JPH0933336A - Vibration analysis method for system having nonlinear element, torsional vibration analysis method for vehicle driving system, torsional angle/torque characteristic decision method for multistage clutch disk, and multistage clutch disk - Google Patents

Abstract

PROBLEM TO BE SOLVED: To operate the vibration analysis of the whole of a system having nonlinear elements surely in a short time with a simple structure by dividing the system into plural sub-systems taking nonlinear elements as boundaries, forming simultaneous motion equations on each sub-system so as to input the characteristics of the nonlinear elements, and obtaining a solution. SOLUTION: For instance, a vibration system is divided into two sub-systems taking a dry single plate clutch 2 of a 6×4 large-size truck as a boundary, from the clutch 2 to a vehicle body 16 is assigned as a sub-system 1, and from the clutch 2 to an engine 1 is assigned as a sub-system 2. On the sub-system 1, an equation of motion expressed by the model coordinate of the natural vibration of a drive system and a boundary freedom obtained by the static contraction of Guyan is formed, and on the system 2, after the equation of motion is set up considering the engine 1, the disk of the multistage clutch 2, and a torque change as external force, simultaneous motion equations are formed out of two motion equations so as to input the characteristics of the multistage clutch disk to obtain a solution. Thereby, torsional vibration analysis on the whole of the driving system can be conducted without conducting design based on experiment, experience, or the like.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a vibration analysis method for a system having a non-linear element, a torsional vibration analysis method for a vehicle drive system, and a multistage clutch disc, which are suitable for designing a drive shaft system of a truck. And a multi-stage clutch disc.

[0002]

2. Description of the Related Art Low vibration and low noise of a drive system of a truck are important themes from the point of view of commerciality. For the torsional vibration of the vibration system, its resonance frequency should be released outside the practical vehicle speed range. Attempts have been made to change the spring constant of the clutch disc, and to suppress the amplitude at resonance below the allowable level by optimizing the hysteresis of the clutch disc when it cannot escape.

It is an effective method to predict design parameters for low vibration and low noise by simulation and confirm them by experiments. The variation of the drive system of the truck is shown in FIG. As shown in FIG.
Rear two-axis drive as shown in (b), FIG. 2 (c), FIG.
In a complex system such as all-wheel drive as shown in (d), when performing non-linear response analysis including multi-stage torsional characteristics of the clutch disc, there are many degrees of freedom, and therefore a very large amount of calculation time is required. I need.

[0004]

That is, in the case of the conventional means, a huge amount of CPU calculation time is required for the computer, and in terms of phenomenon analysis, the model is too large and difficult to analyze, resulting in poor efficiency. Therefore, when trying to simplify the model within a practical range, there is a problem that variations peculiar to the truck cannot be incorporated in the model.

For example, a rear two-axle vehicle is approximated by a rear single-axle vehicle, and there is a problem that it cannot be modeled as a rear two-axle vehicle. The present invention was devised in view of such a problem, the system is divided into several subsystems, the partial structure synthesis method is applied, and only the necessary vibration mode is extracted as a linear inside each subsystem. , Non-linear elements such as multi-stage characteristics of clutch discs and backlash of gears are taken into consideration when connecting subsystems to form the overall system, so that the degree of freedom of the overall system is greatly reduced and the calculation time can be shortened. It is an object of the present invention to provide a vibration analysis method for a system having a non-linear element, a torsion vibration analysis method for a vehicle drive system, and a torsion angle / torque characteristic determination method for a multi-stage clutch disc.

Another object of the present invention is to provide a multi-stage clutch disk which can reduce the calculation time required for its design.

[0007]

Therefore, the vibration analysis method for a system having a non-linear element of the present invention uses the non-linear element as a boundary when performing a vibration analysis for a system having one or more non-linear elements. , The system is divided into a plurality of subsystems, the equations of motion for each subsystem are established, and then the equations of motion for the above subsystems are made simultaneous,
The characteristics of the non-linear element are input and a solution is obtained, whereby vibration analysis of the entire system is performed (claim 1).

A vibration analysis method for a system having a nonlinear element according to the present invention is the method according to claim 1,
Input the characteristics of the non-linear element so that the linear element comes to either the right side or the left side and the non-linear element comes to the other side of the right side or the left side, and the equations of motion for the respective subsystems are made simultaneous. Then, the vibration analysis of the entire system is performed by obtaining a solution (claim 2).

Further, the vibration analysis method for a system having a non-linear element according to the present invention is the method according to claim 1, which is calculated according to the order of the vibration to be obtained when the equation of motion is established for each subsystem. It is characterized in that the error information is added. A vibration analysis method for a system having a non-linear element according to the present invention is the method according to claim 2 or 3, wherein numerical integration of numerical integration is performed according to the order of vibration obtained when solving the equation of motion for each subsystem. It is characterized in that the solution is obtained while selecting the size (claim 4).

The torsional vibration analysis method for a vehicle drive system according to the present invention uses a clutch as a boundary when performing a torsional vibration analysis for a vehicle drive system having a clutch disc as a non-linear element. A first subsystem from the clutch to the end element of the drive system;
After division into a second system from the engine to the clutch, with respect to the first system, the modal coordinates of the natural vibration of the drive system and the boundary degrees of freedom determined by Guyan's static contraction are used. The equation of motion shown is created, and for the second subsystem, an equation of motion is created in consideration of the engine, the clutch disc, and torque fluctuations as external force.
The above equations of motion for the first subsystem and the second subsystem are made simultaneous, and the characteristics of the clutch disk are input,
A torsional vibration analysis is performed for the entire drive system by obtaining a solution (claim 5).

Further, the torsional vibration analysis method for a vehicle drive system according to the present invention is the method according to claim 5, wherein a linear element is provided on one of the right side and the left side and the clutch is provided on the other side of the right side or the left side. The drive system is obtained by making simultaneous the equations of motion for the first subsystem and the second subsystem so that the torque information of the disk comes and inputting the characteristics of the clutch disk to obtain a solution. It is characterized in that a torsional vibration analysis is performed for the whole (Claim 6).

Further, a torsional vibration analysis method for a vehicle drive system according to the present invention is the method according to claim 5, wherein the first subsystem is:

[0013]

(Equation 3)

The error information [Kr] represented by is replaced with the rigidity matrix inside the first subsystem, and the equation of motion combining the constraint modal method with the mode compensation is constructed (claim). 7). The method for determining the twist angle / torque characteristics of the multi-stage clutch disc of the present invention is a method of determining the twist angle / multi-stage clutch disc as a non-linear element provided in a vehicle drive system at low speed / light load and high speed / high load. When obtaining the torque characteristics,
In order to keep the resonance speed below the practical vehicle speed range,
The spring constant defining the low speed / light load region in the torque characteristic is determined, and the torsional vibration analysis method using the partial structure synthesis method, that is, the clutch is used as a boundary to drive the drive system from the clutch It is divided into a first subsystem that reaches the end element of the system and a second subsystem that extends from the engine to the clutch, and then, for the first subsystem, modal coordinates of the natural vibration of the drive system are set. The equation of motion expressed by the boundary degree of freedom obtained by Guyan's static contraction is constructed and
For the subsystem, the equations of motion considering the engine, the multi-stage clutch disc, and torque fluctuation as external force are established, and then the equations of motion for the first subsystem and the second subsystem are made simultaneous. Then, by inputting the characteristics of the multi-stage clutch disc and obtaining a solution, the upper limit torsion angle in the low speed / light load region is determined by using the torsional vibration analysis method for analyzing the torsional vibration of the entire drive system. In addition, the spring constant that defines the high-speed / high-load region in the torsion angle / torque characteristics is determined so that the resonance speed falls below the practical vehicle speed range, and the above partial structure synthesis method is used. The upper limit torsion angle in the high-speed / high-load region is determined by using the torsional vibration analysis method (claim 8).

