Damping tool shank and working method

Abstract

The application provides a damping tool shank and a working method, relates to the field of machine tools, and aims at the poor adaptability of the current passive damping device. A damping assembly carrying a local resonance phononic crystal structure is installed on the shank. The vibration generated in the cutting process is transmitted to the annular shell through the shank, then conducted to the base beam, and finally drives each vibrator group to vibrate. The position change of the vibrator group changes the parameters of the local resonance phononic crystal structure, and then adjusts the frequency range of the vibration suppression. When the rotating speed increases, the centrifugal force increases, the distance between the second vibrator and the first vibrator decreases, and the band gap moves to high frequency. When the rotating speed decreases, the band gap moves to low frequency, the adaptive matching between the band gap and the cutting flutter frequency is realized, the cutting vibration is suppressed, the band gap range is automatically adjusted through the rotating speed change, the flutter frequency under different working conditions is matched, and the adaptability is improved.

Description

Technical Field

[0001] This invention relates to the field of machine tools, and more specifically to a vibration-damping tool holder and its working method. Background Technology

[0002] Machining involves removing material from a workpiece surface using cutting tools. Machining vibrations can cause dimensional and surface quality defects, and even lead to part scrap. Current methods for vibration suppression in machining include passive vibration control and active / semi-active vibration control. Active / semi-active vibration control is highly targeted but requires additional electrical components such as sensors and actuators, making the system more complex and less stable. In passive vibration control, optimizing cutting parameters for suppression significantly limits the selection of machining parameters. Passive vibration suppression devices, which increase system stiffness, damping, or add passive vibration absorbers, offer higher reliability and are commonly used in machining applications.

[0003] Current passive vibration damping devices are only suitable for specific working conditions. After adjustment and installation on the tool holder, their performance parameters are fixed values. They can only play a passive vibration damping role under specific machining parameters and are difficult to adapt to the vibration damping requirements of different machining parameters in various machining scenarios. It is also difficult to make corresponding adjustments according to the machining parameters. When the working conditions change, the passive vibration damping device needs to be adjusted to change the stiffness or mass of the vibration damping structure to adapt to the adjusted working conditions. However, the entire adjustment process is relatively complicated and requires the machine to be stopped for adjustment, which is difficult to meet the cutting needs when the working conditions change frequently. Summary of the Invention

[0004] The purpose of this invention is to address the shortcomings of existing technologies by providing a vibration-damping tool holder and its working method. A vibration-damping component with a localized resonant phonon crystal structure is installed on the shank. Vibrations generated during cutting are transmitted through the shank to the annular housing, then to the base beam, ultimately causing each oscillator group to vibrate. Changes in the position of the oscillator groups alter the parameters of the localized resonant phonon crystal structure, thereby adjusting its frequency range for suppressing vibration. When the rotational speed increases, the centrifugal force increases, the distance between the second and first oscillators decreases, and the bandgap shifts to higher frequencies. When the rotational speed decreases, the bandgap shifts to lower frequencies, achieving adaptive matching between the bandgap and the cutting chatter frequency. Vibration energy within the bandgap range is absorbed or blocked by the localized resonance effect, preventing further transmission to the shank and the tool, thus suppressing cutting vibration.

[0005] The first objective of this invention is to provide a vibration-damping tool holder, which adopts the following solution:

[0006] include:

[0007] The shank has a mounting end for mounting the cutting tool.

[0008] The vibration damping assembly includes an annular shell and a local resonant phonon crystal structure located within the annular shell. The shell is sleeved and connected to a handle. The local resonant phonon crystal structure includes an annular base beam and multiple oscillator groups. The base beam is coaxially fixed to the outer ring of the annular shell. The multiple oscillator groups are evenly distributed along the circumferential direction of the annular shell. Each oscillator group includes a connecting rod, a first oscillator, and a second oscillator. The connecting rod is radially distributed along the base beam and one end is connected to the base beam. The first oscillator is fixed to the connecting rod. The second oscillator is located on the side of the first oscillator facing the axis of the base beam and slides axially with the connecting rod. The first oscillator and the second oscillator repel each other.

