Design support method for dynamic vibration absorbers

Abstract

To provide a technique of supporting design of a dynamic vibration absorber suitable to a processing tool.SOLUTION: A reception unit 112 receives a resonant frequency of a processing tool. A holding unit 120 holds a relation between a spring constant and a crush amount of an elastic body. An information acquisition unit 114 refers to the holding content of the holding unit 120 and acquires at least one of information about the crush amount of the elastic body in a dynamic vibration absorber having a resonant frequency matching or approximate to the received resonant frequency of the processing tool, information about the size of the elastic body, information about the number of elastic bodies and information about the mass of a mass body. An output unit 116 outputs the acquired information to an output device 130.SELECTED DRAWING: Figure 18

Description

Technical Field

[0001] The present disclosure relates to a technique for assisting the design of a dynamic vibration absorber that suppresses vibrations generated in a processing process and a structure capable of adjusting the resonance frequency of the dynamic vibration absorber.

Background Art

[0002] In order to suppress chatter vibration and forced vibration, a technique of attaching a dynamic vibration absorber including a spring element and a mass body (weight) to a processing tool to improve damping is known. By attaching a dynamic vibration absorber having a resonance frequency that matches or approximates the resonance frequency of the processing tool attached to the spindle or feed device of a machine tool, the resonance phenomenon near the resonance frequency of the processing tool is suppressed.

[0003] Patent Document 1 discloses a tool holder provided with an absorber mass disposed in a cavity inside a shank. In this tool holder, a pair of elastic supports that regulate both ends of the absorber mass are pressed against the absorber mass by at least one of a pair of pressure plates disposed on both ends of the absorber mass. This tool holder has an adjustment screw for moving a pressure plate movable along the longitudinal axis with respect to the elastic support, and has a function of adjusting the rigidity of the elastic support.

Prior Art Documents

Patent Documents

[0004]

Patent Document 1

Summary of the Invention

Problems to be Solved by the Invention

[0005] For a dynamic vibration absorber to exert its damping effect, its resonant frequency must match or approximate the resonant frequency of the machining tool. However, in reality, accurately predicting the resonant frequency of a dynamic vibration absorber at the design stage is not easy. For example, an O-ring is a spring element whose contact area with the contact surface changes as it deforms, resulting in a nonlinear relationship between force and displacement. Even if the resonant frequency of a dynamic vibration absorber containing such a spring element is predicted using CAE (Computer-Aided Engineering) analysis, the dynamic behavior of the spring element cannot be reproduced, so the predicted resonant frequency and the resonant frequency of the dynamic vibration absorber actually manufactured according to the prediction usually do not match.

[0006] When designing a dynamic vibration absorber for a particular machining tool, one might consider creating numerous prototypes and experimentally matching their resonant frequencies to those of the machining tool. However, this approach is costly and time-consuming. In particular, when designing a dynamic vibration absorber for a specialized tool used in a specific machining process that is not mass-produced, the method of creating numerous prototypes and experimentally matching resonant frequencies is not cost-effective and is therefore difficult to adopt.

[0007] Furthermore, the resonant frequency of a machining tool is not determined solely by the machining tool itself, but is also influenced by the structure of the machine tool to which the machining tool is attached (spindle or feed mechanism), the mounting force, and the condition of the mounting surface. For example, even with the same machining tool, if the rigidity of the mounting part of the machine tool to which it is attached differs, the resonant frequency of the machining tool will differ (if the rigidity of the mounting part is low, the resonant frequency of the machining tool will be low). In particular, if the rigidity of the mounting part of the machine tool is not sufficiently high compared to the rigidity of the machining tool, the change in resonant frequency will be large, and therefore the resonant frequency of the machining tool will differ depending on the machine tool to which it is attached. For this reason, even for general-purpose tools that are not specialized for a particular machining process, if the rigidity of the mounting part of the machine tool is not sufficiently high, it will not be possible to mass-produce the dynamic vibration absorber. Therefore, it is desirable to establish a method that can suitably support the design of dynamic vibration absorbers suitable for machining tools (more precisely, machining tools attached to the machine tool structure).

[0008] Furthermore, the tool holder disclosed in Patent Document 1 has a function to adjust the rigidity of the elastic support, that is, a function to adjust the spring constant (spring stiffness), but since it can only adjust the rigidity of two elastic supports, the adjustment range is not wide. Therefore, it is also desired to realize a structure that can adjust the resonant frequency of a dynamic vibration absorber attached to a machining tool over a wide range.

[0009] This disclosure is made in view of these circumstances and aims to provide technology to support the design of dynamic vibration absorbers suitable for machining tools, and to provide technology that allows for adjustment of the resonant frequency of the dynamic vibration absorber. [Means for solving the problem]

[0010] To solve the above problems, a design support method in one aspect of the present disclosure is a method for supporting the design of a dynamic vibration absorber to be attached to a machining tool, comprising a mass body and an elastic body held between the mass body and the machining tool, the method comprising: a reception step of receiving the resonant frequency of the machining tool; an acquisition step of obtaining at least one of the following information regarding the amount of compression of the elastic body, the size of the elastic body, the number of elastic bodies, and the mass of the mass body in a dynamic vibration absorber having a resonant frequency that matches or approximates the received resonant frequency of the machining tool, by referring to the relationship between the spring constant and the amount of compression of the elastic body held in the holding part; and an output step of outputting the acquired information.

[0011] Another aspect of the present disclosure is a dynamic vibration absorber structure that is attached to a component and has an adjustable resonant frequency, comprising a mass body and two or more elastic bodies held between the mass body and the component, and further comprising a pressing section for independently changing the amount of compression of at least one of the elastic bodies.

[0012] A further aspect of the present disclosure is a machine tool to which a machining tool is attached, the machining tool having a dynamic vibration absorber comprising a mass body and two or more elastic bodies held between the mass body and the machining tool, the machining tool comprising an adjustment structure capable of adjusting the resonant frequency of the dynamic vibration absorber by independently changing the amount of compression of at least one elastic body.

[0013] Furthermore, any combination of the above components, as well as any conversion of the expressions of this disclosure between methods, apparatus, systems, etc., are also valid as aspects of this disclosure. [Brief explanation of the drawing]