Further, the torsion angle / torque characteristic determining method for the multi-stage clutch disc of the present invention is applied to the multi-stage clutch disc as a non-linear element provided in the vehicle drive system at low speed / light load and high speed / high load. Torsion angle
When obtaining the torque characteristics, the spring constant that defines the low-speed / light-load area in the torsion angle / torque characteristics was determined so that the resonance speed was below the practical vehicle speed range, and the partial structure synthesis method was used. A first torsional vibration analysis method, that is, with the clutch serving as a boundary, the drive system includes a first subsystem that extends from the clutch to an end element of the drive system and a second subsystem that connects the engine to the clutch. Then, the first system is divided into and the equation of motion represented by the modal coordinates of the natural vibration of the drive system and the boundary degrees of freedom obtained by Guyan's static contraction is created, and For the second subsystem, a motion equation considering the engine, the multi-stage clutch disc, and torque fluctuation as an external force is created, and then the equations of motion for the first subsystem and the second subsystem are simultaneously expressed. Let the multi-stage Enter the characteristics of the Tchidisuku,
By determining the solution, the first torsional vibration analysis method that analyzes the torsional vibration of the entire drive system is used to determine the upper limit torsion angle in the low-speed / light-load region, and the resonance rotational speed is practically used. A spring constant that defines the high-speed / high-load region in the torsion angle / torque characteristics is determined so as to be equal to or lower than the vehicle speed range, and a second torsional vibration analysis method using a partial structure synthesis method, that is, With the clutch as a boundary, the drive system is divided into a first system from the clutch to the end element of the drive system and a second system from the engine to the clutch, and then the first system For the system of, the equation of motion expressed by the modal coordinates of the natural vibration of the drive system and the boundary degrees of freedom obtained by Guyan's static reduction,

[0016]

(Equation 4)

The error information [Kr] represented by is replaced with the stiffness matrix inside the first subsystem to combine the modal compensation with the modal compensation and the second subsystem with respect to the second subsystem. The engine, the multi-stage clutch disc, and the equation of motion considering the torque fluctuation as the external force are created, and then the equations of motion for the first subsystem and the second subsystem are made simultaneous to form the multi-stage clutch disc. By determining the solution by inputting the characteristics of the above, a second torsional vibration analysis method for performing torsional vibration analysis for the entire drive system is used to determine the upper limit torsion angle in the high speed / high load region. (Claim 9).

Furthermore, the multi-stage clutch disc of the present invention is a multi-stage clutch disc provided in a vehicle drive system, in which the upper limit of the torsion angle / torque characteristic in the high-speed / high-load region is that of the high-speed / high-load condition. The value is set to a value corresponding to about 1.7 times the engine average shaft torque, and the lower limit of the torsion angle in the high speed / high load region in the torsion angle / torque characteristic is the engine average shaft at the high speed / high load. It is characterized in that it is set to a value corresponding to around 0.5 times the torque (claim 10).

The multi-stage clutch disc of the present invention is a multi-stage clutch disc provided in a vehicle drive system, wherein the upper limit torsion angle in the high-speed / high-load region of the torsion angle / torque characteristics is at high speed / high load. The value is set to a value corresponding to around 1.7 times the engine average shaft torque, and the lower limit twist angle in the high speed / high load region and the upper limit twist angle in the low speed / light load region in the twist angle / torque characteristic are respectively set. The value is set to a value corresponding to around 0.5 times the engine average shaft torque at the time of high speed / high load, and the lower limit of the twist angle / torque characteristics in the low speed / light load region is It is characterized in that it is set to a value corresponding to around 0.3 times the engine average shaft torque under high load (claim 11).

[0020]

Embodiments of the present invention will be described below with reference to the drawings. 1 to 18 show an embodiment of the present invention, and FIG. 1 is a schematic diagram showing a simulation model of a drive system in a 6 × 4 track, FIG.
4D are schematic diagrams showing each model of the drive system, FIG. 3 is a schematic diagram showing simulation results, and FIGS. 4A to 4C are schematic graphs showing hysteresis characteristics of the clutch device.
5A to 5C are graphs showing the response characteristics of the drive system in the simulation, FIGS. 6 to 8 are graphs showing the frequency response characteristics of the drive system in the simulation, and FIG. 9 is a frequency response experiment of the drive system with respect to the simulation results. FIG. 10 is a graph showing the frequency response characteristic of the drive system in the simulation, FIG. 11 is a graph showing the noise level characteristic of the drive system, and FIGS. 12 and 13 are graphs showing the hysteresis characteristic of the clutch device of this embodiment. ,
FIG. 14 is a diagram for explaining a CPU calculation time characteristic in a simulation, FIG. 15 is a diagram for explaining a CPU time calculation characteristic in comparison with a conventional method as an embodiment of the present invention, and FIG. 16 is an example of a clutch device. FIG. 17 is a partially cutaway front view showing an example of the clutch device with its front structure partially broken, and FIG. 18 is a graph showing hysteresis characteristics of the example of the clutch device.

First, a simulation should be carried out using the partial structure synthesis method BBA of the present invention to perform design.
The dry single-plate clutch attached to the V-type 8-cylinder 4-cycle diesel engine or the like of the four heavy-duty trucks is configured as follows, for example. That is, as shown in FIGS. 16 and 17, the spline hub 101 is provided so as to be capable of spline fitting to a main shaft (not shown), and the annular flange portion of the spline hub 101 is provided with a fan-shaped notch. , Fan-shaped windows are provided.

The drive clutch plates 102, 10
3 is provided so as to face each of both end surfaces of the annular flange portion of the spline hub 101, and the drive clutch plates 102 and 103 are provided with fan-shaped notches, fan-shaped windows, and pins 107. Are connected via. Then, the sub plates 104a, 1
04b is provided inside the drive clutch plates 102 and 103, is coupled by the pin 106, is loosely fitted to the outer periphery of the spline hub 101, and has a fan-shaped window and a fan-shaped notch on the outer peripheral side. .

Further, the bush 105 is the spline hub 1.
01 is loosely fitted to the drive clutch plate 102. The pin 107 connects the drive clutch plates 102 and 103, and the sub-plate 104a,
It is arranged so as to engage with the fan-shaped notch of 104b. Further, the fan-shaped notches of the sub-plates 104a and 104b are formed and arranged so as to have a rotation permissible width of θ2 in the counterclockwise direction and θ2 ′ in the clockwise direction with respect to the pin 107.

Further, the fan-shaped notch of the spline hub 101 with which the pin 106 engages is formed so as to have a rotation permissible width of θ1 in the counterclockwise direction and θ1 ′ in the clockwise direction with respect to the pin 106. It is set up. Then, the non-metal friction washers 108a, 108b are
The flange portion of the spline hub 1 and the sub plate 104a,
The friction washer 10 is interposed between the friction washer 10 and 104b.
8a, 108b and the spline hub 101 are configured to have a small friction coefficient.

The non-metallic friction washer 1
09a is interposed between the sub plate 104a and the drive clutch plate 102, and the friction washer 10
9b is interposed between the sub plate 104b and the drive clutch plate 103. Here, friction coefficients between the friction washer 109a and the drive clutch plate 102 and between the friction washer 109b and the drive clutch plate 103 are set to be large.

Further, the wave spring 111 is interposed between the friction washer 109b and the sub-plate 104b, and is prevented from rotating by the sub-plate 104b. The friction plate 110 is provided between the wave spring 111 and the friction washer 108b, and the sub plate 1
It is stopped by 04b.

A first-stage spring 112 is provided, and the spring seats 116 and 117 are used to form the first-stage spring 11.
2, the spring seat 116 receives the edge of the fan window provided on the spline hub 101 and the sub plate 10 at the position shown in the figure.
It is in contact with the edges of the fan windows provided at 4a and 104b.
The bottom portions of the spring seats 116 and 117 are in contact with the protrusions of the fan window provided on the flange portion of the spline hub 101 that face in the opening direction.