[0009] Furthermore, two sets of localized resonant phonon crystal structures are installed inside the annular shell, and are distributed sequentially upward along the axial direction of the annular shell.

[0010] Furthermore, one end of the connecting rod is connected to the base beam to form a cantilever structure, and the end of the connecting rod away from the base beam forms a limiting block. A sliding space for the second oscillator is formed between the limiting block and the first oscillator.

[0011] Furthermore, the first oscillator is sleeved and fixed outside the connecting rod, and the second oscillator is slidably sleeved outside the connecting rod. The first and second oscillators corresponding to the same oscillator group are coaxially distributed.

[0012] Furthermore, the first oscillator includes a housing and a magnet. The housing is fitted and fixed to the outside of the connecting rod and forms an annular groove for accommodating the magnet, which is embedded in the annular groove.

[0013] Furthermore, the second oscillator is a ring magnet, with one end of the first oscillator and one end of the second oscillator facing each other.

[0014] Furthermore, the annular shell and the handle are coaxially distributed, and the inner ring of the annular shell is attached to and connected to the handle.

[0015] Furthermore, the segment in which the connecting rod slides into the second oscillator is the optical axis segment.

[0016] A second objective of the present invention is to provide a method for operating a vibration-damping tool holder, utilizing the vibration-damping tool holder provided as in the first objective, comprising:

[0017] The cutting tool is installed onto the mounting end of the shank and then mounted as a whole onto the tool holder;

[0018] The cutting tool generates vibration during the cutting process. The vibration is first transmitted to the annular shell, and then transmitted to the annular base beam fixed to the outer ring of the shell.

[0019] Initially, the second oscillator is located at the end of the connecting rod due to the magnetic repulsion of the first oscillator;

[0020] When the damping tool holder starts to rotate, the second oscillator will be subjected to centrifugal force, which will cause the second oscillator to slide along the connecting rod towards the first oscillator. As the second oscillator gradually approaches the first oscillator, the magnetic repulsion between the two will gradually increase until the centrifugal force and the magnetic repulsion reach a balance, and the second oscillator will stabilize at this specific position.

[0021] The vibration damping components generate local resonance to absorb the vibration energy transmitted from the handle and suppress vibration.

[0022] Furthermore, when the rotational speed of the damping tool holder changes, the position of the second oscillator on the connecting rod changes, altering the local resonance effect of the damping component to adapt it to the vibration at that rotational speed.

[0023] Compared with the prior art, the advantages and positive effects of this invention are:

[0024] To address the poor adaptability of current passive vibration damping devices, a vibration damping component equipped with a local resonant phonon crystal structure is installed on the shank. Vibrations generated during cutting are transmitted through the shank to the annular shell, then to the base beam, ultimately driving the vibration of each oscillator group. Changes in the position of the oscillator group alter the parameters of the local resonant phonon crystal structure, thereby adjusting its frequency range for suppressing vibration. As the rotational speed increases, the centrifugal force increases, the distance between the second and first oscillators decreases, and the bandgap shifts to higher frequencies; as the rotational speed decreases, the bandgap shifts to lower frequencies, achieving adaptive matching between the bandgap and the chatter frequency. Vibration energy within the bandgap range is absorbed or blocked by the local resonance effect, preventing further transmission to the shank and tool, thus suppressing cutting vibration. By automatically adjusting the bandgap range through changes in rotational speed, it matches the chatter frequency under different working conditions, improving its adaptability.

[0025] Two sets of locally resonant phonon crystal structures are sequentially distributed along the axial direction within the annular shell to strengthen the vibration damping components. The two sets of structures form a superposition effect in the axial direction, covering a wider range of vibration frequencies. When cutting vibrations are transmitted to the annular shell, the two sets of structures can act on vibrations in different frequency bands respectively, synergistically suppressing vibrations through their respective locally resonant band gaps, improving the overall comprehensiveness of the vibration damping effect, and avoiding the vibration damping blind spots that may exist in a single structure at a specific frequency band.