[0014] [Figure 1] This diagram shows a machining tool attached to the spindle of a machine tool. [Figure 2] This diagram shows the overall structure of the machining tool before redesign. [Figure 3] This figure shows an example of a cross-sectional structure of a machining tool incorporating a dynamic vibration absorber. [Figure 4] This figure shows another example of the cross-sectional structure of a machining tool incorporating a dynamic vibration absorber. [Figure 5] This figure shows another example of the cross-sectional structure of a machining tool incorporating a dynamic vibration absorber. [Figure 6] This diagram shows the configuration of an experimental dynamic vibration absorber. [Figure 7] This is a diagram showing photographs of the experimental environment. [Figure 8] This diagram shows the combinations of parameter values ​​for the prototype. [Figure 9] This figure shows an example of the results of a frequency analysis. [Figure 10] This diagram shows the relationship between the amount of compression and the spring constant. [Figure 11] This diagram shows the relationship between the amount of compression and the spring constant. [Figure 12] This diagram shows the relationship between the amount of compression and the spring constant. [Figure 13] This diagram shows the relationship between the amount of compression and the spring constant. [Figure 14] This diagram shows the relationship between the amount of compression and the spring constant. [Figure 15] It is a diagram showing the relationship between the amount of crushing and the spring constant. [Figure 16] It is a diagram showing the relationship between the amount of crushing and the spring constant. [Figure 17] It is a diagram showing the relationship between the amount of crushing and the spring constant. [Figure 18] It is a diagram showing the functional blocks of the design support device. [Figure 19] It is a diagram showing the flowchart of the design support method for the dynamic vibration absorber. [Figure 20] It is a diagram showing a virtual hole part and a virtual mass body. [Figure 21] It is a diagram showing an example of information displayed on the output device. [Figure 22] It is a diagram showing the experimental results demonstrating the effect of the dynamic vibration absorber. [Figure 23] It is a diagram showing an example of an adjustment structure capable of adjusting the resonance frequency of the dynamic vibration absorber. [Figure 24] It is a diagram showing another example of an adjustment structure capable of adjusting the resonance frequency of the dynamic vibration absorber. [Figure 25] It is a diagram showing a state where some elastic bodies are pressed by the tapered part. [Figure 26] It is a diagram showing a state where all elastic bodies are pressed by the tapered part.

Embodiments for Carrying Out the Invention

[0015] FIG. 1 shows a machining tool 3 attached to the spindle 2 of a machine tool 1. The machining tool 3 has a tool body (tool holder) 4, and the base end portion of the tool body 4 is rotatably attached to the spindle 2. A plurality of blade portions 5 having cutting edges are attached to the tip end portion of the tool body 4. In this example, four blade portions 5 are provided at the tip end portion of the tool body 4, but three or less, or five or more blade portions 5 may be provided.

[0016] In production sites for automotive parts and other components, specialized tools tailored to specific machining processes are frequently used to shorten processing time through increased efficiency and reduce the number of tools owned through process consolidation. In recent years, small machine tools have become widespread to save space, and an increasing number of production sites are using a combination of specialized tools and small machine tools for machining. Increased efficiency, process consolidation, and the miniaturization of machine tools are factors that cause regenerative chatter vibrations, affecting cutting stability. Furthermore, in small machine tools, the rigidity of the mounting section of the machine tool to which the machining tool is attached is often not sufficiently high compared to the rigidity of the machining tool, and the resonant frequency of the machining tool tends to vary from machine tool to machine tool. Therefore, in this embodiment, we propose a method to support the design of a dynamic vibration absorber attached to the machining tool 3 attached to the machine tool.

[0017] The design support method for dynamic vibration absorbers according to this embodiment can be used, for example, in a production site when chatter vibration or forced vibration becomes a problem during cutting using a machining tool 3. This method can be used to design a dynamic vibration absorber to be attached to the machining tool 3 and to redesign the machining tool 3 to have space for attaching the dynamic vibration absorber. Naturally, the design support method for dynamic vibration absorbers according to this embodiment can also be used when designing a new machining tool (for example, a special tool) equipped with a dynamic vibration absorber from scratch.

[0018] Figure 2 shows the overall structure of the machining tool 3 before redesign. The machining tool 3 has a mounting portion 6 at the base end of the tool body 4 that is attached to and detached from the mounting portion of the spindle 2. The space shown by the dashed line inside the tool body 4 is a hole 7 for housing the dynamic vibration absorber. In Figure 2, the hypothetical hole 7 is shown by the dashed line for the purpose of explaining the location and shape of the space in which the dynamic vibration absorber is housed, but the machining tool 3 before redesign may not actually have a hole 7 and may have a solid structure. The machining tool 3 before redesign may have a hole 7 formed in it, for example, for the purpose of weight reduction.

[0019] The hole 7 opens at the tip of the tool body 4 and may have a bottomed cylindrical shape centered on the axis of rotation. As shown in the figure, if the tool body 4 is composed of a large diameter portion 8 at the tip and a small diameter portion 9 at the base, the hole 7 may be configured as an internal space of the large diameter portion 8, or it may be configured as an internal space connecting the large diameter portion 8 to the small diameter portion 9. In the deflection vibration mode (or "bending vibration mode") in which the base is fixed and the tip vibrates, the base mainly plays the role of the stiffness (spring constant) of the vibration system, and the tip mainly plays the role of mass. Placing a dynamic vibration absorber in the hole 7 formed on the tip side is reasonable because it does not significantly reduce the stiffness and obtains an effective damping effect at positions of large displacement. However, even when the hole 7 is formed on the tip side of the tool body 4, the stiffness of the tool body 4 is reduced to some extent, so it is preferable that the hole 7 be designed in a shape and size that does not significantly reduce the second moment of area of ​​the tool body 4.

[0020] Figure 3 shows an example of the cross-sectional structure of a machining tool 3 incorporating a dynamic vibration absorber 20 designed using the design support method according to the embodiment. The machining tool 3 has a hole 7 (hole diameter D) for housing the dynamic vibration absorber 20, and the dynamic vibration absorber 20 is placed in the housing space where the opening of the hole 7 is closed with a lid 10. A pair of thrust bearings 28 may be provided at both ends of the dynamic vibration absorber 20, but different types of bearings may be provided.

[0021] The dynamic vibration absorber 20 of this embodiment comprises a substantially cylindrical mass body 26 and one or more elastic bodies 24 that function as spring elements. The elastic bodies 24 employ O-rings, which are annular rubber packings with a circular cross-section. Four annular grooves 22 with a diameter d are formed on the outer circumferential surface of the mass body 26, and the O-rings, which are the elastic bodies 24, are held in the annular grooves 22 and held between the mass body 26 and the machining tool 3. In this embodiment, the bottom surface of the annular groove 22 is a cylindrical surface, and the inner circumferential surface of the hole 7 is a cylindrical surface, and by making the difference in diameter between these cylindrical surfaces less than twice the diameter of the O-ring, the O-ring is subjected to radial deformation. The O-ring has nonlinear rigidity with respect to this deformation. In the tool body 4, the dynamic vibration absorber 20 is made to match or approximate the resonance frequency of the machining tool 3 attached to the spindle 2, thereby providing a function to suppress chatter vibration and forced vibration.

[0022] Figure 4 shows another example of the cross-sectional structure of a machining tool 3 incorporating a dynamic vibration absorber 20. In this example, elastic supports 29a and 29b, such as rubber balls, are provided at both ends of the dynamic vibration absorber 20 to restrict the axial movement of the dynamic vibration absorber 20. Support 29a is held between a recess provided at the rotational center of one end of the mass body 26 and a recess provided in the tool body 4, while support 29b is held between a recess provided at the rotational center of the other end of the mass body 26 and a recess provided in the cover 10.

[0023] Figure 5 shows another example of the cross-sectional structure of a machining tool 3 incorporating a dynamic vibration absorber 20. In this example, a stopper 27a provided on the tool body 4 and a stopper 27b provided on the cover 10 restrict the axial movement of the dynamic vibration absorber 20. In Figure 5, stopper 27a functions as the side wall of the right-side annular groove 22, and stopper 27b functions as the side wall of the left-side annular groove 22.