The spring seats 116 and 117 have fitting portions (not shown) with the sub-plates 104a and 104b, and the distance from the center of the spline hub 101 is kept constant. A structure similar to that of the first-stage spring 112 is provided symmetrically with respect to the center of the spline hub 101.

The fan window 125 is used for the drive clutch plate 10.
2, 103 and the sub plates 104a and 104b. The spring seat 118 is the sub plate 10
4a, 104b and a fitting portion (not shown) are provided so that the distance from the center is kept constant.
The second stage outer spring 114 has its inner diameter guided by the outer diameter of the spring seat 118.
It is fitted into the fan window provided on 04b.

The second stage inner spring 115 is the spring seat 1
Both ends are supported by 18. The second step spring 113 is 1
The step spring 112 is concentrically provided on the outer side of the step spring 112 and is inserted into a fan window 125 provided on the sub plates 104a and 104b. The fan window 126 penetrates the drive clutch plates 102, 103 and the sub-plates 104a, 104b provided at positions perpendicular to the fan window 125.

The notch edge 119 is provided on the collar-shaped portion of the spline hub 101 to abut the pin 106, and 121 is the portion of the fan-shaped notch pin 107 provided on the collar of the spline hub 101 in the counterclockwise rotation direction. A notch edge, 120 is a notch edge in the clockwise direction of the portion where the pin 106 of the fan notch provided on the flange of the spline hub 101 abuts, and 122 is a counterclockwise fan-notch provided on the sub plates 104a and 104b. A notch edge in the rotational direction, and 123 is a notch edge in the clockwise direction of the pin 107 portion of the fan-shaped notch provided on the flange of the spline hub 101.

With this structure, when the drive clutch plates 102 and 103 are pressed against the clutch surface of the flywheel by a pressure plate (not shown), the rotation of the flywheel is transmitted and the drive clutch plates 102 and 1 are transmitted.
03 turns counterclockwise. Drive clutch plate 103
Is locked together with the drive clutch plate 102 by a pin 107, so that both rotate together.

A sub-plate 104a between the drive clutch plate 102 and the drive clutch plate 103,
104b, friction washers 108a, 108b,
The friction washers 109a and 109b, the friction plate 110, and the collar portion of the spline hub 101 are
Since the surface pressure is applied by the wave spring 111, torque is transmitted by friction with each other, but since friction between the friction washers 108a, 108b and the spline hub 101 is small, slippage occurs at this portion and the sub plates 104a, 104b, Friction washer 1
08a, 108b, friction washers 109a, 1
09b, the wave spring 111, and the friction plate 110 rotate integrally with the drive clutch plates 102 and 103.

When the sub-plates 104a and 104b rotate counterclockwise, the first-stage spring 112 is moved to the spline hub 101.
The torque is transmitted to the spline hub 101 because it is compressed between At this time, the torque fluctuation is alleviated by the deflection of the first-stage spring 112. The torque is also transmitted by the sliding friction washers 108a and 108b as described above, and this action provides the hysteresis characteristic of the first stage as shown in FIG.

Then, the sub plates 104a and 104b
When the counterclockwise rotation of the spline hub 101 with respect to the spline hub 101 reaches θ1, the pin 106 comes into contact with the notch edge 119 of the fan-shaped notch of the spline hub 101, and the sub plate 1
04a and 104b are integrated with the spline hub 101. When the drive clutch plate 102 further rotates, the friction washers 109a, 1 having the next largest friction coefficient are generated.
09b and drive clutch plates 102 and 103, respectively
Sliding between them, the drive clutch plates 102, 103 and the bush 105 rotate integrally.

At this time, the second step spring 113, the second step outer spring 114, and the second step inner spring 115 are connected to the spline hub 101.
The torque is transmitted to the spline hub 101 because it is compressed between On the other hand, the friction washers 109a,
Torque is also transmitted on the sliding surface of 109b.

When the drive clutch plates 102 and 103 further rotate by θ2, the pin 107 rotating together abuts on the notched edge 122 of the sub plates 104a and 104b, so that the rotation occurs between the sub plates 104a and 104b and the pin 106. The torque is directly transmitted to the spline hub 101 via the, and the relative rotation is lost, so that a large torque is transmitted.

At this time, the sub plates 104a, 104
b is fixed integrally with the flange portion of the spline hub 101, and only the drive clutch plates 102 and 103, the spring seats 116 and 117, and the pin 107 rotate integrally. At that time, the friction washer 109
The second stage hysteresis is generated at the portions a and 109b.

It is necessary to appropriately set the hysteresis characteristics of the first and second stages as described above in order to reduce the noise related to the clutch disc, and further, if necessary, the third and fourth stages. , But the hysteresis is set by the following means. That is, in order to design a clutch device and the like for a 6 × 4 large truck equipped with a V-type 8-cylinder 4-cycle diesel engine and a dry single plate clutch, the following analysis is performed.

In this analysis, the constrained modal method, which is one of the substructure synthesis methods, and the mode compensation method are used. The mode compensation method is a method of combining mode compensation with a constraint modal method, and the operation of the partial structure synthesis method BBA is developed in the simulation model shown in FIG. 1, for example.

First, the drive system is divided into two subsystems with the clutch serving as a boundary. Subsystem 1 is from the clutch to the vehicle body and subsystem 2 is from the clutch to the engine. The subsystem 1 includes a clutch (clutch hub) 2, transmissions 3, 4, propeller shafts 5, 6, propeller shaft 7, through shaft 8, propeller shaft 9, propeller shaft 9,
10, differential 11, wheel 12, car body 1
3, the differential 14, the wheel 15, and the vehicle body 16 are provided.

Further, the subsystem 2 is constructed so as to include the engine 1 and the clutch (clutch hub) 2. Here, when the equation of motion of the subsystem 1 is divided into the dependent degree of freedom Xf inside the boundary and the independent degree of freedom Xb on the boundary,

[0043]

(Equation 5)

Here, M is the mass, f is the degree of freedom inside the subsystem 1, b is the boundary degree of freedom, X is the displacement, X to dot is the acceleration, K is the stiffness spring coefficient, "0" is the internal force, and Tc. Is the torque of the clutch disc. When the Guyan static contraction is applied to the above equation, [Xf ′] = − [Kff ⁻¹ Kfb] Xb (2), and the equation excluding the mass M is obtained.

On the other hand, the eigenvalue problem when the boundary is fixed − ([Kff] −ω ² [Mff]) [Xf ′] =

[0] ... Assuming that the constraint mode obtained by solving (3) is [φ], the physical degree of freedom Xf is converted into modal coordinates ξ by the following equation. [Xf ″] = [φ] ξ (4) According to the constraint mode method, the degree of freedom inside the boundary is calculated by the following equation: [Xf] = [Xf ′] + [Xf ″] = [φ −Kff ⁻¹ Kfb] [ξXb] ^t (5), the following equation is obtained.

[0046]

(Equation 6)

Then, by substituting the equation (6) into the equation (1) and multiplying by [T] ^t from the previous, the coordinate transformation of the equation (1) is performed.

[0048]

(Equation 7)

Here,

[0050]

(Equation 8)

Next, the calculation for the subsystem 2 will be described.

[0052]

[Equation 9]

Here, T _E0 represents the average shaft torque of the engine, and ΔT _E represents the torque fluctuation. The amplitude of the torque fluctuation is given so that the rotation fluctuation of the flywheel becomes almost the same value as the measured value, and the frequency of the torque fluctuation is simplified for the sake of explosion 1
Only the next component is calculated. The average shaft torque is a measured value.

Next, the calculation for the entire system will be described. First, in order to calculate the nonlinear transient response by inputting the multistage characteristics of the clutch disc, the following expansion is performed.
From equations (7) and (8),

[0055]

(Equation 10)

Here, TE-TC on the right side is the torque acting on the subsystem 1 side, and TC is the torque acting on the subsystem 2 side. The multi-stage characteristic of the clutch disc is given by the following equation (11).