[0026] The annular shell and the handle are coaxially distributed, with the inner ring fitted and connected to the handle. This coaxial distribution ensures a symmetrical vibration transmission path from the handle to the annular shell, avoiding uneven energy distribution caused by eccentricity. This allows each oscillator assembly to be evenly stressed, achieving a synergistic vibration damping effect. The fitted inner ring enhances the connection rigidity, reduces energy loss during vibration transmission, and ensures that more vibration energy is absorbed by the internal localized resonant phonon crystal structure, improving damping efficiency. Attached Figure Description

[0027] The accompanying drawings, which form part of this invention, are used to provide a further understanding of the invention. The illustrative embodiments of the invention and their descriptions are used to explain the invention and do not constitute an improper limitation of the invention.

[0028] Figure 1 This is a schematic diagram of a vibration damping tool holder in one or more embodiments of the present invention.

[0029] Figure 2 This is a schematic diagram of the internal structure of the vibration damping component in one or more embodiments of the present invention.

[0030] Figure 3 This is a schematic diagram of the structure of a localized resonant phonon crystal structure in one or more embodiments of the present invention.

[0031] Figure 4 This is a schematic diagram of the vibration damping component installed on the handle in one or more embodiments of the present invention.

[0032] Figure 5 This is a schematic diagram of the connecting rod in one or more embodiments of the present invention.

[0033] Figure 6 This is a schematic diagram of the structure of the first oscillator in one or more embodiments of the present invention.

[0034] Figure 7 This is a schematic diagram of the external structure of the first oscillator in one or more embodiments of the present invention.

[0035] Figure 8 This is a top view schematic diagram of a localized resonant phonon crystal structure in one or more embodiments of the present invention.

[0036] Figure 9 For one or more embodiments of the present invention, the magnetic force F1 and the distance between the magnets are... d The relationship curve.

[0037] Figure 10 Rotational speed in one or more embodiments of the present invention n Distance from magnet d The relationship curve.

[0038] Figure 11 This is a schematic diagram of a periodic unit cell structure with four different equilibrium positions in one or more embodiments of the present invention.

[0039] Figure 12 This is a diagram of the band structure and mode shape in one or more embodiments of the present invention.

[0040] Figure 13 This is a schematic diagram illustrating the relationship between the bandgap boundary frequency and the magnet distance in one or more embodiments of the present invention.

[0041] Figure 14 This is a schematic diagram illustrating the relationship between the bandgap boundary range and rotational speed in one or more embodiments of the present invention.

[0042] Among them, 1. Annular shell; 2. Vibration damping component; 3. Handle; 4. Local resonant phonon crystal structure; 5. Second oscillator; 6. First oscillator; 7. Connecting rod; 8. Base beam; 9. Fastener; 601. Outer shell; 602. First magnet; 701. Threaded section; 702. Limiting block. Detailed Implementation

[0043] Example 1

[0044] In a typical embodiment of the present invention, such as Figures 1-14 As shown, a vibration-damping tool holder is presented.

[0045] Passive vibration damping devices have fixed performance parameters and are only applicable to specific working conditions. When machining parameters (such as rotational speed) change, the stiffness or mass of the damping structure must be actively adjusted to adapt to the new working conditions. This is not only complex to operate, but also cannot meet the high-efficiency cutting requirements under varying working conditions. Based on this, this embodiment provides a vibration damping tool holder that, through the design of a structure that can automatically adjust the damping frequency range according to machining parameters, achieves wide-frequency and adaptive vibration suppression.

[0046] like Figures 1-14 As shown, the vibration damping tool holder includes a shank 3 and a vibration damping assembly 2.

