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1 finite element model In order to achieve accurate modeling, according to the involute equation expression, the parametric language of ANSYS software is combined with the corresponding command stream to create the model of the gear pair.
1) Units and materials The gear material studied in this paper uses 20CrMoH, the modulus of elasticity is 210GPa, and the Poisson's ratio is 0.27.
There are two types of units used for meshing:
Solidl64 solid unit and shelll63 thin shell unit. Selecting the default Const.stress algorithm for the solidl64 unit can improve the efficiency of the analysis problem. The S/RCo-rotationalHughes.Liu in-plane multi-integration point improved unit algorithm is selected for the shelll63 unit to eliminate the hourglass condition.
A rigid body definition is used for one of the gears in order to reduce the time for explicit analysis and improve the efficiency of the solution. At the same time, since a rigid body is defined, the degrees of freedom of all nodes in the rigid body are coupled to the center of mass of the rigid body. There are 6 degrees of freedom to facilitate the application of the load.
2) Dividing the mesh Considering the force characteristics of the gear meshing process, the mesh at the gear teeth should be dense when meshing, and the mesh density can be appropriately reduced away from the gear teeth. It should be noted that:
a. Try to avoid the use of degenerate shell elements and solid elements, such as triangular shell elements and tetrahedral solid elements, the calculation accuracy is poor relative to quadrilateral shell elements and hexahedral solid elements; b. the unit size is as uniform as possible to avoid relatively A small unit area will result in a small solution time step, resulting in a long solution time; c. Try to avoid bad shape elements that may cause hourglasses; d. Use material parameters that are as realistic as possible in the material model. A high unrealistic modulus of elasticity is used to express the rigid body. The gear mesh model is shown in Figure 1.
3) PART definition PART is a unit set with the same material, unit properties and unit type. In ANSYS/LS-DYNA, PART must be defined reasonably. Contact interface definition, loading, constraints and boundary conditions need to be passed. PART to operate, here you need to create 2 PART, corresponding to the big gear and pinion.
4) Defining the contact contact problem is a highly nonlinear behavior. It is important to understand the characteristics of the problem and to establish a reasonable model. Here, according to the situation of the two gear models, the gear contact surfaces that may be in contact are respectively defined. It is also necessary to define the contact type and some parameters related to the contact, including static friction coefficient, dynamic friction coefficient, stiffness factor and penetration tolerance. .
5) Constraint and load Since the application of the load in ANSYS/LS-DYNA software is done by means of an array, according to the normal operation of the gear, it is necessary to define the driving torque of the driving wheel and the blocked torque of the driven wheel. An array of time.
2 Solve the solution parameters before the solution process, mainly the solution time control, the result file output time interval, the result file output format, and so on.
According to the rotation speed of the gear and only the analysis of the engagement of a pair of gears, the solution time is set to 0.001 s. The result file output time interval is 40 steps, and the output result file format is . Rst and. D3plot can perform the query of the results in the ANSYS postprocessor and the postprocessor module that comes with LS-DYNA.
3 Analysis results 3.1 The maximum equivalent stress 2 in the gear contact process is the equivalent stress cloud diagram of the gear at different times in one meshing period. The whole process lasts for 0.001 s.3, which is the maximum equivalent stress of the gear with time. The peak of the secondary equivalent stress, the first time is due to the impact of the gear drive start, the stress at the moment of the gear tooth contact is very large, the maximum stress reaches 833MPa, and the stress caused by the impact load decays rapidly with the increase of time. At 0.0001s, the maximum stress attenuation between the gears is 181 MPa, and as the contact is stable, the contact force gradually increases. At the time of 0.00015 s, the second equivalent stress peak occurs, and the maximum stress reaches 853 MPa. It can be seen from the above data that the stress value caused by the impact load during the start of the gear transmission is large, which is impossible to predict and calculate by the empirical formula, and various gear failure modes may occur due to the large stress, so It is important to determine the impact load during the gearing process. It can be seen that for accurate design and calibration, dynamic simulation of the gear drive is important.
3.2 Gear meshing characteristics Select the maximum contact stress moment, 4 is the equivalent stress distribution diagram of the involute gear meshing transmission. It can be seen from the stress cloud diagram that stress is generated only near the teeth that contact each other, and the closer to the contact point, the greater the stress The stress in the area away from the contact point is basically zero, which is consistent with the actual stress distribution of the gear transmission, which also conforms to the Saint-Venant law in elastic mechanics. In order to analyze the influence of the impact load on the tooth contact portion and the root portion, a group of units is selected in the tooth and root portion, respectively, as shown in FIG.
In the X-direction (radial) stress versus time curve and the Y-direction (tangential) stress versus time curve, it can be seen that there are two impacts in the gear contact process, and the second impact load is much larger, and the stress caused by it When reaching 250 MPa, the stress on the pinion in the X direction is much smaller than that on the large gear on the driven wheel. The stress generated by the drive wheel in the Y direction is much larger than the stress on the driven wheel, close to 1000 MPa. Although there are two peaks in the Y-direction stress, the stress generated in the second impact is small, close to 0.
The equivalent stress versus time curve of the two elements is shown in Fig. 6c. From the equivalent stress diagram, the equivalent stress on the driving wheel is larger than that of the driven wheel, and the stress after the second wave is rapidly reduced.
Figure 7ab shows the curve of the X-direction (radial) stress versus time and the Y-direction (tangential) stress with time for the selected group of the roots. It can be seen that the stress in both directions has one inverse. The process of the direction, this is because the upper wheel of the driving wheel and the driven wheel tooth contact during the rebound after the first impact, so a reverse stress is generated.
The curve of the equivalent stress of the root unit with time is shown in Fig. 7c. It can be seen that the 21685 element equivalent stress curve and the 8101 unit stress curve basically coincide, while the 21685 unit and the 8101 unit are basically related to the tooth center line in the geometric position. Symmetrical, so the equivalent stress at the root of the teeth on both sides of the gear teeth is symmetrically distributed about the center line of the gear teeth during the gear meshing process. At the same time, the effective stress of the root is pulsed with the operation of the gear, and the stress value fluctuates due to the existence of the flank clearance and the like, which is very serious for the damage of the gear.
4 Conclusions The steps of the dynamic contact simulation analysis of the involute spur gear and the method for determining the parameters are introduced, and the criteria to be paid attention to during the analysis process are proposed. The variation law of the stress of a pair of involute spur gears during the meshing process was calculated by ANSYS/LS-DYNA software. The dynamic simulation results of the involute spur gear were analyzed in detail, and the impact load in the gear transmission was discussed. The effect shows that at the initial moment of contact, the stress value due to the impact is large, and the impact load needs to be considered in the gear design process. The selection of contact parameters involved in the analysis process is more accurate if it can be corrected by means of experimental results.