[0057]

[Equation 11]

After solving the equation (10), it is converted to the physical coordinate system by the following equation (12).

[0059]

(Equation 12)

The attenuation is calculated by using Rayleigh damping according to the following equation.

[0061]

(Equation 13)

There is a relationship of ζr = (α / ωr + βωr) / 2 (14) between the proportional constants α and β and the modal damping ratio ζr (reference document; Nagamatsu and Okuma: partial structure synthesis method). , Baifukan, 1991), so give the measured value to ζr,
It is calculated by back-calculating α and β. Further, ωr is the angular frequency of the target vibration.

On the other hand, in the case of combining modal compensation with the constrained modal method,

[0064]

[Equation 14]

In addition, the upper right component of the transformation matrix [T] of the equation (6) is changed from −Kff ⁻¹ · Kfb to −Kr ·
It may be replaced with Kfb. Here, the first term on the right side [Kf
f] is the internal stiffness matrix, the second term on the right side [Mf
f] indicates the internal mass matrix, which is the third part on the right side.
The term [φi] represents a column vector, and the third term on the right side <φi> represents a row vector.

Further, ωi represents the i-th order natural angular frequency, and ωc is a parameter of a certain constant natural angular frequency, and the angular frequency of the target vibration may be given. (6)
Formulation The following formulation proceeds in the same way as the constrained modal method. By the way, the application of the partial structure synthesis method BBA will be described. First, the calculation is performed for the torsional fourth-order vibration at the time of high-speed / high-load traveling (Reference: Automotive Handbook (Basic / Theoretical Edition), JSME, 1990. ).

This is because it is not necessary to consider a non-linear element such as the backlash of the gear under a high load, and the calculation is easy to handle. Then, assuming the actual operating range of the clutch disc, the natural vibration is obtained by linear approximation (reference: K. Suzuki, Y. Tozawa: Influence of Powertrain Tors
ional Rigidityon NVH of 6 × 4 Trucks, SAE Paper 9224
82).

Then, many vibration modes are required, but the object is only the torsional fourth-order vibration. Here, the natural vibration analysis by the partial structure synthesizing method BBA will be compared with a conventional method not using the partial structure synthesizing method BBA. As a conventional method, a general-purpose code NASTRAN (reference document; NA
We are comparing the natural frequency and vibration mode when using STRAN User's Manual). The result of the comparison is obtained as shown in FIG.

The numerals 1 to 16 in the figure correspond to the elements of the same numeral in FIG. 1, and the displacements of the corresponding elements are shown at the positions of the respective numerals. The vibration characteristics are represented by connecting with a broken line. Here, [φ] = (φ1 to φ4), and [φ] in the figure
= (Φ4), [φ] = (φ4 + compensation), indicates the case of the conventional NASTRAN,
[Φ] = (φ1 to φ4) means the case of selecting the 1st to 4th order vibrations from the constraint mode [φ] of the equation (4).

Further, the solid line shows the characteristics in the case of, and the vibration modes have overlapping characteristics.
In addition, the natural frequency characteristic in the case of, is 32.0
It is the characteristic in the vibration of Hz, and 32.0 Hz is 500.
It corresponds to an idle speed of around rpm. on the other hand,
The characteristic of the broken line indicates the characteristic with respect to the vibration of 29.6 Hz.

That is, in the case of [φ] = (φ1 to φ4), the torsional fourth-order vibration matches the natural frequency and the vibration mode by 3 to 4 digits in effective digits as compared with the conventional method. Also, [φ] = (φ
Even in the case of 4), if compensation is performed (in this case, [φ] =
(Φ4 + compensation). ), The natural frequency and the vibration mode match with the conventional method by 3 to 4 significant digits.

As a result, it is considered that the natural frequency and the accuracy of the vibration mode (effective digits) can be considered to be almost the same.
From the constraint mode [φ] of equation (4), the required degrees of freedom φm to φn
Mode is selected, the modes φm to φn are set to the whole system ((10)
It is difficult to find out which vibration of the equation) the vibration corresponds to, but in the present embodiment, it corresponds to the order of the frequency.

By the way, the nonlinear response analysis by the partial structure synthesis method BBA will be explained. First, the modal parameter of each subsystem is calculated, and the calculated modal parameter is used as the general-purpose code ACSL.
(Reference: ACSL User's Guide / Reference Manual), and the analysis is performed as a system for obtaining the time response. The clutch disc under test is a one-stage type (Fig. 4
(A)), 4-stage type (Fig. 4 (b)), 3-stage type (Fig. 4)
(C)), three types of clutches A, B, C respectively
Call.

The hysteresis H1a of the clutch A and the hysteresis H2c and H3c of the clutch C are twice the hysteresis H2b to H4b of the clutch B. FIG.
In (a) to (c), the horizontal axis represents the twist angle θ of the clutch disk and the vertical axis represents the transmission torque Tc, which shows the hysteresis characteristics of the clutch disk, and the spring constants classified by θ1 to θ4 are shown. The characteristics corresponding to k1 to k4 are shown.

First, the clutch A was tested, and the torsional fourth-order natural angular frequency ω0 calculated by the spring constant was obtained, and ω0
The frequency response before and after is set as the excitation frequency ω and the time response is compared with the conventional method. First, the clutch A was tested with the theoretical solution in the one-degree-of-freedom system (reference; CYRIL M.HARRIS; SHOCK AND
VIBRATION HANDBOOK, THIRD EDITION, McGRAW-HILL BOOK
COMPANY), and it is easy to conduct theoretical studies.

In this case, since damping is not taken into consideration other than the clutch, the response is in a divergent state in which the amplitude increases with time when the natural frequency and the excitation frequency match (FIG. 5 (b )). Here, in FIGS. 5A to 5C, the horizontal axis represents time (seconds), and the vertical axis represents rotation fluctuation.
When 0 = 0.98, when ω / ω0 = 1, ω / ω0 =
The time-dependent change of the rotational fluctuation in each of the cases of 1.02 is shown.

[Φ] = (φ1 to φ4), [φ] = (φ4
In the case of (+ compensation), the effective digits are 2 to 3 digits in agreement with the conventional method at any excitation frequency, but FIG. 5 shows [φ] = (φ1
~ Φ4) shows the response. On the other hand, [φ] =
At (φ4), the natural frequency is 29.6 Hz and the conventional method 32
Since it shifts to the low frequency side by 10% or more with respect to Hz (see FIG. 3), correspondingly, ω / ω0 = 0.98 to 1.
The response value at 02 also differs from the conventional method from the first significant digit.

From this result, [φ] is calculated for the fourth vibration.
= (Φ1 to φ4) or [φ] = (φ4 + compensation) is used. Next, the clutch B is tested and compared.
In the case of [φ] = (φ1 to φ4), as in the case of the clutch A, the three digits of the effective digits match those of the conventional method (see FIG. 6).

In the figure, Ne is the engine rotational speed, N0 is the idle rotational speed, ΔN0 is the rotational fluctuation (actually measured value) of the flywheel during idling, and ΔNp is the rotational fluctuation of the differential pinion. Also, a comparison between the calculation and the experiment is shown in FIG.
As can be seen from FIG. 7, in the calculation, the hysteresis ratio of the clutch disk is added to the damping ratio obtained in the experiment, but the resonance rotational speed and the amplitude at resonance almost correspond to each other.

Next, the torsional vibration characteristics of the clutch disc and the fourth torsional vibration will be described. The clutch C has double the hysteresis of the clutch B in the actual operating range. A decrease of about 25% in the amplitude at resonance was predicted due to the increase in hysteresis (FIG. 8), and the effect was confirmed in the experiment (FIG. 9).

In both clutches B and C, the resonance occurrence region is outside the practical vehicle speed range, and the feeling is good.
Next, the result of analyzing the torsional fourth-order vibration during low speed / light load running will be described. In the case of light load, the backlash of the gears was ignored as in the case of high load.