[0047] The shank 3 is equipped with a mounting end for the tool, which serves as the connection base between the tool and the machine tool, and directly transmits cutting force and vibration.

[0048] The main body of the vibration damping component 2 is an annular shell 1, which is fitted onto the handle 3. An annular cavity is formed inside the annular shell 1, providing protection and fixed support for the internal structure. A localized resonant phonon crystal structure 4 is installed inside the annular shell 1. The localized resonant phonon crystal structure 4 includes a base beam 8 and oscillator groups. The base beam 8 is annular and coaxially fixed to the outer ring of the annular shell 1, serving as the mounting carrier for the oscillator groups and transmitting vibrations to each oscillator group. Multiple oscillator groups are evenly distributed along the circumferential direction of the annular shell 1. Each oscillator group includes a connecting rod 7, a first oscillator 6, and a second oscillator 5. The connecting rod 7 is radially distributed along the base beam 8, with one end fixed to the base beam 8, providing a sliding and supporting track for the oscillator. The first oscillator 6 is fixed to the connecting rod 7, creating a repulsive force with the second oscillator 5. The second oscillator 5 is located on the side of the first oscillator 6 facing the axis of the base beam 8 and can slide axially along the connecting rod 7, repelling the first oscillator 6 under magnetic force.

[0049] The vibration generated during the cutting process is transmitted to the annular shell 1 through the handle 3, and then to the base beam 8, which ultimately drives the vibration of each oscillator group.

[0050] The position of the oscillator enables adaptive adjustment. Initially, due to the repulsive force between the first oscillator 6 and the second oscillator 5, the second oscillator 5 is located at the end of the connecting rod 7 and fixed by the limiting device. When the handle rotates, the second oscillator 5 is subjected to centrifugal force and slides along the connecting rod 7 toward the first oscillator 6. As the distance decreases, the repulsive force between the two oscillators increases until the centrifugal force and the repulsive force are balanced, and the second oscillator 5 stabilizes in a specific position.

[0051] The change in the position of the oscillator group alters the parameters of the local resonant phonon crystal structure 4, thereby adjusting its bandgap range, which in turn suppresses the frequency range of vibration. As the rotational speed increases, the centrifugal force increases, the distance between the second oscillator 5 and the first oscillator 6 decreases, and the bandgap shifts towards higher frequencies; as the rotational speed decreases, the bandgap shifts towards lower frequencies, achieving adaptive matching between the bandgap and the cutting chatter frequency. Vibrational energy within the bandgap range is absorbed or blocked by the local resonance effect, preventing further transmission to the shank 3 and the tool, thus suppressing cutting vibration.

[0052] This embodiment achieves wideband adaptive vibration reduction without active adjustment. It automatically adjusts the bandgap range based on speed changes to match chatter frequencies under different operating conditions, adapting to vibration suppression requirements in commonly used speed ranges such as 1200-6000 r / min. It eliminates the limitations of passive vibration reduction devices on machining parameters, expanding the range of cutting parameters that can be selected. Relying on mechanical structures and physical forces (centrifugal force, magnetic force) for adjustment, it eliminates the need for sensors, actuators, and other electrical components. The system is simple and reliable, suitable for complex operating conditions such as high-speed, high-power cutting. By adjusting the oscillator's magnetism and structural parameters, it can adapt to different cutting materials and machine tools, further expanding its application scenarios.

[0053] like Figure 2 , Figure 4 As shown, two sets of localized resonant phonon crystal structures 4 are installed inside the annular shell 1, distributed sequentially along the axial direction of the annular shell 1. The two sets of structures form a superposition effect in the axial direction, which can cover a wider range of vibration frequencies. When cutting vibration is transmitted to the annular shell 1, the two sets of structures can act on vibrations in different frequency bands respectively, and suppress vibrations synergistically through their respective localized resonant band gaps, thereby improving the overall comprehensiveness of the vibration reduction effect and avoiding the vibration reduction blind spots that may exist in a single structure in a specific frequency band.

