertia is concentrated, these components can be simplified as inertial elements.

[0162] (2) Semi-axle model

[0163] The inertia of the half shaft is small, and it is usually simplified as an elastic element, and the asymmetry of the left and right half shafts is taken into consideration. Its inertia is evenly distributed between the main reducer and differential at both ends and the wheels.

[0164] (3) Body model

[0165] The driving wheel has a large moment of inertia, so the tire part can be regarded as an inertial element and is connected to the vehicle body through an elastic element. The moment of inertia of the vehicle body is equivalently converted into the moment of inertia at the base speed of the engine crankshaft through the principle of conservation of kinetic energy. The calculation formula is as follows:

[0166]

[0167]

[0168] In the formula, m is the vehicle body mass and is the set value;

[0169] r w is the radius of the driving wheel, is the set value;

[0170] i g i0 and i1 are the transmission ratios of the gearbox and the main reducer in the main reducer-differential model, respectively, and ω is the set value; w is the angular velocity of the driving wheel, which is an intermediate quantity;

[0171] v is the speed of the driving wheel, which is an intermediate quantity;

[0172] (4) Drive motor MG2

[0173] The output torque of the drive motor does not consider the harmonic effect, and uses the first-order transfer function to simulate the hysteresis effect of the motor torque. The output torque calculation formula is:

[0174]

[0175] Among them, T tar is the motor target torque, is the set value;

[0176] t is the simulation time, is the set value;

[0177] τ is the motor torque response time, which is the set value;

[0178] Tmax It is the maximum torque output by the drive motor MG2. This value is determined by the motor external characteristic curve and is a function of the motor speed.

[0179] (5) Driving resistance model

[0180] The resistance encountered by a vehicle during driving mainly includes rolling resistance and air resistance. The formula is as follows:

[0181] F f =fmg (19)

[0182]

[0183] T l =(F f +F w )·r w (twenty one)

[0184] Among them, F f is the rolling resistance, which is the calculated intermediate value;

[0185] F w is the air resistance, which is the intermediate value of the calculation;

[0186] f is the rolling resistance coefficient between the driving wheel and the road surface, which is the set value;

[0187] m is the vehicle mass, which is the set value;

[0188] C D is the air resistance coefficient, which is the set value;

[0189] A is the frontal area of ​​the vehicle, which is the set value;

[0190] u is the vehicle speed, which is the simulation output value;

[0191] 6. Typical working mode

[0192] The response of the powertrain system in the more complex hybrid mode is optimized. In ML or MH gear, the engine and the drive motor (MG2) work together, and the power output of the engine and the motor is optimized and matched through the torque coupling system.

[0193] like Figure 4 As shown in the figure, in this mode, the dynamic response of the power system is more complex, and the model is expanded from the original ten degrees of freedom to eleven degrees of freedom, fully considering the influence of the motor's moment of inertia and torque output on the overall system. The torque and speed between the engine and the motor interact according to a certain coupling relationship to achieve optimal power distribution. Its motion differential equations are shown as follows:

[0194]

[0195] In the formula, J i is the equivalent moment of inertia of the ith component, except for J3, J4 and J body Except for θ, the rest are set values; i is the rotation angle of the ith component, which is the set value;

[0196] k ij is the equivalent torsional stiffness between the i-th component and the j-th component, which is the set value;

[0197] c ij is the equivalent viscous damping between the i-th component and the j-th component, which is the set value;

[0198] The specific meanings of the values ​​of i and j can be found in Table 1 below:

[0199] Table 1 Specific meanings of the values ​​of i and j

[0200] The values ​​of i and j meaning 1 engine 2 Torsional vibration damper 3 Transmission input shaft 4 Transmission output shaft 5 Final reducer 6 Left drive wheel 7 Right drive wheel 8 Half of the vehicle mass is equivalent to the left drive shaft 9 Half of the vehicle mass is equivalent to the right drive shaft 10 Generator MG1 rotor 11 Drive motor MG2 rotor

[0201] Among them, formula (22) is a whole, which is used to solve the engine speed Angular acceleration of the gearbox input shaft and the angular acceleration of the left and right driving wheels

[0202] In summary, the dynamic response of the engine and motor working together in hybrid mode is considered. By expanding the model's degrees of freedom, the influence of the motor's moment of inertia and torque output on the entire system is fully considered, the torque and speed coupling relationship between the engine and motor is established, and the power transmission characteristics under actual working conditions are simulated.