[0024] The following describes a method for supporting the design of the dynamic vibration absorber 20 attached to the machining tool 3 and the hole 7 formed in the machining tool 3. As a prerequisite, experiments are conducted in which multiple types of parameters that determine the resonant frequency of the dynamic vibration absorber are varied, and the resonant frequency is measured for each combination of these multiple types of parameters. This is used to create a database for deriving multiple types of parameters for the dynamic vibration absorber that generate the desired spring constant.

[0025] <Creating a spring constant database> The resonant frequency f of a dynamic vibration absorber can be calculated using the following equation (1), where m is the mass and k is the spring constant.

number

number

[0026] When designing a dynamic vibration absorber for the machining tool 3, the density ρ is determined once the material of the mass is determined. Furthermore, since there is a limit to the volume of the housing space for the dynamic vibration absorber that can be formed in the machining tool 3 (the hole 7 shown in Figures 2 to 5) from the standpoint of tool rigidity, the volume V of the mass 26 is determined within that limit. When the spring element in the dynamic vibration absorber 20 is an O-ring fitted into the annular groove 22, the spring constant k is affected by several types of parameters as follows. (a) Number of O-rings (b) Amount of compression of the O-ring (compression allowance) (c) O-ring size The spring constant k is also affected by the shape of the crushed surface (contact surface) that the O-ring contacts, but in this embodiment, the crushed surface is a cylindrical surface, and therefore there is no difference in the shape of the crushed surface.

[0027] In the experiments described below, prototypes were fabricated with various parameter combinations while changing several types of parameters that affect the resonant frequency, and the resonant frequency f of the prototypes was measured by frequency analysis.

[0028] Figure 6 shows the configuration of the experimental dynamic vibration absorber 30. The experimental dynamic vibration absorber 30 comprises a substantially cylindrical mass body 36 and O-rings 34. The outer surface of the mass body 36 has first annular grooves 32a, second annular grooves 32b, third annular grooves 32c, fourth annular grooves 32d, fifth annular grooves 32e, and sixth annular grooves 32f (hereinafter referred to as "annular grooves 32" unless otherwise specified) formed thereon to hold the O-rings 34, allowing the number of O-rings 34 provided on the mass body 36 to be changed. All annular grooves 32 have the same groove diameter d. In the mass body 36, at least the region in which annular grooves 32 are formed is located inside a pair of housing tubes 38a, 38b having an inner diameter D.

[0029] Figure 7 shows a photograph of the experimental environment. In this experiment, the housing tubes 38a and 38b were fixed in a vise, and the longitudinal central region of the mass body 36 was vibrated with an impulse hammer. Fast Fourier Transformation (FFT) was performed on the values ​​detected by an acceleration sensor attached to the mass body 36 to measure the resonant frequency. To mitigate measurement errors, the average of 10 measurements was derived as the resonant frequency f. In the experiment, prototypes (dynamic vibration absorbers 30) with several different parameters were fabricated, and the resonant frequency f of each prototype was measured. <Experimental Parameters> • Mass of mass 36 • Groove diameter d of mass body 36 • Hole diameter D of the housing tube 38 • O-ring size 34 • Number of O-rings (34) • Amount of crushing

[0030] The O-ring 34 is made of nitrile rubber conforming to the JIS P standard, with a wire diameter of 3.5 mm. The O-ring 34 is available in five sizes according to the P standard: P22A (inner diameter 21.7 mm), P29 (inner diameter 28.7 mm), P35 (inner diameter 34.7 mm), P41 (inner diameter 40.7 mm), and P48 (inner diameter 47.7 mm).

[0031] Figure 8 shows the combinations of parameter values ​​for the prototypes fabricated in the experiment. In the experiment, five different masses 36 (610g, 980g, 1380g, 1840g, 2460g) were prepared, and the groove diameter d of the mass 36 was formed to be slightly larger than the inner diameter of the O-ring 34 used. In the experiment, the amount of compression of the O-ring 34 was changed by changing the hole diameter D of the housing tubes 38a and 38b in relation to the groove diameter d of the mass 36. The amount of compression is calculated from the groove diameter d, hole diameter D, and wire diameter of the O-ring 34 using the following relationship. Crushing amount = (wire diameter × 2) - (hole diameter D - groove diameter d) (3)

[0032] As shown in Figure 8, in the experiment, five different hole diameters D were prepared for the housing tubes 38a and 38b, generating five different compression levels for each combination of mass 36 and O-ring 34. Since an O-ring 34 with a wire diameter of 3.5 mm was used, the maximum compression level was kept below 1.0 mm. For each pattern shown in Figure 8, the number of O-rings 34 was varied in three stages (2, 4, and 6), and the resonant frequency of the dynamic vibration absorber 30 was measured. Therefore, in the experiment, a total of 75 patterns of dynamic vibration absorbers 30 were fabricated, each consisting of five types of mass 36 (O-rings 34), five compression levels, and three quantities of O-rings 34, and the resonant frequencies were measured.

[0033] Figure 9 shows an example of the results of frequency analysis of acceleration. Figure 9(a) shows the excitation force of the impulse hammer. Figure 9(b) shows the detected values ​​of an acceleration sensor attached to a mass. The detected values ​​shown in Figure 9(b) are the acceleration detected when the dynamic vibration absorber 30 of pattern 1 shown in Figure 8 (mass: 610g, groove diameter d of mass: 22.03mm, hole diameter D of housing tube: 28.08mm, O-ring size: 21.7mm, O-ring wire diameter: 3.5mm, compression amount: 0.95mm) is vibrated. Figure 9(c) shows the analysis results obtained by performing an FFT (Fast Fourier Transformation) on the detected acceleration. The sampling frequency was 6.4 kHz, and the number of sampling points was 2048. This analysis result shows that the resonance frequency f of the dynamic vibration absorber 30 in pattern 1 was measured to be 309.375 [Hz].

[0034] The above frequency analysis was performed for all 75 patterns of dynamic vibration absorbers 30, and the obtained resonant frequency f and the mass m of the mass body 36 were substituted into equation (1) to calculate the spring constant k for each pattern.

[0035] Figures 10 to 14 show the relationship between the amount of compression and the spring constant derived for each O-ring size. The numerical values ​​shown at each point in the graph represent the resonant frequency f identified for each pattern. Equation (4) was used as a fitting function to express the relationship between the amount of compression and the spring constant, and the optimal coefficients a and b for the cases of 2, 4, and 6 O-rings were determined using the least squares method. (Spring constant) = a × (Amount of compression) b (4) The dashed lines in Figures 10 to 14 show the fitting curves for cases where there are 2, 4, and 6 O-rings 34.

[0036] Figure 10 shows the relationship between the compression amount and the spring constant derived for the P22A O-ring. The numerical values ​​shown at each point in the graph represent the resonant frequencies f identified in patterns 1 to 5 shown in Figure 8. From the graph in Figure 10, it can be seen that when using a P22A O-ring with a wire diameter of 3.5 mm, a dynamic vibration absorber 30 with a spring constant of up to approximately 0.8 [N / m] can be designed under the conditions of a compression amount of less than 1.0 mm and no more than 6 O-rings.