[0054] like Figure 2 and Figure 3 As shown, one end of the connecting rod 7 is connected to the base beam 8 to form a cantilever structure, and the end of the connecting rod 7 away from the base beam 8 forms a limiting block 702. The limiting block 702 and the first oscillator 6 form a sliding space for the second oscillator 5.

[0055] The cantilever structure allows the connecting rod 7 to transmit vibrations more flexibly, ensuring that the oscillator assembly is more sensitive to the vibration response of the base beam 8 and enhancing the triggering efficiency of local resonance. The function of the limiting block 702 is to limit the sliding range of the second oscillator 5 in the initial state or at low speeds, preventing it from moving too far away from the first oscillator 6 due to magnetic repulsion and detaching from the connecting rod 7, thus ensuring the stability of the structure; at the same time, the existence of the sliding space provides the necessary margin of motion for the second oscillator 5 to adjust its position as the rotational speed changes, ensuring that the centrifugal force and magnetic force can be balanced through position changes.

[0056] like Figure 2 and Figure 3 As shown, the first oscillator 6 is sleeved and fixed outside the connecting rod 7, and the second oscillator 5 is slidably sleeved outside the connecting rod 7. The first oscillator 6 and the second oscillator 5 of the same oscillator group are coaxially distributed. The sleeved structure makes the connection between the oscillator and the connecting rod 7 more stable, avoids displacement or loosening during vibration, ensures the stability of the force transmission direction, and allows the second oscillator 5 to slide stably along the axial direction of the connecting rod 7.

[0057] The coaxial distribution ensures that the repulsive force between the first oscillator 6 and the second oscillator 5 always acts along the axial direction, avoiding the generation of additional radial force or torque due to eccentricity, ensuring the stability of the sliding trajectory of the second oscillator 5, and thus ensuring the accuracy of the oscillator position adjustment, so that the change in the bandgap range is more in line with the design expectations.

[0058] like Figure 5 , Figure 6 , Figure 7 As shown, the first oscillator 6 includes a housing 601 and a first magnet 602. The housing 601 is sleeved and fixed to the outside of the connecting rod 7 and forms an annular groove for accommodating the magnet. The first magnet 602 is embedded in the annular groove formed by the housing 601.

[0059] The outer casing 601 can be made of steel to fix the first magnet 602, avoid wear or positional displacement caused by direct contact between the first magnet 602 and the connecting rod 7, and enhance the structural strength of the first oscillator 6 so that it remains stable under high-frequency vibration.

[0060] The design of the annular groove ensures that the magnet can be accurately positioned and that the area facing the second oscillator 5 is stable, thereby keeping the repulsive force characteristics between the two consistent and providing a stable physical parameter basis for the balance calculation of centrifugal force and magnetic force.

[0061] The second oscillator 5 is a second magnet, which is a ring magnet. One end of the first magnet 602 of the first oscillator 6 and one end of the second magnet of the second oscillator 5 are directly opposite each other. The ring magnet allows it to fit more snugly around the connecting rod 7, improving guidance during sliding and reducing frictional interference with the connecting rod 7. The structure with both ends directly opposite ensures the maximum utilization of the repulsive force between them, avoiding force loss due to misalignment. This makes the change of magnetic force with distance more significant, thereby making the position adjustment of the second oscillator 5 more sensitive. Even a small change in rotation speed can drive the second oscillator 5 to slide through the change of centrifugal force, achieving a rapid response within the bandgap range.

[0062] The annular shell 1 and the handle 3 are coaxially distributed. The inner ring of the annular shell 1 is attached to and connected to the handle 3. The coaxial distribution ensures that the vibration transmission path from the handle 3 to the annular shell 1 is symmetrical, avoiding uneven distribution of vibration energy caused by eccentricity, so that each oscillator group can be evenly stressed and play a synergistic vibration reduction role.