The clutch B has a light load stage added to the clutch C. FIG. 10 shows a prediction of the torsional vibration reduction effect of the clutch B with respect to the clutch C at low speed and light load, and FIG. 11 shows a confirmation result in an actual vehicle. FIG.
It can be seen that 1 is the noise level in the resonance range, which is greatly reduced by the clutch B.

In FIG. 10, [φ] = (φ1 to φ4)
Is the result in the case of. By the way, when comparing the calculation time in the computer required for the analysis, it becomes as shown in FIG. 14, and it is considered as follows. That is, as a result of comparing the CPU time when the partial structure synthesis method BBA is used with the conventional method, 12% when [φ] = (φ1 to φ4) and [φ] = (φ4 + compensation) when [φ] = (φ1 to φ4) In the case of 48
%, And the calculation time is significantly shortened.

In the calculation of this time, which is a comparison target, the same integral step size was used. However, since higher-order vibrations are eliminated, the step size may be increased. In that case, the step size is further shortened. it can. In fact, as shown in FIG. 15, the calculation time is reduced to 1/10 or less as compared with the conventional method by increasing the step size by removing the higher-order vibration. The horizontal axis NST in FIG. 15 is a parameter of the integration size, and the integration size is 0.005 / NST. The vertical axis of FIG. 15 is CP
U time ratio.

That is, as described above, for each subsystem, if the solution is obtained while selecting the size of the numerical integration according to the order of the oscillation obtained when solving the equation of motion, the calculation time is further shortened. It becomes possible. As described above, the so-called partial structure synthesizing method B in which the vibration analysis is performed by extracting only the necessary vibration in consideration of the non-linearity such as the multistage characteristic of the clutch disc with respect to the drive system torsional vibration
BA is put into practical use.

According to the partial structure synthesizing method BBA, the conventional method and comparison with experimental values are performed for high speed / high load and low speed / light load running as follows. . The calculation accuracy is almost the same as the conventional method, and the calculation time is about 1 /
It can be shortened to 2.

A vibration / noise reduction effect by changing the twisting characteristic of the clutch disc was predicted and confirmed in an actual vehicle. Thus, the drive system torsional vibration analysis of the truck by the partial structure synthesis method BBA is performed. The design of the multi-stage torsional characteristic of the clutch disk as shown in FIG. 12 performed by using the analysis method is described later in (1). It is performed according to each procedure of (5).

That is, regarding the design of the multi-stage torsional characteristic of the clutch disc, the vibration caused by the torsional vibration of the drive system
In order not to generate noise, idle noise region R1, low speed / light load region R2, high speed / high load region R shown in FIG.
3. For each of the stopper regions R4, the spring constants k1 to k4 are set so that the resonance rotational speed is equal to or lower than the practical vehicle speed region, and the torsion angles θ1 to θ4 that determine each region are set.

Basically, vibration caused by torsional vibration of the drive system
In order not to generate noise, it is necessary to reduce the spring constants k1 to k4, but if the spring constant is reduced, the torque loaded in each of the regions R1 to R4 is cleared,
It is necessary to increase the twist angle θ. Increasing the twist angle θ means making a larger hole in the clutch hub, and the expansion of θ is limited by the strength of the clutch hub.

Therefore, the spring constants k1 to k4 are set while preventing the twist angles θ1 to θ4 from becoming larger than necessary.
It is necessary to have a balanced structure so that is as low as possible. Conventionally, the twisting characteristic of the clutch disc is determined based on experience and experiments, and there is no standard for determining the twisting characteristic.Therefore, the twisting characteristic of the clutch disc is reviewed each time a new vehicle is developed or the engine output increases. However, in either case, development efficiency cannot be increased because it is decided by experience and experiment.

However, as will be described later, the range of operating angles θ2 to θ3 in the high-speed / high-load running region R3 having the largest torque fluctuation range among the torsional characteristics is determined by using the average axial torque TE0 of the engine. , Θ1 to θ4 can be set by computer simulation. As a result, the resonance of the torsional vibration at high speed and high load can be surely kept within the practical vehicle speed range.

Assuming that the engine torque fluctuation is ΔT, since 0.5 · TE0 <TE0 + ΔT <1.7 · TE0, the operating range does not dig into the stopper region and vibration and noise are prevented. . Each setting is performed according to the following procedures.

(1) First stage: In the idling noise region, the spring constant k is set so that the resonance becomes equal to or less than the idle speed.
Decide on 1. Next, by computer simulation, the actual operating ranges θ1 ′ to θ1 of the clutch disc are determined. Where θ
1'is the operating range limit on the negative side of the first stage, and θ1 is the operating range limit on the positive side of the first stage.

Incidentally, θ1 is, as will be described later, the average shaft torque TE0 of 0.
The torsion angle θ value corresponding to triple the torque is set. Then, the torsional characteristics are decided for the first stage, the second stage, the third stage,
When the stopper torque cannot be ensured or the strength of the clutch hub becomes insufficient, the hysteresis H1 is increased to narrow the actual operating range and set it within the narrower operating range θ1 ′ to θ1.

However, note that if the hysteresis H1 is set too large, the idle noise will deteriorate.
With the above cycle, the spring constant k1, the hysteresis H1,
Determine the operating range θ1 'to θ1 of the first stage. (2) Second stage: At low speed and light load, the spring constant k2 is determined so that the fourth-order torsional rotational speed is equal to or lower than the practical vehicle speed.

Next, the ranges θ1 and θ2 of the spring constant k2 are determined. According to the method described above, the actual operating range of the clutch disk is checked, and θ2 is set so that the operating range does not exceed θ2 during the fourth-order torsional resonance at low speed and light load. (3) Third stage: At the time of high speed and high load, the spring constant k3 is determined so that the resonance rotational speed of the fourth order of torsion is below the practical vehicle speed range.

The range θ2 to θ3 of the spring constant k3 is determined by the above-mentioned analysis method. Looking at the actual operating range of the clutch disk, the operating range is θ during torsional fourth resonance at high speed and high load.
Set θ3 so that it does not exceed 3. That is, high speed
As shown in FIG. 13, the upper limit twist angle θ3 under high load corresponds to a torque that is larger than the average axial torque TE0 by 0.7 times the average axial torque TE0, and at the θ position corresponding to the inclination of the spring constant k3. Is set to the value of.

This is because, as shown in FIG. 8, the peak of the rotational fluctuation ratio in the clutch B (see FIG. 4B) in which the torque that shifts to the final stage is relatively small, and the torque that shifts to the final stage are relatively small. Comparing with the peak of the rotation fluctuation ratio in the large clutch C (see FIG. 4C), the clutch B
In this case, the characteristic is that it is on the lower rotation speed side, and as the torque shifting to the final stage increases, the peak shifts to the higher rotation speed side.

As shown in FIG. 6, when the twist angle θ that shifts to the final stage is set so as to correspond to a torque 0.8 times larger than the average shaft torque TE0, the characteristic of the rotational fluctuation ratio is 1. The state shown by the dotted line is reached, and the peak of the rotation fluctuation ratio reaches the position where the rotation speed ratio is 1.0. Therefore, in such a set state, the resonance peak reaches the idle speed, which is not preferable.

Therefore, a value corresponding to a torque that is 0.7 times larger than the average axial torque TE0 from the average axial torque TE0 for obtaining the characteristic shown by the solid line in FIG. The upper limit value that can be set is set. On the other hand, the lower limit torsion angle θ2 at high speed and high load is shown in Fig. 1.
As shown in FIG. 3, the value of the torsion angle θ is set to a value corresponding to the vicinity of a torque that is smaller than the average axial torque TE0 by 0.5 times the average axial torque TE0 and to the inclination of the spring constant k3.