[0037] Figure 11 shows the relationship between the compression amount and the spring constant derived for the P29 O-ring. The numerical values ​​shown at each point in the graph represent the resonant frequencies f identified in patterns 6 to 10 shown in Figure 8. From the graph in Figure 11, it can be seen that when using a P29 O-ring with a wire diameter of 3.5 mm, a dynamic vibration absorber 30 with a spring constant of up to approximately 1.05 [N / m] can be designed under the conditions of a compression amount of less than 1.0 mm and no more than 6 O-rings.

[0038] Figure 12 shows the relationship between the compression amount and the spring constant derived for the P35 O-ring. The numerical values ​​shown at each point in the graph represent the resonant frequencies f identified in patterns 11 to 15 shown in Figure 8. From the graph in Figure 12, it can be seen that when using a P35 O-ring with a wire diameter of 3.5 mm, a dynamic vibration absorber 30 with a spring constant of up to approximately 1.15 [N / m] can be designed under the conditions of a compression amount of less than 1.0 mm and no more than 6 O-rings.

[0039] Figure 13 shows the relationship between the compression amount and the spring constant derived for the P41 O-ring. The numerical values ​​shown at each point in the graph represent the resonant frequencies f identified in patterns 16-20 shown in Figure 8. From the graph in Figure 13, it can be seen that when using a P41 O-ring with a wire diameter of 3.5 mm, a dynamic vibration absorber 30 with a spring constant of up to approximately 1.45 [N / m] can be designed under the conditions of a compression amount of less than 1.0 mm and no more than 6 O-rings.

[0040] Figure 14 shows the relationship between the compression amount and the spring constant derived for the P48 O-ring. The numerical values ​​shown at each point in the graph represent the resonant frequencies f identified in patterns 21 to 25 shown in Figure 8. From the graph in Figure 14, it can be seen that when using a P48 O-ring with a wire diameter of 3.5 mm, a dynamic vibration absorber 30 with a spring constant of up to approximately 1.6 [N / m] can be designed under the conditions of a compression amount of less than 1.0 mm and no more than 6 O-rings.

[0041] Next, Figures 15 to 17 show the results of linear interpolation for O-ring sizes that were not experimented with. Figure 15 shows the relationship between the amount of compression and the spring constant derived for two O-rings. The fitting curves for O-ring sizes "22", "29", "35", "41", and "48" are fitting curves derived from measured values, while the fitting curves shown between them are interpolation curves derived by linear interpolation for each O-ring size.

[0042] Figure 16 shows the relationship between compression amount and spring constant derived for four O-rings. The fitting curves for O-ring sizes "22," "29," "35," "41," and "48" are derived from measured values, while the fitting curves between them are interpolated curves derived by linear interpolation for each O-ring size.

[0043] Figure 17 shows the relationship between compression amount and spring constant derived for six O-rings. The fitting curves for O-ring sizes "22," "29," "35," "41," and "48" are derived from measured values, while the fitting curves between them are interpolated curves derived by linear interpolation for each O-ring size.

[0044] The relationships shown in Figures 10 to 17 demonstrate that as the amount of compression of the O-ring 34 increases, the spring constant increases according to the fitting function shown in equation (4). Furthermore, it is shown that the spring constant also increases when the number of O-rings 34 is increased and when their size is increased.

[0045] From the above experiments, a database can be created to derive multiple types of parameters for the dynamic vibration absorber 30 that generate a desired spring constant. As shown in Figures 10 to 17, the spring constant k is determined by the combination of the size of the O-ring 34, the number of O-rings 34, and the amount of compression. Therefore, if the size of the O-ring 34 is determined, the spring constant k is determined by the combination of the number of O-rings 34 and the amount of compression, and conversely, if the number of O-rings 34 is determined, the spring constant k is determined by the combination of the size of the O-rings 34 and the amount of compression.

[0046] For example, when designing a dynamic vibration absorber 30 in which the size of the O-ring 34 is fixed as P41 (inner diameter 40.7 mm) and the spring constant k is 0.4, referring to the relationship shown in Figure 13, the point at which each of the three fitting functions has a spring constant of 0.4 corresponds to the combination of the number of O-rings 34 and the amount of compression that achieves a spring constant of 0.4.

[0047] When two O-rings 34 are provided, the compression amount is set to approximately 0.71 in order to achieve a spring constant of 0.4 for the dynamic vibration absorber 30. When four O-rings 34 are provided, the compression amount is set to approximately 0.21 in order to achieve a spring constant of 0.4 for the dynamic vibration absorber 30. When six O-rings 34 are provided, the compression amount is set to approximately 0.07 in order to achieve a spring constant of 0.4 for the dynamic vibration absorber 30. In this embodiment, the database stores a fitting function derived from experiments, that is, a fitting function that defines the relationship between the compression amount and the spring constant for the size and number of O-rings 34. The database may also store the relationship between the compression amount and the spring constant in a format other than the fitting function. For example, the database may maintain a table representing the relationship between the amount of compression and the spring constant, where the spring constant is given in predetermined increments (e.g., 0.001 [N / m]) from 0 to the maximum value (e.g., the spring constant when the compression amount is 1 mm), and the amount of compression required to achieve each spring constant.

[0048] Figure 18 shows the functional blocks of the design support device 100 in an embodiment. The design support device 100 comprises an input unit 102, a processing unit 110, a holding unit 120, and an output device 130. The processing unit 110 has a receiving unit 112, an information acquisition unit 114, and an output unit 116.

[0049] The holding unit 120 stores a database containing information about the spring constants of dynamic vibration absorbers generated based on information identified through experiments. Specifically, the holding unit 120 stores a database that defines the relationships between multiple types of parameters of dynamic vibration absorbers that generate spring constants within a predetermined range.

[0050] The retaining unit 120 may store fitting functions derived from experiments, that is, fitting functions that define the relationship between the amount of compression and the spring constant for the size and number of O-rings. Furthermore, as explained with respect to Figures 15 to 17, the retaining unit 120 may also store fitting functions for O-ring sizes generated by interpolating the fitting functions derived from experiments. By interpolating the fitting functions derived from experiments to generate fitting functions for O-ring sizes that have not been experimented with, the effort of experimenting with all sizes of O-rings can be reduced. In order to obtain a highly accurate fitting function for each size of O-ring, it is preferable to experiment with all sizes of O-rings and derive fitting functions that accurately define the relationship between the amount of compression and the spring constant for all sizes of O-rings.