[0063] The inner ring of the annular shell 1 fits into the connecting handle 3, which enhances the connection rigidity between the two, reduces energy loss during vibration transmission, and ensures that more vibration energy can be absorbed by the internal local resonant phonon crystal structure 4, thereby improving vibration reduction efficiency.

[0064] The section where the connecting rod 7 slides with the second oscillator 5 is a smooth shaft section. The connecting rod 7 has a threaded section 701, which can be connected to the base beam 8. At the same time, the base beam 8 can be fixed to the annular housing 1 by fasteners 9 such as screws. The smooth shaft section reduces the frictional resistance when the second oscillator 5 slides, allowing the second oscillator 5 to adjust its position more smoothly according to the balance between centrifugal force and magnetic force, avoiding position adjustment lag caused by jamming, ensuring that the bandgap range can match the flutter frequency changes caused by speed changes in real time, and improving the response speed and accuracy of adaptive vibration reduction.

[0065] Taking the HSK 63 tool holder as an example, such as Figure 1 As shown, the middle section has a diameter of 46mm, and the inner diameter of the annular shell 1 is 46mm. It is clamped in the middle section of the handle 3, as shown. Figure 8As shown, the base beam 8 has an inner diameter of 94 mm and an outer diameter of 98 mm. The localized resonant phonon crystal structure 4 is fixed to the outer ring of the inner wall of the annular shell 1 using circumferentially distributed fastening screws as fasteners 9. The base beam 8 has a thickness of 2 mm. The connecting rod 7 is a non-ferromagnetic aluminum alloy stud, including a body and a limiting block 702. One end of the connecting rod 7 is connected and fixed to the threaded hole on the base beam 8 through a threaded section 701. The first oscillator 6 is a bowl-shaped structure formed by the outer shell 601 surrounding the annular first magnet 602. It has a thickness of 5 mm and a diameter of 12 mm. The first magnet 602 is a neodymium iron boron strong magnet. The center of the outer shell 601 of the first oscillator 6 has an M2 internal threaded hole, which is screwed onto the connecting rod 7. The second oscillator 5 can also be a neodymium iron boron strong magnet with a diameter of 12 mm, an inner diameter of 2 mm, and a thickness of 3 mm. It is ensured to slide freely on the optical axis section of the connecting rod 7 through a clearance fit.

[0066] The distance and magnetic force relationship between the first oscillator 6 and the second oscillator 5 were calculated using the finite element method. Neodymium iron boron (N54) magnets were used, with a density of 7600 kg / m³. Elastic modulus 160 GPa, Poisson's ratio 0.24, remanent flux density modulus 1.47 T, restoring permeability 1.05, magnetic force F1 and distance between magnets d Relationship curves as follows Figure 9 As shown, the magnetic force can be obtained by exponential fitting. Centrifugal force can be generated by... Calculated, where m The mass of the second oscillator 5 is approximately 0.0025 kg. n It refers to the rotational speed of the tool holder, and the distance between the magnets corresponding to the first oscillator 6 and the second oscillator 5. d and the radius of the oscillator's center of mass r The relationship is:

[0067] r =37.5- d (mm);

[0068] When the tool holder speed is at the commonly used... n Within the range of 1200-6000 r / min, reaching F 1= F At 2 o'clock, the rotational speed n The distance between the magnets when the oscillator is in equilibrium d Relationship curves as follows Figure 10 As shown, the mathematical relational expression is:

[0069] ;

[0070] The base beam 8 and connecting rod 7 are made of aluminum alloy 6061, with an elastic modulus of 76.5 GPa, Poisson's ratio of 0.32, and a density of 2733 kg / m³. The steel outer shell of the first oscillator 6 has an elastic modulus of 200 GPa, a Poisson's ratio of 0.3, and a density of 7850 kg / m³. .