That is, as shown in FIG. 4, two stages (θ1 to θ2, θ2 to θ) including a relatively small spring constant k2 are included.
Clutch B divided into 3) and one stage (θ1 to θ2)
With the clutch C, the characteristic of the clutch disk C approaches the characteristic of the clutch B by increasing the value of the torsion angle θ2. By the way, the rotation variation ratio characteristics of the clutch B and the clutch C at low speed and light load are as shown in FIG.

The characteristic of the alternate long and short dash line shown in FIG.
It is a rotation variation ratio characteristic when the torsion angle θ is set to correspond to a torque 0.6 times the average shaft torque TE0.
The characteristic of the dotted line is that the twist angle θ is 0.5 of the average axial torque TE0.
This is a rotation fluctuation ratio characteristic when it is set to correspond to double the torque. As shown by these characteristics, from the rotation variation ratio characteristics when the torsion angle θ is set to correspond to the torque of 0.5 times the average shaft torque TE0, when the state of the clutch C is shifted, the peak is obtained. Therefore, the torsion angle θ value corresponding to a torque in the vicinity of 0.6 times the average shaft torque TE0 is set as the upper limit of the torsion angle θ in the low speed / light load region, and the torsion angle θ becomes 0.5 times or less. The angle θ2 is set.

The characteristic of the clutch B in FIG. 10 is the characteristic when the value of the torsion angle θ1 at the lower limit of the low speed / light load region is set to 0.3 times the average shaft torque TE0, and is shown in FIG. Thus, the value of 0.3 times the average shaft torque TE0 is set as the lower limit of the torsion angle θ1. By the way, if the twist angle θ3 is within the strength of the clutch hub, it is OK.
Then, proceed as it is, but if the strength is insufficient,
By narrowing the torsion angle θ3 and optimizing the hysteresis (generally increasing the hysteresis), the actual operating range of the clutch disc is narrowed so that the operating range does not exceed θ3 during torsional fourth-order resonance at high speed and high load. Reset the twist angle θ3.

(4) Examine whether or not the stopper torque Ts is secured. In the torsional characteristic, when the torque when θ = θ3 does not reach Ts, a region having a spring constant k3 ′ slightly higher than the spring constant k2 is provided between the second step and the third step. Therefore, in the range of θ1 ′ to θ1,
The spring constant is k1 in the range of θ1 to new θ2, the spring constant is k2 in the range of new θ2 to θ2, the spring constant is k2 ′ in the range of θ2 to θ3, and the spring constant is k3, and the characteristics of the clutch disc are determined. .

(5) On the negative side, first, θ1 'is determined as a measure against idling noise. Then, from the strength of the clutch hub, θ4 'is decided, the new second stage, the new third stage,
Decide the new 4th stage. (6) The design standard of the stopper torque is set as follows.

Positive side: 1.7 times or more the maximum engine torque. Negative side: More than engine maximum torque. By the setting of the spring constants k1 to k4 and the operating angles θ1 to θ4 that determine each operating region by computer simulation as described above, the vibration caused by the torsional vibration of the drive system
It is possible to eliminate noise in a well-balanced manner and improve development efficiency when developing a new car.

[0107]

As described above in detail, according to the vibration analysis method for a system having a non-linear element (claim 1) of the present invention, when performing a vibration analysis for a system having one or more non-linear elements, The system is divided into a plurality of subsystems with the non-linear element as a boundary, and then the equations of motion for each subsystem are established, and then the equations of motion for the above subsystems are made simultaneous and the nonlinear equations are calculated. By inputting the characteristic of the element and obtaining the solution, the vibration analysis of the whole system can be performed, and the vibration analysis of the system having the nonlinear element can be performed by the CPU calculation in a short time. There is an advantage that it is possible to perform design based on reliable calculation without designing based on experiments and experience.

Further, according to the vibration analysis method for a system having a non-linear element of the present invention (claim 2), in the method according to claim 1, a linear element is provided on one of the right side and the left side, and On the other hand, by inputting the characteristics of the non-linear element so that the non-linear element comes, the equations of motion for each of the above-mentioned subsystems are simultaneous, and the solution is obtained, whereby the vibration analysis of the entire system is performed. With a simple configuration, it becomes possible to easily input the characteristics of the nonlinear element in the vibration analysis of the system having nonlinear elements, and it becomes possible to perform the vibration analysis of the system by a short-time CPU operation. Therefore, there is an advantage that a design based on reliable calculation can be performed without designing.

Furthermore, in the vibration analysis method (Claim 3) for a system having a non-linear element of the present invention, the order of the vibration to be obtained when the equation of motion is made for each subsystem in the method described in Claim 1. It is possible to easily add the error information in the vibration analysis of the system having non-linear elements with a simple configuration that adds the error information calculated according to the above, and the vibration analysis of the system can be performed by the CPU calculation in a short time. As a result, there is an advantage that design based on reliable calculation can be performed without designing based on experiments and experience.

Further, in the vibration analysis method (Claim 4) for a system having a non-linear element of the present invention, in the method according to Claim 2 or Claim 3, the vibration obtained when solving the equation of motion for each subsystem Since the solution is obtained while selecting the size of the numerical integration according to the order of, the calculation time can be further shortened. According to the torsional vibration analysis method for a vehicle drive system of the present invention (Claim 5), when the torsional vibration analysis is performed for a vehicle drive system having a clutch disc as a non-linear element, a clutch is used as a boundary. , The drive system is divided into a first subsystem that extends from the clutch to the end element of the drive system and a second subsystem that extends from the engine to the clutch. , The equation of motion expressed by the modal coordinates of the natural vibration of the drive system and the boundary degree of freedom obtained by Guyan's static contraction, and the engine, the clutch disk, the external force for the second subsystem. After creating the equation of motion considering the torque fluctuation, the equations of motion for the first subsystem and the second subsystem are made simultaneous, and the characteristic of the clutch disk is input to obtain the solution. With the simple structure of performing torsional vibration analysis for the entire drive system, the torsional vibration analysis for the entire drive system having non-linear elements can be performed by a short-time CPU calculation. Therefore, there is an advantage that a design based on reliable calculation can be performed without designing.

According to the torsional vibration analysis method for a vehicle drive system of the present invention (claim 6), in the method according to claim 5, a linear element is provided on one of the right side and the left side, and On the other hand, the equations of motion for the first subsystem and the second subsystem are made simultaneous so that the torque information of the clutch disc comes, and the characteristic of the clutch disc is input to obtain a solution. Thus, the torsional vibration analysis of the entire drive system can be performed by a short-time CPU calculation in consideration of the torque information of the clutch disk with a simple configuration of performing the torsional vibration analysis of the entire drive system. As a result, there is an advantage that it is possible to perform design based on reliable calculation without designing based on experiments and experience.

Furthermore, in the torsional vibration analysis method for a vehicle drive system according to the present invention (claim 7), in the method according to claim 5, the first subsystem is

[0113]

(Equation 15)

The error information [Kr] represented by is replaced with the rigidity matrix inside the first subsystem, and a nonlinear equation is formed with a simple structure in which a motion equation in which the modal compensation is combined with the constraint modal method is created. The torsional vibration analysis of the entire drive system that has elements can be performed by a short-time CPU operation in the state where the modal compensation is combined with the constraint modal method, and reliable calculation can be performed without designing based on experiments and experience. There is an advantage that it becomes possible to design based on.