[0051] As described above, the retaining unit 120 may maintain the relationship between the amount of compression and the spring constant in a format other than the fitting function. For example, the retaining unit 120 may maintain a table representing the correspondence between spring constants in predetermined increments (e.g., 0.001 [N / m]) from 0 to a maximum value (e.g., the spring constant when the amount of compression is 1 mm) and the amount of compression required to achieve each spring constant, as the relationship between the amount of compression and the spring constant. In this case, the retaining unit 120 may maintain in a table format the values ​​of the amount of compression for spring constants in increments of 0.001 [N / m], such as the value of the amount of compression for spring constants achieving a spring constant of 0.001 [N / m], the value of the amount of compression for spring constants achieving a spring constant of 0.002 [N / m], and the value of the amount of compression for spring constants achieving a spring constant of 0.003 [N / m]. The retaining unit 120 may maintain in a table format the values ​​of the amount of compression for spring constants in increments of 0.001 [N / m] within the range of spring constants that are actually required.

[0052] In this embodiment, the retaining part 120 maintains the relationship between the O-ring size, the amount of O-ring compression, and the number of O-rings and the spring constant. However, it is known that the spring constant changes nonlinearly with respect to the amount of compression, while it changes roughly linearly with respect to the size and number of O-rings. Therefore, the retaining part 120 only needs to maintain the relationship between the spring constant and the amount of O-ring compression for a reference O-ring size (e.g., P35) and a reference number (e.g., 2), and does not need to maintain the relationship between the spring constant and the amount of compression for the sizes and numbers of other O-rings. In this case, when determining the amount of compression for the size and number of O-rings, the amount of compression can be determined by multiplying the spring constant by the ratio to the reference size and the ratio to the reference number.

[0053] Figure 19 shows a flowchart of the design support method for a dynamic vibration absorber in an embodiment. First, the designer mounts the pre-redesign machining tool 3 to the machine tool 1 that will actually be used, and then excites the machining tool 3 with an impulse hammer or the like. The designer then performs a frequency analysis on the values ​​detected by the acceleration sensor attached to the machining tool 3 to measure the resonant frequency f (S10).

[0054] Figure 1 shows the machining tool 3 before redesign, mounted on the machine tool 1 that will actually be used. The discloser measured the resonant frequencies of machining tools of various shapes mounted on machine tools and found that the resonant frequencies of low-rigidity machining tools (e.g., machining tools with a long length in the direction of the rotation axis) are less affected by the machine tool, while the resonant frequencies of high-rigidity machining tools (e.g., short machining tools) are more easily affected by the machine tool. Therefore, especially in the case of short machining tools, it was found that when mounted on a different machine tool, the resonant frequency of the machining tool changes due to the influence of the machine tool to which it is mounted. Accordingly, in this embodiment, the resonant frequency f of the machining tool 3 is measured with the machining tool 3 mounted on the machine tool 1 that will actually be used.

[0055] Figure 20 shows a hypothetical hole 7 in the machining tool 3 before redesign, and a hypothetical mass body 26 to be housed in the hole 7. The designer designs the approximate shape of the hole 7 that can be formed within a range that does not significantly reduce the second moment of area of ​​the tool body 4. Here, the designer tentatively determines the hole diameter D of the hole 7 and sets the groove diameter d of the mass body 26 to be slightly smaller than the hole diameter D. In this embodiment, the groove diameter d of the mass body 26 and the O-ring to be fitted into the annular groove of groove diameter d are treated as a set, so for example, the size of the O-ring available in the P standard can be determined from the tentatively determined hole diameter D, and the corresponding groove diameter d can be determined. In this way, the designer determines the groove diameter d of the mass body 26 and the O-ring to be used from the shape of the hole 7 that can be formed in the machining tool 3 (S12).

[0056] Once the resonant frequency f of the machining tool 3, the groove diameter d of the mass body 26, and the O-ring to be used are determined, the designer inputs information specifying the resonant frequency f, groove diameter d, and the size of the O-ring to be used from the input unit 102 of the design support device 100 (S14). If the processing unit 110 can automatically calculate the groove diameter d for the O-ring size, the designer may omit inputting the groove diameter d. The information specifying the size of the O-ring to be used may be information indicating the size of the O-ring itself (including the inner diameter and wire diameter), or it may be the standard name (nominal number) of the O-ring; in any case, it is sufficient if the information allows the processing unit 110 to identify the size of the O-ring (including the inner diameter and wire diameter). The receiving unit 112 receives the resonant frequency f, groove diameter d, and information specifying the size of the O-ring to be used.

[0057] The information acquisition unit 114 derives the relationship between mass m and spring constant k from the received resonant frequency f by referring to equation (1). Since the target resonant frequency f is specified, the relationship between mass m and spring constant k is determined as follows. k / m = (2π·f) 2 (5)

[0058] The information acquisition unit 114 reads a fitting function from the holding unit 120 that defines the relationship between the amount of compression and the spring constant for the size of the received O-ring, and obtains the relationship between the mass m that satisfies equation (5) and the amount of compression required to generate the spring constant k. In other words, the information acquisition unit 114 refers to the relationship between the O-ring size and the amount of compression of the O-ring with respect to the spring constant, which is held in the holding unit 120, and obtains information regarding the amount of compression of the O-ring in a dynamic vibration absorber having a resonant frequency that matches or approximates the resonant frequency of the processing tool 3, and information regarding the mass m of the mass body (S16). The output unit 116 outputs the information regarding the amount of compression of the O-ring and the information regarding the mass m of the mass body to the output device 130 (S18). The output device 130 may be a display device.

[0059] As described above, the holding unit 120 stores fitting functions corresponding to the number of O-rings (2, 4, or 6). Therefore, the information acquisition unit 114 acquires information regarding the amount of O-ring compression and the mass m of the mass body, corresponding to the number of O-rings, and the output unit 116 outputs information regarding the amount of O-ring compression and the mass m of the mass body, corresponding to the number of O-rings.

[0060] Figure 21 shows an example of information displayed on the output device 130. The output unit 116 displays information regarding the amount of O-ring compression and information regarding the mass m of the mass body in graph form. In this example, the information regarding the amount of O-ring compression is displayed as information regarding the hole diameter D of the hole 7 formed in the machining tool 3. From equation (3), the hole diameter D is, Hole diameter D = Groove diameter d + (Wire diameter × 2) - Crushing amount (6) It is calculated as follows.

[0061] In the example shown in Figure 21, the relationship curves between the hole diameter D and mass m are displayed for the cases of 2, 4, and 6 O-rings. The combinations of hole diameter D and mass m on this line realize the resonant frequency f of the dynamic vibration absorber. For example, if 4 O-rings are used and the hole diameter D is 50.5 mm, the mass m of the mass body should be approximately 600 g. In this way, the processing unit 110 supports the design of a dynamic vibration absorber having a resonant frequency that matches or approximates the resonant frequency f of the processing tool 3 by presenting the designer with a graph that represents the combinations of hole diameter D and mass m that realize a resonant frequency that matches the resonant frequency f of the processing tool 3. Since this dynamic vibration absorber uses standard O-rings, the manufacturing cost can be kept low.

[0062] The processing unit 110 may, instead of displaying the graph on the output device 130, display on the output device 130 the combination of one or more hole diameters D and masses m on the graph. In the embodiment, the case where the hole portion 7 is not formed in the processing tool 3 has been described, but if a cylindrical hole portion 7 is already formed in the processing tool 3, the information acquisition unit 114 may acquire information about the groove diameter d of the mass body as information about the amount of crushing. If the hole diameter D is fixed, the groove diameter d is calculated by the following formula. Groove diameter d = Hole diameter D - (Wire diameter × 2) + Crushing amount (7) In this case, the output unit 116 may display a curve representing the combination of the groove diameter d of the mass body and the mass m of the mass body that realizes the resonant frequency f, for each number of O-rings.