[0071] The base beam 8, the circumferentially distributed connecting rods 7, the first oscillator 6, and the second oscillator 5 are considered as a periodically distributed Euler beam-additional oscillator structure. A localized resonant phononic crystal unit cell structure containing one period is extracted, which is an equivalent straight beam segment with a length of 1 / 8 of the base beam 8 (because there are 8 circumferentially distributed oscillators, hence one-eighth of the length), connected by a connecting rod 7, a first oscillator 6, and a second oscillator 5. Comsol finite element analysis is performed on the periodic unit cell structure of the localized resonant phononic crystal. Equivalent periodic unit cell structures with various equilibrium positions of the second oscillator 5 are selected, such as... Figure 11 As shown. Bloch boundary conditions are applied to the unit cell to simulate wave propagation in an ideal periodic structure. The boundary conditions are given by the following equation:

[0072] ;

[0073] Where u(r) is the nodal displacement vector, a is the period constant (20 mm), and r is the nodal vector. k is the wave vector. By taking values ​​of the wave vector along the irreducible Brillouin zone boundary (ГХ) and solving for eigenvalues, the band structure of the phononic crystal can be obtained. The frequency range in the band structure through which no elastic wave mode passes is the band gap of the phononic crystal. Figure 12 As shown, with d Taking a value of 10mm as an example, the modal shapes at points A and B are as follows: Figure 12 As shown.

[0074] from Figure 12 It can be seen that the four typical structures exhibit vibration-damping band gaps at 516-712Hz, 633-911Hz, 777-1150Hz, and 946-1283Hz, respectively. The structural oscillator absorbs vibrational energy through local resonance, preventing vibration propagation and thus creating the band gap. The presence of the band gap can suppress vibrations within a specific frequency range of the tool holder. Simulations show the band gap boundary range as follows: Figure 13 As shown.

[0075] Based on the mechanism of centrifugal force and magnetic force balancing and thus changing the position of the oscillator, chatter induced by low rotational speed often corresponds to the low-frequency mode of the spindle system, while chatter induced by high rotational speed corresponds to the high-frequency mode. As the rotational speed increases, the chatter frequency also increases. The bandgap range corresponding to different rotational speeds can be obtained through curve fitting, such as... Figure 14 As shown, this is consistent with the flutter frequency-rotation speed matching relationship, and the bandgap frequency obtained at different rotational speeds ( f Range of variation With a range of up to 150% and a wide variation range, by rationally selecting the rotational speed or designing structural parameters, the bandgap frequency and chatter frequency can correspond to different rotational speeds, achieving adaptive adjustment of the tool holder bandgap range during the cutting process. This eliminates the need for continuous manual adjustment of the vibration suppression structure, effectively suppressing chatter under varying parameter conditions. Furthermore, the size and material of the strong magnet or initial structural parameters can be adjusted as needed to adapt to different cutting materials and machine tools.

[0076] Example 2

[0077] In another typical embodiment of the present invention, such as Figures 1-14 As shown, a method for operating a vibration-damping tool holder is provided, utilizing the vibration-damping tool holder as described in Example 1. The specific steps include:

[0078] The cutting tool is installed to the mounting end of the shank 3 and then as a whole to the tool holder to form a cutting assembly;

[0079] The cutting tool generates vibration during the cutting process. The vibration is first transmitted to the annular shell 1, and then transmitted to the annular base beam 8 which is fixed to the outer ring of the shell.

[0080] In the initial state, the second oscillator 5 is located at the end of the connecting rod 7 due to the magnetic repulsion of the first oscillator 6;

[0081] When the vibration damping tool holder starts to rotate, the second oscillator 5 is subjected to centrifugal force, which causes the second oscillator 5 to slide along the connecting rod 7 toward the first oscillator 6. As the second oscillator 5 gradually approaches the first oscillator 6, the magnetic repulsion between the two gradually increases until the centrifugal force and the magnetic repulsion reach a balance, and the second oscillator 5 stabilizes at this specific position. When the rotational speed of the vibration damping tool holder changes, the position of the second oscillator 5 on the connecting rod 7 changes, altering the local resonance effect of the vibration damping component 2 to adapt it to the vibration at that rotational speed.