According to the torsion angle / torque characteristic determining method for the multi-stage clutch disc of the present invention (claim 8), the multi-stage clutch disc as a non-linear element provided in the vehicle drive system is operated at a low speed / light load. When determining the torsion angle / torque characteristics at high speed / high load, the spring constant that defines the low speed / light load area in the torsion angle / torque characteristics is determined so that the resonance speed falls below the practical vehicle speed range. In addition, the torsional vibration analysis method using the partial structure synthesis method, that is, the first branch system from the clutch to the end element of the drive system with the clutch as the boundary, and the engine to the clutch The second system leading to the first system is divided into the second system and the operability expressed by the modal coordinates of the natural vibration of the drive system and the boundary degrees of freedom obtained by Guyan's static contraction. With make a equation, second
For the subsystem, the equations of motion considering the engine, the multi-stage clutch disc, and torque fluctuation as external force are established, and then the equations of motion for the first subsystem and the second subsystem are made simultaneous. Then, by inputting the characteristics of the multi-stage clutch disc and obtaining a solution, the upper limit torsion angle in the low speed / light load region is determined by using the torsional vibration analysis method for analyzing the torsional vibration of the entire drive system. In addition, the spring constant that defines the high-speed / high-load region in the torsion angle / torque characteristics is determined so that the resonance speed falls below the practical vehicle speed range, and the above partial structure synthesis method is used. The multi-stage clutch disc as a non-linear element provided in the drive system for a vehicle has a simple structure in which the upper limit torsion angle in the high-speed / high-load region is determined by using the torsional vibration analysis method. Torsional vibration analysis of the entire drive system equipped with a latch disk can be performed by CPU calculation in a short time, so that design based on reliable calculation can be performed without designing by experiment or experience. There are advantages.

According to the torsion angle / torque characteristic determining method for the multi-stage clutch disc of the present invention (claim 9), the multi-stage clutch disc as a non-linear element provided in the vehicle drive system is operated at a low speed / light load. And high speed
When obtaining the torsion angle / torque characteristic at high load, the spring constant that defines the low speed / light load region in the torsion angle / torque characteristic is determined so that the resonance speed falls below the practical vehicle speed range. A first torsional vibration analysis method using a partial structure synthesis method, that is, a first branch system from the clutch to the end element of the drive system with the clutch as a boundary, and from the engine to the clutch To the second system leading to the first system, the first system is expressed by the equation of motion expressed by the modal coordinates of the natural vibration of the drive system and the boundary degrees of freedom obtained by Guyan's static contraction. With respect to the second subsystem, the equation of motion considering the engine, the multi-stage clutch disc, and torque fluctuation as an external force is established, and then the first subsystem and the second subsystem are formed. The equation of motion for the system is Then, by inputting the characteristics of the multi-stage clutch disc and obtaining a solution, the first torsional vibration analysis method for analyzing the torsional vibration of the entire drive system is used, Determine the upper limit twist angle and make sure that the resonance speed is below the practical vehicle speed range.
A spring constant that defines the high-speed / high-load region in the torsion angle / torque characteristics is determined, and a second torsional vibration analysis method using a partial structure synthesis method, that is, the clutch is used as a boundary to drive the drive. The system is divided into a first subsystem that extends from the clutch to the end element of the drive system and a second subsystem that extends from the engine to the clutch, and then the drive system for the first subsystem. The equation of motion expressed by the modal coordinates of the natural vibration of the system and the boundary degrees of freedom obtained by Guyan's static contraction,

[0117]

(Equation 16)

The error information [Kr] represented by is replaced with the rigidity matrix inside the first subsystem, and the constraint modal method is combined with the mode compensation, and the second subsystem is The engine, the multi-stage clutch disc, and the equation of motion considering the torque fluctuation as the external force are created, and then the equations of motion for the first subsystem and the second subsystem are made simultaneous to form the multi-stage clutch disc. By inputting the characteristics of the above, and obtaining a solution, the upper limit torsion angle in the high-speed / high-load region is determined by using the second torsional vibration analysis method that analyzes the torsional vibration of the entire drive system. With a simple configuration, regarding the multi-stage clutch disc as a non-linear element provided in the vehicle drive system, in order to reliably design the characteristics of the clutch disc, the torsional vibration analysis of the entire drive system is performed. Become so performed by brief CPU operation, without performing the design by experiments and experience, etc., there is an advantage to be so performed a design based on reliable calculations.

Furthermore, according to the multi-stage clutch disc of the present invention (claim 10), in the multi-stage clutch disc provided in the vehicle drive system, the upper limit torsion angle in the high-speed / high-load region in the torsion angle / torque characteristics is high. Is set to a value corresponding to 1.7 times the engine average shaft torque at high speed and high load, and the lower limit of the twist angle / torque characteristic in the high speed / high load region is The multistage clutch disc as a non-linear element provided in the vehicle drive system has a simple configuration in which the value is set to a value corresponding to about 0.5 times the engine average shaft torque under high load, and the high speed of the clutch disc・ Characteristics in the high load area can be designed reliably and easily, and noise can be reliably reduced. Torsional vibration analysis of the entire drive system can be performed in a short time. Become so performed by the CPU operation, without designing the clutch disc by experiments and experience, etc., there is an advantage to be so performed a design based on reliable calculations.

According to the multistage clutch disc of the present invention (claim 11), in the multistage clutch disc provided in the vehicle drive system, the upper limit torsion angle in the high-speed / high-load region in the torsion angle / torque characteristics is high. , A value corresponding to 1.7 times the engine average shaft torque at high speed / high load, the lower limit of the torsion angle / torque characteristic of the high speed / high load region, and the low speed / light load region. The upper limit of the twist angle is set to a value corresponding to around 0.5 times the engine average shaft torque at the high speed / high load, and the lower limit of the low speed / light load region in the twist angle / torque characteristic is set. With a simple configuration in which the torsion angle is set to a value corresponding to about 0.3 times the average engine shaft torque at the time of high speed and high load, the multistage clutch as a non-linear element provided in the vehicle drive system is provided. With regard to the Chi-disc, the characteristics of the clutch disc in the high-speed / high-load region and the low-speed / light-load region can be reliably and easily designed, and reliable noise reduction can be achieved. The torsional vibration analysis can be performed by a CPU calculation in a short time, and there is an advantage that a design based on reliable calculation can be performed without designing a clutch disc through experiments, experience, or the like.

[Brief description of drawings]

FIG. 1 is a schematic diagram showing a simulation model of a drive system in a 6 × 4 track as an embodiment of the present invention.

2A to 2D are schematic diagrams showing respective models of a drive system as an embodiment of the present invention.

FIG. 3 is a schematic diagram showing a simulation result of a drive system in a 6 × 4 track as an embodiment of the present invention.

4A to 4C are schematic graphs showing hysteresis characteristics of the clutch device as the embodiment of the present invention.

5A to 5C are graphs showing response characteristics of a drive system in a simulation as an embodiment of the present invention.

FIG. 6 is a graph showing frequency response characteristics of a drive system in a simulation as an embodiment of the present invention.

FIG. 7 is a graph showing frequency response characteristics of a drive system in a simulation as an embodiment of the present invention.

FIG. 8 is a graph showing frequency response characteristics of a drive system in a simulation as an embodiment of the present invention.

FIG. 9 is a graph showing frequency response experimental characteristics of a drive system with respect to a simulation result as an embodiment of the present invention.

FIG. 10 is a graph showing frequency response characteristics of a drive system in a simulation as an embodiment of the present invention.

FIG. 11 is a graph showing noise level characteristics of the drive system according to the embodiment of the present invention.

FIG. 12 is a graph showing a hysteresis characteristic of the clutch device according to the embodiment of the present invention.

FIG. 13 is a graph showing a hysteresis characteristic of the clutch device according to the embodiment of the present invention.

FIG. 14 is a diagram for explaining a CPU operation time characteristic in a simulation as an embodiment of the present invention.

FIG. 15 is a diagram for explaining CPU time calculation characteristics in comparison with a conventional method as an embodiment of the present invention.

FIG. 16 is a vertical cross-sectional view showing an example of a clutch device as an embodiment of the present invention.

FIG. 17 is a partially cutaway front view showing an example of the clutch device according to the embodiment of the present invention with its front structure partially broken.

FIG. 18 is a graph showing hysteresis characteristics in an example of the clutch device according to the embodiment of the present invention.