[0063] Figure 22 shows experimental results demonstrating the effectiveness of the dynamic vibration absorber designed using the design support method of the embodiment. Here, an impulse response analysis was performed using the excitation force measurement results of the impulse hammer and the measurement results of an acceleration sensor attached to the machining tool, and the compliance transfer function of the machining tool was obtained and illustrated. As shown in Figure 22, it can be seen that the peak value around 350 Hz was significantly improved by attaching the dynamic vibration absorber to the machining tool.

[0064] As described above, according to this embodiment, by pre-measuring the spring constant k of a dynamic vibration absorber comprising a mass body and elastic bodies held on the outer surface of the mass body in combination with the size of the elastic bodies, the number of elastic bodies, and the amount of compression (compression allowance) of the elastic bodies, it becomes possible to easily design a dynamic vibration absorber having a target resonant frequency f. Furthermore, as mentioned above, by utilizing the fact that the spring constant k changes linearly (proportionally) with respect to the size and number of elastic bodies, it is also possible to design the dynamic vibration absorber more simply by pre-measuring the relationship between the spring constant k of the elastic bodies and the amount of compression with respect to the size and number of reference elastic bodies.

[0065] The present disclosure has been described above based on embodiments. These embodiments are illustrative, and it will be understood by those skilled in the art that various modifications are possible in combinations of their components and processing processes, and that such modifications are also within the scope of the present disclosure. In the embodiments, the elastic body used as a spring element was described as an O-ring that is compressed between two cylindrical surfaces (the inner circumferential surface of the hole 7 and the bottom surface of the annular groove 22). However, the design support method of the embodiments can also be suitably applied to the design of dynamic vibration absorbers that include elastic body structures other than O-rings compressed between two cylindrical surfaces, in which the relationship between force and displacement exhibits nonlinear behavior.

[0066] In the embodiment, an example was shown in which a mass body 26 is housed inside the machining tool 3 and an O-ring, which is an elastic body 24, is held in an annular groove 22 formed on the outer circumferential surface of the mass body 26. However, if the machining tool 3 is an elongated tool, the mass body 26 may be provided so as to surround the machining tool 3. In this case, a hole may be formed in the mass body 26, the machining tool 3 may be housed in the hole of the mass body 26, and the O-ring, which is an elastic body 24, may be held in an annular groove 22 formed on the outer circumferential surface of the machining tool 3.

[0067] The following modified example describes a structure in which the resonant frequency of a dynamic vibration absorber can be adjusted. Figure 23 shows an example of a dynamic vibration absorber with an adjustable resonant frequency. The modified dynamic vibration absorber 50 comprises a substantially cylindrical mass body 52 and a plurality of elastic bodies 54a, 54b, 54c, 54d (hereinafter referred to as "elastic bodies 54" unless otherwise specified) that function as spring elements. The elastic bodies 54 are components having nonlinear rigidity with respect to deformation and may be O-rings, which are annular rubber packings with a circular cross-section. A plurality of annular grooves 56a, 56b, 56c, 56d (hereinafter referred to as "annular grooves 56" unless otherwise specified) are formed on the outer circumferential surface of the mass body 52, and the elastic bodies 54 are held in the annular grooves 56.

[0068] The dynamic vibration absorber 50 is housed in a case body 60, which is a component (vibrating body) surrounding the mass body 52. ​​The case body 60 may be a tool body 4 with a hole 7 formed therein, and the dynamic vibration absorber 50 may be housed in the hole 7 of the tool body 4. The dynamic vibration absorber 50 shown in Figure 23 has four elastic bodies 54, but the number of elastic bodies 54 is not limited to four; there may be more. In other words, it is preferable for the dynamic vibration absorber 50 to have two or more elastic bodies 54, and the more elastic bodies 54 there are, the wider the range of adjustment for the resonance frequency can be. In Figure 23, an O-ring is shown as an elastic body 54 held on the outer circumferential surface of the mass body 52, but elastic bodies 54 other than O-rings may be held at the longitudinal end of the mass body 52.

[0069] The adjustment structure 40 includes a pressing section for independently changing the amount of compression of each of the multiple elastic bodies 54. In the adjustment structure 40 shown in Figure 23, elastic pieces 58 capable of pressing the elastic bodies 54 are provided on the inside of the case body 60 at positions opposite to the elastic bodies 54. In this example, elastic pieces 58a and 58b are provided opposite to elastic body 54a, elastic pieces 58c and 58d are provided opposite to elastic body 54b, elastic pieces 58e and 58f are provided opposite to elastic body 54c, and elastic pieces 58g and 58h are provided opposite to elastic body 54d. Each elastic piece 58 is provided with an adjustment screw 62 for adjusting the amount of movement of the elastic piece 58. The pressing section consists of an elastic piece 58 and an adjustment screw 62, and is provided for each elastic body 54. Note that the elastic piece 58 maintains a position where it does not contact the elastic body 54 unless it is pushed radially inward by the adjustment screw 62.

[0070] In the state shown in Figure 23, the adjustment screws 62a and 62b are pushed radially inward, and the elastic pieces 58a and 58b are pressing against the elastic body 54a. Similarly, the adjustment screws 62c and 62d are pushed radially inward, and the elastic pieces 58c and 58d are pressing against the elastic body 54b, and the adjustment screws 62g and 62h are pushed radially inward, and the elastic pieces 58g and 58h are pressing against the elastic body 54d. On the other hand, the adjustment screws 62e and 62f are not pushed radially inward, and the elastic pieces 58e and 58f maintain a position where they do not contact the elastic body 54c, and do not press against the elastic body 54c (the amount of compression of the elastic body 54c is zero). Thus, with the adjustment structure 40, the amount of compression of the elastic body 54 can be independently adjusted by the amount of indentation of the adjustment screw 62, which enables fine adjustment of the spring constant k of the dynamic vibration absorber 50, and therefore fine adjustment of the resonant frequency f. Furthermore, since the adjustment structure 40 allows independent adjustment of the amount of compression of three or more elastic bodies 54, the spring constant k can be adjusted over a wide range, and therefore the resonant frequency f can be set over a wide range.

[0071] In the example shown in Figure 23, the elastic body 54 is pressed from two directions that are 180 degrees apart from each other, and it is preferable that these pressing directions be aligned with the cutting direction that affects the regenerative chatter vibration. Alternatively, the other side may be fixed to create a structure that presses from only one side. The adjustment structure 40 may also press the elastic body 54 from four directions that are 90 degrees apart. Furthermore, in this example, an annular groove 56 is formed only on the outer circumferential surface of the mass body 52 to hold the O-ring in the longitudinal direction, but annular grooves may also be formed on the inner circumferential surface of each elastic piece 58 to hold the mass body 52 in the longitudinal direction.