[0082] The vibration damping component 2 generates a local resonance effect to absorb the vibration energy transmitted by the handle 3 and suppress vibration.

[0083] Without the need for an external control system, the resonant frequency can be dynamically adjusted simply through the coupling of mechanical structure and physical force, and it can adapt to changes in vibration characteristics under different cutting parameters.

[0084] The above description is merely a preferred embodiment of the present invention and is not intended to limit the invention. Various modifications and variations can be made to the present invention by those skilled in the art. Any modifications, equivalent substitutions, improvements, etc., made within the spirit and principles of the present invention should be included within the scope of protection of the present invention.

Claims

1. A vibration-damping tool holder, characterized in that, include: The shank has a mounting end for mounting the cutting tool. The vibration damping assembly includes an annular shell and a local resonant phonon crystal structure located within the annular shell. The shell is sleeved and connected to a handle. The local resonant phonon crystal structure includes an annular base beam and multiple oscillator groups. The base beam is coaxially fixed to the outer ring of the annular shell. The multiple oscillator groups are evenly distributed along the circumferential direction of the annular shell. Each oscillator group includes a connecting rod, a first oscillator, and a second oscillator. The connecting rod is radially distributed along the base beam and one end is connected to the base beam. The first oscillator is fixed to the connecting rod. The second oscillator is located on the side of the first oscillator facing the axis of the base beam and slides axially with the connecting rod. The first oscillator and the second oscillator repel each other. The first oscillator includes a housing and a magnet. The housing is fitted and fixed to the outside of the connecting rod and has an annular groove for accommodating the magnet, which is embedded in the annular groove. The second oscillator is a ring magnet, with one end of the first oscillator and one end of the second oscillator facing each other; The cutting tool generates vibration during the cutting process. The vibration is first transmitted to the annular shell, and then transmitted to the annular base beam fixed to the outer ring of the shell. Initially, the second oscillator is located at the end of the connecting rod due to the magnetic repulsion of the first oscillator; When the damping tool holder starts to rotate, the second oscillator will be subjected to centrifugal force, which will cause the second oscillator to slide along the connecting rod towards the first oscillator. As the second oscillator gradually approaches the first oscillator, the magnetic repulsion between the two will gradually increase until the centrifugal force and the magnetic repulsion reach a balance, and the second oscillator will stabilize at this specific position. The vibration damping components generate local resonance to absorb the vibration energy transmitted from the handle and suppress vibration.

2. The vibration-damping tool holder as described in claim 1, characterized in that, Two sets of localized resonant phonon crystal structures are installed inside the annular shell, and are distributed sequentially upward along the axial direction of the annular shell.

3. The vibration-damping tool holder as described in claim 1, characterized in that, One end of the connecting rod is connected to the base beam to form a cantilever structure, and the end of the connecting rod away from the base beam forms a limiting block. The limiting block and the first oscillator form a sliding space for the second oscillator.

4. The vibration-damping tool holder as described in claim 3, characterized in that, The first oscillator is fixed outside the connecting rod, and the second oscillator is slidably fitted outside the connecting rod. The first and second oscillators in the same oscillator group are coaxially distributed.

5. The vibration-damping tool holder as described in claim 1, characterized in that, The annular shell and the handle are coaxially distributed, and the inner ring of the annular shell is attached to and connected to the handle.

6. The vibration-damping tool holder as described in claim 1, characterized in that, The segment in which the connecting rod slides into the second oscillator is the optical axis segment.

7. A method for operating a vibration-damping tool holder, utilizing the vibration-damping tool holder as described in any one of claims 1-6, characterized in that, include: The cutting tool is installed onto the mounting end of the shank and then mounted as a whole onto the tool holder; When the rotational speed of the vibration damping tool holder changes, the position of the second oscillator on the connecting rod changes, altering the local resonance effect of the vibration damping component to match the vibration at that rotational speed.

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