[0203] 7. Time-frequency analysis

[0204] Time-frequency analysis evaluates the vibration characteristics of the vehicle in hybrid drive mode by calculating the second-order angular acceleration of the powertrain.

[0205] (1) As shown in Figure 5(a), the vehicle dynamics simulation is performed based on formula (22) to obtain the engine speed Angular acceleration of the gearbox input shaft and the angular acceleration of the left and right driving wheels And extract the time domain signal in the corresponding simulation time. Among them, the horizontal axis is time, and the vertical axis is speed or angular acceleration.

[0206] (2) As shown in Figure 5(b), the time domain signals of the engine speed and the three angular accelerations are subjected to fast Fourier transform (FFT) to obtain the corresponding frequency domain signals, revealing the distribution characteristics of the angular acceleration and speed at different frequencies. The horizontal axis is the frequency and the vertical axis is the engine speed.

[0207] (3) The frequency spectrum is further subjected to order tracking analysis to extract the main vibration frequency of the system and its relationship with the engine speed. As shown in Figure 5(c), the order is set to 2 and the frequency band is 15 Hz. The second-order angular acceleration of the gearbox input shaft and the left and right drive wheels is analyzed and extracted to identify the key characteristics of the system vibration. The horizontal axis is the engine speed and the vertical axis is the second-order angular acceleration.

[0208] The amplitudes (maximum values) of the three second-order angular accelerations can be obtained from Figure 5(c).

[0209] 8. Select key parameters for multi-objective optimization

[0210] Based on Figure 5(c), the amplitudes of the three second-order angular accelerations can be obtained. Then, the parameters in formula (22) are changed and multiple simulations are performed to identify the parameters that have a greater impact on the amplitudes of the three second-order angular accelerations. These parameters are used as key parameters that have a significant impact on the dynamic response of the system, such as flywheel inertia, torsional vibration damper stiffness, and motor inertia, and are used as optimization variables. The genetic algorithm is used to perform multi-objective optimization of the key parameters of the hybrid powertrain system, aiming to simultaneously reduce the second-order angular acceleration amplitudes of the gearbox input shaft and the left and right drive wheels, and improve the vibration performance of the system.

[0211] (1) Optimization objectives and objective functions

[0212] The optimization goal is to minimize the root mean square value of the second-order angular acceleration of the transmission input shaft and the left and right drive wheels within the engine speed range of 1000rpm to 6000rpm. The specific expression of the objective function is as follows:

[0213]

[0214] Where RMS is the calculated root mean square value;

[0215] Among them, there is a situation where the three objective functions can be minimized at the same time, which corresponds to one solution.

[0216] When the three objective functions do not have the minimum at the same time, there are multiple solutions.

[0217] (2) Optimization variables and constraints

[0218] Based on the above analysis, the following eight key parameters are selected as optimization variables:

[0219] x=[J2,k 12 ,c 12 ,k 23 ,k 34 ,k 56 ,k 57 ,J10 ]

[0220] The constraints of each optimization variable are shown in the following table:

[0221] Table 2 Constraints of optimization variables

[0222] Variable Name Lower limit Upper limit unit <![CDATA[J2]]> -5％ +5％ kg·m2 <![CDATA[k 12 ]]> -20％ +20％ Nm / rad <![CDATA[c 12 ]]> -20％ +20％ Nm <![CDATA[k 23 ]]> -15％ +15％ Nm / rad <![CDATA[k 34 ]]> -15％ +15％ Nm / rad <![CDATA[k 56 ]]> -15％ +15％ Nm / rad <![CDATA[k 57 ]]> -15％ +15％ Nm / rad <![CDATA[J 10 ]]> -5％ +5％ kg·m2

[0223] (3) Genetic algorithm optimization process

[0224] like Figure 6 As shown in Figure 1, the genetic algorithm is used to deal with this multi-objective optimization problem. First, individuals are represented by real number coding, and the fitness function is used to evaluate the quality of each solution. The main steps of the genetic algorithm include selection, crossover, and mutation. The optimization process continuously generates a new generation of individuals until it converges to the Pareto optimal solution. The Pareto optimal solution means that a certain goal cannot be further optimized without compromising any other goals.

[0225] Combining time-frequency analysis and the above-mentioned multi-objective genetic algorithm, the parameters of the hybrid vehicle transmission system are optimized, achieving significant improvements in the system vibration characteristics. By optimizing key parameters such as flywheel inertia, torsional vibration damper stiffness, and motor inertia, the second-order angular acceleration of the gearbox input shaft and the drive wheel is effectively reduced, thereby reducing the vibration amplitude and noise, and improving the NVH performance and driving comfort of the vehicle. At the same time, the optimization process ensures the reasonable range and constraints of various parameters, and finally obtains a Pareto optimal solution that meets the requirements of multi-objective optimization, providing effective technical support for the design of hybrid powertrain systems. The optimization results can be used as design parameters to help achieve a more efficient, low-vibration, and low-noise powertrain system.

[0226] In addition, after the optimization process is completed, simulation verification is carried out to verify the effect of the optimized parameters on the system vibration behavior, torsional vibration reduction effect and overall dynamic performance improvement, ensuring that the optimization results meet the performance requirements under different working conditions, achieving the purpose of reducing vibration and noise and improving system reliability.

[0227] Step A: First, adjust the values ​​of key parameters within the constraints and generate a new parameter array containing 8 values.

[0228] Step B: Then, substitute the array into formula (22) to perform system dynamics simulation, and finally perform time-frequency analysis, and extract the relationship curve between the engine speed and the second-order angular acceleration corresponding to each simulation.

[0229] According to the engine speed-second-order angular acceleration curve of each simulation, its root mean square value is calculated to quantify the dynamic characteristics of the system vibration.

[0230] Step C: After traversing the root mean square values ​​of all simulation results, select the parameter array that minimizes the two objective functions at the same time, and use it as the optimal solution for multi-objective optimization.

[0231] Among them, step A and step C are completed by genetic algorithm.

[0232] The above is only a preferred embodiment of the present invention and does not limit the present invention in any way. Any technician in the relevant technical field, without departing from the scope of the technical solution of the present invention, makes any form of equivalent replacement or modification to the technical solution and technical content disclosed in the present invention, which does not depart from the content of the technical solution of the present invention and still falls within the protection scope of the present invention.

Claims

1. A hybrid powertrain system parameter optimization method based on time-frequency analysis, characterized in that: The following steps are involved: Step 1: Establish a dynamic model of the hybrid powertrain system, which includes: An engine and a flywheel, wherein the flywheel is connected to an output shaft of the engine; A torsional vibration damper, wherein the primary mass is fixedly connected to the flywheel and the secondary mass is connected to the input shaft of the gearbox; a vibration damping spring is arranged between the secondary mass and the primary mass; The gearbox comprises an output shaft and an input shaft; The main reducer-differential includes two output shafts, each of which is connected to a half shaft, and the half shaft is equipped with a driving wheel; the main reducer-differential is connected to the gearbox and the driving motor MG2 through a gear set; Step 2: Establish the total transient torque model T of the engine e ; Step 3: Establish the transmission torque model T of the torsional vibration damper d ; Step 4, calculate the total rotational inertia J3 of the gears on the input shaft of the gearbox and the total rotational inertia J4 of the gears on the output shaft; Step 5: Calculate the moment of inertia J of the vehicle body body , output torque T of drive motor MG2 mg2 , resistance torque T1; Step 6: Create a file containing T e 、T d , J3, J4, J body 、T mg2 , the differential equations of motion of the powertrain in hybrid mode of T1; Step 7: Perform time-frequency analysis based on the differential equations of motion; Step 8: Select several key parameters as optimization variables; Step 9: Set constraints and objective function; Step 10: Execute the genetic algorithm to obtain the optimal solution.