[0072] In the above example, the dynamic vibration absorber 50 is housed in a case body 60 which is a component, but the dynamic vibration absorber 50 may be configured to surround a cylindrical component. In this case, the annular groove 56 may be formed in the component, and the pressing portion may be provided on the dynamic vibration absorber 50.

[0073] Furthermore, although the above example explained that the pressing unit can independently change the amount of compression of each of the multiple elastic bodies 54, it is sufficient for the pressing unit to be able to independently change the amount of compression of at least one elastic body 54. By allowing the pressing unit to independently change the amount of compression of one or more elastic bodies 54, the resonant frequency f can be set appropriately. In addition, the number of elastic bodies 54 may be two, but by using three or more, it becomes possible to set the resonant frequency f over a wide range.

[0074] Figure 24 shows another example of a dynamic vibration absorber with an adjustable resonant frequency. The modified dynamic vibration absorber 80 comprises a substantially cylindrical mass body 82 and a plurality of elastic bodies 84a, 84b, 84c, 84d (hereinafter referred to as "elastic bodies 84" unless otherwise specified) that function as spring elements. The elastic bodies 84 are components having nonlinear rigidity with respect to deformation and may be O-rings, which are annular rubber packings with a circular cross-section. A plurality of annular grooves 86a, 86b, 86c, 86d (hereinafter referred to as "annular grooves 86" unless otherwise specified) are formed on the outer circumferential surface of the mass body 82, and the elastic bodies 84 are held in the annular grooves 86.

[0075] The dynamic vibration absorber 80 is housed in a case body 90, which is a component that surrounds the mass body 82. The case body 90 may be a tool body 4 with a hole 7 formed therein, and the dynamic vibration absorber 80 may be housed in the hole 7 of the tool body 4. The dynamic vibration absorber 80 shown in Figure 24 has four elastic bodies 84, but the number of elastic bodies 84 is not limited to four; there may be more. In other words, it is preferable for the dynamic vibration absorber 80 to have two or more elastic bodies 84, and the more elastic bodies 84 there are, the wider the range of adjustment for the resonance frequency can be. In Figure 24, an O-ring is shown as an elastic body 84 held on the outer circumferential surface of the mass body 82, but elastic bodies 84 other than O-rings may be held at the longitudinal end of the mass body 82.

[0076] The adjustment structure 70 includes pressing parts for independently changing the amount of compression of each of the multiple elastic bodies 84. In the adjustment structure 70 shown in Figure 24, tapered portions 88 capable of pressing the elastic bodies 54 are provided on the inside of the case body 90 at positions facing the elastic bodies 54. In this example, tapered portions 88a are provided at positions facing elastic body 54a, tapered portions 88b are provided at positions facing elastic body 54b, tapered portions 88c are provided at positions facing elastic body 54c, and tapered portions 88d are provided at positions facing elastic body 54d. The case body 90 is provided with adjustment screws 92a and 92b for adjusting the amount of axial movement of the mass body 82. The tapered portions 88, which are pressing parts, are provided for each elastic body 84. Note that the adjustment screw 92b may be omitted if the longitudinal component of the elastic force applied to the tapered portion 88 is large. Furthermore, the adjustment structure 70 only needs to be capable of adjusting the longitudinal position of the mass body 82, and may be a structure that pulls with a wire or the like instead of pushing with an adjustment screw.

[0077] The tapered portions 88a and 88d have inclined surfaces that are tilted radially inward at an angle α with respect to the axial direction. The tapered portions 88b and 88c also have inclined surfaces that are tilted radially inward at an angle β with respect to the same axial direction. The angles α and β may be the same or different, but it is preferable that the inclination angles of all tapered portions 88 are in the same direction with respect to the axial direction. When the length of the tapered portion 88a is La, La × sinα is the maximum compression amount of the elastic body 84a. Similarly, when the length of the tapered portion 88b is Lb, Lb × sinβ is the maximum compression amount of the elastic body 84b. As described in the embodiment, for example, when using an O-ring with a wire diameter of 3.5 mm, it is preferable that the maximum compression amount be designed to be less than 1 mm.

[0078] In the state shown in Figure 24, neither tapered portion 88 is pressing against the elastic body 84. In the adjustment structure 70, the adjustment screw 92a is pushed in to the right in the axial direction, and the adjustment screw 92b is pulled out to the right in the axial direction, causing the case body 90 to move relative to the dynamic vibration absorber 80, so that the tapered portion 88 presses against the elastic body 84. At this time, the tapered portion 88 deforms the elastic body 84 by an amount of compression corresponding to the amount of movement.

[0079] Figure 25 shows a state in which some of the elastic bodies 84 are pressed by the tapered portion 88. In this state, elastic body 84a is pressed by the tapered portion 88a, and elastic body 84d is pressed by the tapered portion 88d. Elastic bodies 84b and 84c are not pressed. Thus, in the adjustment structure 70, the multiple tapered portions 88 are formed so that they begin pressing the elastic bodies 84 at different timings.

[0080] Figure 26 shows the state in which all elastic bodies 84 are pressed by the tapered portion 88. This state represents the state in which the adjustment screw 92a is pushed further to the right in the axial direction from the state shown in Figure 25. In the state shown in Figure 26, elastic body 84a is pressed by the flat portion 89a, elastic body 84b is pressed by the tapered portion 88b, elastic body 84c is pressed by the tapered portion 88c, and elastic body 84d is pressed by the flat portion 89d.

[0081] Thus, with the adjustment structure 70, the amount of compression of the elastic body 84 can be independently adjusted by the shape (angle) of the tapered portion 88 and the position where the tapered portion 88 is provided, making it possible to fine-tune the spring constant k of the dynamic vibration absorber 80, and therefore fine-tune the resonance frequency f. Furthermore, since the adjustment structure 70 allows for independent adjustment of the compression amounts of multiple elastic bodies 84, the spring constant k can be adjusted over a wide range, and therefore the resonance frequency f can be set over a wide range.

[0082] In the above example, the dynamic vibration absorber 80 is housed in a case body 90, which is a component. However, the dynamic vibration absorber 80 may be configured to surround a cylindrical component. In this case, the annular groove 86 may be formed in the component, and the pressing portion may be provided on the dynamic vibration absorber 80.

[0083] Furthermore, while the above example explained that the pressing unit can independently change the amount of compression of each of the multiple elastic bodies 84, it is sufficient for the pressing unit to be able to independently change the amount of compression of at least one elastic body 84. By allowing the pressing unit to independently change the amount of compression of one or more elastic bodies 84, the resonant frequency f can be set appropriately. In addition, the number of elastic bodies 84 may be two, but by using three or more, it becomes possible to set the resonant frequency f over a wide range.

[0084] The details of the nature of this disclosure are as follows: A design support method in one aspect of the present disclosure is a method for designing a dynamic vibration absorber to be attached to a machining tool, comprising a mass body and an elastic body held between the mass body and the machining tool, the method comprising: a reception step of receiving the resonant frequency of the machining tool; an acquisition step of obtaining at least one of the following information regarding the amount of compression of the elastic body, the size of the elastic body, the number of elastic bodies, and the mass of the mass body in a dynamic vibration absorber having a resonant frequency that matches or approximates the received resonant frequency of the machining tool, by referring to the relationship between the spring constant and the amount of compression of the elastic body held in the holder; and an output step of outputting the acquired information.