[Explanation of symbols]

 1 engine 2 clutch 3,4 transmission 5-7 propeller shaft 8 through shaft 9,10 propeller shaft 11 differential 12 wheel 13 vehicle body 14 differential 15 wheel 16 vehicle body 101 spline hub 101a fan notch 102,103 drive clutch plate 104a, 104a sub Plate 105 Bush 106,107 Pin 108a, 108b Friction washer 109a Non-metallic friction washer 109b Friction washer 110 Friction plate 111 Wave spring 112 First step spring 113 Second step spring 114 Second step outer spring 115 Second step inner spring 116, 117 Spring seat 118 Spring seat 119-123 Notch edge 125 126 Ogimado

Claims (11)

[Claims] 1. When performing vibration analysis on a system having one or more non-linear elements, the non-linear element is used as a boundary to divide the system into a plurality of sub-systems, and a motion equation is calculated for each sub-system. After completion, the equations of motion for each of the above-mentioned subsystems are made simultaneous, the characteristics of the non-linear elements are input, and the solution is obtained to perform the vibration analysis of the entire system. Vibration analysis method for systems with elements. 2. The characteristics of the non-linear element are input so that the linear element comes to one of the right side and the left side and the non-linear element comes to the other of the right side and the left side. A characteristic is that vibration analysis of the entire system is performed by making simultaneous equations of motion and obtaining a solution.
A vibration analysis method for a system having a nonlinear element according to claim 1. 3. A system having a non-linear element according to claim 1, wherein error information calculated in accordance with the order of vibration to be calculated is taken into consideration when the equation of motion is created for each subsystem. Vibration analysis method. 4. The solution according to claim 2, wherein the solution is obtained for each subsystem while selecting the size of the numerical integration according to the order of the oscillation obtained when solving the equation of motion. Vibration analysis method for systems with nonlinear elements. 5. When performing torsional vibration analysis on a vehicle drive system having a clutch disc as a non-linear element, the first drive system extends from the clutch to the end element of the drive system with the clutch serving as a boundary. And a second subsystem extending from the engine to the clutch, and then the first subsystem was obtained by modal coordinates of natural vibration of the drive system and Guyan's static contraction. The equation of motion represented by the boundary degree of freedom is created, and the equation of motion considering the engine, the clutch disc, and the torque fluctuation as an external force is created for the second subsystem, and then the first The equations of motion for the second system and the second system are made simultaneous, and the characteristics of the clutch disk are input,
A torsional vibration analysis method for a vehicle drive system, characterized by performing a torsional vibration analysis for the entire drive system by obtaining a solution. 6. The first subsystem and the second subsystem so that a linear element is provided on one of the right side and the left side and torque information of the clutch disc is provided on the other side of the right side or the left side.
6. The torsional vibration analysis of the entire drive system is performed by simultaneous equations of motion equations for the subsystem, inputting the characteristics of the clutch disc, and obtaining a solution. Torsional vibration analysis method for vehicle drive system. 7. For the first subsystem, The error information [Kr] indicated by is replaced with the rigidity matrix inside the first subsystem to form a motion equation in which a modal compensation is combined with a constrained modal method. Torsional vibration analysis method for vehicle drive system. 8. A resonance speed of a multi-stage clutch disc as a non-linear element provided in a vehicle drive system at a low speed / light load and a high speed / high load when a torsional angle / torque characteristic is obtained is equal to or lower than a practical vehicle speed range. So that the twist angle
The spring constant defining the low speed / light load region in the torque characteristic is determined, and the torsional vibration analysis method using the partial structure synthesis method, that is, the clutch is used as a boundary to drive the drive system from the clutch It is divided into a first subsystem that reaches the end element of the system and a second subsystem that extends from the engine to the clutch, and then, for the first subsystem, modal coordinates of the natural vibration of the drive system are set. The equation of motion expressed by the boundary degree of freedom obtained by Guyan's static contraction is constructed and
For the subsystem, the equations of motion considering the engine, the multi-stage clutch disc, and torque fluctuation as external force are established, and then the equations of motion for the first subsystem and the second subsystem are made simultaneous. Then, by inputting the characteristics of the multi-stage clutch disc and obtaining a solution, the upper limit torsion angle in the low speed / light load region is determined by using the torsional vibration analysis method for analyzing the torsional vibration of the entire drive system. Furthermore, the spring constant that defines the high-speed / high-load region in the torsion angle / torque characteristics is determined so that the resonance speed falls below the practical vehicle speed range, and the above partial structure synthesis method is used. A method of determining a torsion angle / torque characteristic for a multi-stage clutch disc, characterized in that the upper limit torsion angle in the high speed / high load region is determined by using a torsion vibration analysis method. 9. A resonance rotational speed is not more than a practical vehicle speed range when determining torsion angle / torque characteristics at low speed / light load and high speed / high load of a multi-stage clutch disk as a non-linear element provided in a vehicle drive system. So that the twist angle
The spring constant that defines the low-speed / light-load region in the torque characteristic is determined, and the first torsional vibration analysis method using the partial structure synthesis method, that is, with the clutch as the boundary, the drive system is set to the clutch. From the engine to the end element of the drive system and the second system from the engine to the clutch, the first subsystem is divided into the natural vibrations of the drive system. The equation of motion expressed by the modal coordinates and the boundary degrees of freedom obtained by Guyan's static contraction is constructed, and
For the second subsystem, a motion equation considering the engine, the multi-stage clutch disc, and torque fluctuation as an external force is created, and then the equations of motion for the first subsystem and the second subsystem are created. At the low speed / light load by using the first torsional vibration analysis method for conducting torsional vibration analysis of the entire drive system by inputting the characteristics of the multi-stage clutch disc and obtaining a solution. The upper limit torsion angle of the region is determined, and further, the spring constant that defines the high-speed / high-load region in the torsion angle / torque characteristics is determined so that the resonance rotational speed is equal to or lower than the practical vehicle speed range, and the partial structure is also determined. A second torsional vibration analysis method using a synthesis method, that is, a first branch system from the clutch to the end element of the drive system with the clutch as a boundary, and an engine to the clutch. With the second subsystem Then, for the first subsystem,
The equation of motion expressed by the modal coordinates of the natural vibration of the drive system and the boundary degrees of freedom obtained by Guyan's static contraction is given by The error information [Kr] indicated by is replaced with the stiffness matrix inside the first subsystem to combine modal compensation with modal compensation, and the second subsystem is provided with the engine, After the multi-stage clutch disc and the equation of motion considering the torque fluctuation as the external force are created,
Torsional vibration analysis of the entire drive system is performed by simultaneously establishing simultaneous equations of motion for the first and second subsystems, inputting the characteristics of the multistage clutch disc, and obtaining a solution. A method for determining a torsion angle / torque characteristic for a multi-stage clutch disc, characterized in that an upper limit torsion angle in the high speed / high load region is determined by using a second torsion vibration analysis method. 10. In a multi-stage clutch disc provided in a vehicle drive system, an upper limit twist angle in a high-speed / high-load region in a twist angle / torque characteristic thereof is 1.7 of an engine average shaft torque at high-speed / high load. The value is set to a value close to double, and the lower limit twist angle of the high-speed / high-load region in the twist angle / torque characteristic is close to 0.5 times the engine average shaft torque at the high-speed / high load. A multi-stage clutch disc characterized by being set to corresponding values. 11. In a multi-stage clutch disc provided in a vehicle drive system, an upper limit twist angle in a high-speed / high-load region in a twist angle / torque characteristic thereof is 1.7 of an engine average shaft torque at high-speed / high load. It is set to a value corresponding to approximately twice, and the lower limit twist angle in the high speed / high load region and the upper limit twist angle in the low speed / light load region in the twist angle / torque characteristic are the engine at the high speed / high load, respectively. The value is set to a value corresponding to around 0.5 times the average shaft torque, and the lower limit of the torsion angle / torque characteristic in the low speed / light load region is the engine average shaft torque at the high speed / high load. A multi-stage clutch disc, which is set to a value corresponding to about 0.3 times.