2. The hybrid powertrain system parameter optimization method based on time-frequency analysis according to claim 1, characterized in that: The specific expressions of J3 and J4 in step 4 are: J a1 is the moment of inertia of the ML gear driving gear (1); J a2 is the moment of inertia of the ML gear driven gear (2); J a3 is the moment of inertia of the MH gear driving gear (3); J a4 is the moment of inertia of the MH gear driven gear (4); J a5 is the moment of inertia of the MG1 motor driving gear (5); J a6 is the moment of inertia of the driven gear (6) of the MG1 motor; i ml It is the transmission ratio of ML gear; i mh It is the transmission ratio of MH gear; i mg1 is the gear ratio of the generator MG1.

3. The hybrid powertrain system parameter optimization method based on time-frequency analysis according to claim 1, characterized in that: Step 5 J body 、T mg2 The specific expressions of and T1 are: T l =(F f +F w )·r w F f =fmg m is the vehicle mass, which is the set value; r w is the radius of the driving wheel, is the set value; i g and i0 are the transmission ratios of the gearbox and the final reducer respectively; t is the simulation time, is the set value; T tar is the target torque of the motor, is the set value; τ is the motor torque response time; T max The maximum torque output by the drive motor MG2; F f is the rolling resistance; F w is air resistance; f is the rolling resistance coefficient between the driving wheel and the road surface; m is the vehicle mass; C D is the air resistance coefficient; A is the windward area of ​​the vehicle; u is the vehicle speed.

4. The hybrid powertrain system parameter optimization method based on time-frequency analysis according to claim 1, characterized in that: The differential equations of motion in step 6 are: J i is the equivalent moment of inertia of the ith component; θ i is the rotation angle of the i-th component; k ij is the equivalent torsional stiffness between the i-th component and the j-th component; c ij is the equivalent viscous damping between the i-th component and the j-th component; Among them, the value set of i and j includes the engine torsion reducer, the gearbox input shaft, the gearbox output shaft, the main reducer, the left driving force, the right driving wheel, the generator MG1 rotor, and the drive motor MG2 rotor.

5. The hybrid powertrain system parameter optimization method based on time-frequency analysis according to claim 1, characterized in that: The optimization variables in step 8 are: x=[J2,k 12 ,c 12 ,k 23 ,k 34 ,k 56 ,k 57 ,J 10 ] J2-equivalent moment of inertia of the engine; k 12 -Equivalent torsional stiffness between the engine and the torsional vibrator; c 12 -Equivalent viscous stiffness between the engine and the torsional vibrator; k 23 -Equivalent torsional stiffness between the torsional vibrator and the gearbox input shaft; k 34 -Equivalent torsional stiffness between the gearbox input shaft and the gearbox output shaft; k 56 -Equivalent torsional stiffness between the final reducer and the left drive wheel; k 57 -Equivalent torsional stiffness between the final reducer and the right drive wheel; J 10 -Equivalent moment of inertia of the rotor of the generator MG1.

6. The hybrid powertrain system parameter optimization method based on time-frequency analysis according to claim 1, characterized in that: The objective function in step 9 is to minimize the root mean square value of the second-order angular acceleration of the transmission input shaft, the left drive wheel, and the right drive wheel.

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