[0085] According to this embodiment, it becomes possible to suitably support the design of a dynamic vibration absorber suitable for a machining tool.

[0086] The reception step receives at least one of the following: information specifying the amount of compression of the elastic body, information specifying the size of the elastic body, information specifying the number of elastic bodies, and information specifying the mass of the mass body. The acquisition step refers to the relationship between the spring constant and the amount of compression of the elastic body held in the holding part and acquires information from among the information specifying the amount of compression of the elastic body, information specifying the size of the elastic body, information specifying the number of elastic bodies, and information specifying the mass of the mass body that was not received by the reception step. The output step may output the acquired information.

[0087] For example, if the reception step receives information specifying the amount of compression of the elastic body and information specifying the size of the elastic body, the acquisition step acquires information regarding the number of elastic bodies and information regarding the mass of the mass body. Also, if the reception step receives information specifying the mass of the mass body, the acquisition step acquires information regarding the amount of compression of the elastic body, information regarding the size of the elastic body, and information regarding the number of elastic bodies. Also, if the reception step receives information specifying the size of the elastic body, the acquisition step acquires information regarding the amount of compression of the elastic body, information regarding the number of elastic bodies, and information regarding the mass of the mass body. In this way, by having the acquisition step acquire information about information not received by the reception step, and the output step output the acquired information, it becomes possible to suitably support the design of a dynamic vibration absorber suitable for the machining tool.

[0088] The holding unit maintains the relationship between the size of the elastic body, the amount of compression of the elastic body, and the number of elastic bodies with respect to the spring constant. The acquisition step may acquire information regarding the amount of compression of the elastic body and information regarding the mass of the mass body, corresponding to the size and number of elastic bodies. This can suitably support the design of a dynamic vibration absorber, including the number of elastic bodies. The reception step may accept information specifying the size of the elastic bodies to be used.

[0089] The elastic body is an O-ring, held in an annular groove formed in a mass or machining tool. The receiving step receives the groove diameter d of the annular groove, the acquisition step acquires information regarding the hole diameter D of the hole to be formed in the machining tool or mass as information regarding the amount of compression of the elastic body, and the output step may output information regarding the hole diameter D of the hole as information regarding the amount of compression of the elastic body. By outputting information regarding the hole diameter D of the hole as information regarding the amount of compression of the elastic body, information about the hole can be provided to the designer.

[0090] The resonant frequency of the machining tool is preferably the resonant frequency obtained with the machining tool attached to the machine tool. Obtaining the resonant frequency of the machining tool with the machining tool attached to the machine tool in use can effectively support the design of a dynamic vibration absorber suitable for that machine tool.

[0091] Another aspect of the present disclosure is a dynamic vibration absorber to be attached to a component, comprising a mass body and two or more elastic bodies held between the mass body and the component, the dynamic vibration absorber having an adjustable resonant frequency, and comprising a pressing section for independently changing the amount of compression of at least one of the elastic bodies.

[0092] According to this embodiment, the resonant frequency can be suitably adjusted by independently changing the amount of compression of at least one elastic body.

[0093] The dynamic vibration absorber comprises three or more elastic bodies, and the pressing section may allow for independent adjustment of the compression amount of each of the three or more elastic bodies. By allowing for independent adjustment of the compression amount of each of the three or more elastic bodies, it becomes possible to adjust the resonant frequency over a wide range.

[0094] A pressing portion may be provided for each elastic body. The pressing portion is formed on a mass or component, and when the component moves relative to the mass, the pressing portion may deform the elastic body by an amount of compression corresponding to the amount of movement.

[0095] A further aspect of the present disclosure is a machine tool to which a machining tool is attached, the machining tool having a dynamic vibration absorber comprising a mass body and two or more elastic bodies held between the mass body and the machining tool, the machining tool comprising an adjustment structure capable of adjusting the resonant frequency of the dynamic vibration absorber by independently changing the amount of compression of at least one elastic body. [Explanation of symbols]

[0096] 1...Machine tool, 2...Spindle, 3...Processing tool, 4...Tool body, 5...Blade section, 6...Mounting section, 7...Hole section, 8...Large diameter section, 9...Small diameter section, 10...Lid, 20...Dynamic vibration absorber, 22...Annular groove, 24...Elastic body, 26...Mass body, 28...Thrust bearing, 30...Dynamic vibration absorber, 32...Annular groove, 34...O-ring, 36...Mass body, 38a,38b...Housing tube, 40...Adjustment structure, 50...Dynamic vibration absorber, 52...Mass body, 54...Elastic body, 56...Annular groove, 58...Elastic piece, 60...Case body, 62...Adjustment screw, 70...Adjustment structure, 80...Dynamic vibration absorber, 82...Mass body, 84...Elastic body, 86...Annular groove, 88...Tapered section, 89a, 89d...Flat section, 90...Case body, 92a, 92b...Adjustment screw, 100...Design support device, 102...Input section, 110...Processing section, 112...Reception section, 114...Information acquisition section, 116...Output section, 120...Holding section, 130...Output device.

Claims

1. A method for designing a dynamic vibration absorber to be attached to a machining tool, the dynamic vibration absorber comprising a mass body and an elastic body held between the mass body and the machining tool, A receiving step that receives information specifying the resonant frequency of the processing tool and the size of the elastic body to be used, An acquisition step to obtain information regarding the amount of compression of the elastic body, information regarding the number of elastic bodies, and information regarding the mass of the mass body in the dynamic vibration absorber having a resonant frequency that matches or approximates the resonant frequency of the received processing tool, by referring to the relationship between the size of the elastic body, the amount of compression of the elastic body, and the number of elastic bodies with respect to the spring constant of the elastic body held in the holding part, An output step that outputs the acquired information, A method for supporting the design of a dynamic vibration absorber, characterized by including the following:

2. The acquisition step acquires information relating to the amount of compression of the elastic body and the mass of the mass body, corresponding to the size and number of the elastic body. The output step displays information regarding the amount of compression of the elastic body and information regarding the mass of the mass body in the form of a graph. The design support method for a dynamic vibration absorber according to feature 1.

3. The acquisition step acquires information regarding the diameter of the hole to be formed in the processing tool or the mass body as information regarding the amount of compression of the elastic body. The design support method for a dynamic vibration absorber according to feature 2.

4. The elastic body is an O-ring, which is held in an annular groove formed in the mass body or the processing tool. The reception step receives the groove diameter d of the annular groove, The acquisition step involves acquiring information regarding the amount of compression of the elastic body, specifically information regarding the hole diameter D of the hole formed in the processing tool or the mass body. The output step outputs information regarding the hole diameter D of the hole as information regarding the amount of compression of the elastic body. The design support method for a dynamic vibration absorber according to feature 1.

5. The resonant frequency of the aforementioned machining tool is the resonant frequency obtained while the machining tool is attached to the machine tool. The design support method for a dynamic vibration absorber according to feature 1.

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