KR101764233B1 - The manufacturing method of a dynamic vibration absorber - Google Patents

The manufacturing method of a dynamic vibration absorber Download PDF

Info

Publication number
KR101764233B1
KR101764233B1 KR1020150143692A KR20150143692A KR101764233B1 KR 101764233 B1 KR101764233 B1 KR 101764233B1 KR 1020150143692 A KR1020150143692 A KR 1020150143692A KR 20150143692 A KR20150143692 A KR 20150143692A KR 101764233 B1 KR101764233 B1 KR 101764233B1
Authority
KR
South Korea
Prior art keywords
frequency
mechanical device
vibration
dynamic damper
mass
Prior art date
Application number
KR1020150143692A
Other languages
Korean (ko)
Other versions
KR20170043961A (en
Inventor
권혁
Original Assignee
삼성중공업 주식회사
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by 삼성중공업 주식회사 filed Critical 삼성중공업 주식회사
Priority to KR1020150143692A priority Critical patent/KR101764233B1/en
Publication of KR20170043961A publication Critical patent/KR20170043961A/en
Application granted granted Critical
Publication of KR101764233B1 publication Critical patent/KR101764233B1/en

Links

Images

Classifications

    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B63—SHIPS OR OTHER WATERBORNE VESSELS; RELATED EQUIPMENT
    • B63B—SHIPS OR OTHER WATERBORNE VESSELS; EQUIPMENT FOR SHIPPING 
    • B63B17/00—Vessels parts, details, or accessories, not otherwise provided for
    • B63B17/0081—Vibration isolation or damping elements or arrangements, e.g. elastic support of deck-houses
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16F—SPRINGS; SHOCK-ABSORBERS; MEANS FOR DAMPING VIBRATION
    • F16F15/00—Suppression of vibrations in systems; Means or arrangements for avoiding or reducing out-of-balance forces, e.g. due to motion
    • F16F15/02—Suppression of vibrations of non-rotating, e.g. reciprocating systems; Suppression of vibrations of rotating systems by use of members not moving with the rotating systems
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16F—SPRINGS; SHOCK-ABSORBERS; MEANS FOR DAMPING VIBRATION
    • F16F15/00—Suppression of vibrations in systems; Means or arrangements for avoiding or reducing out-of-balance forces, e.g. due to motion
    • F16F15/02—Suppression of vibrations of non-rotating, e.g. reciprocating systems; Suppression of vibrations of rotating systems by use of members not moving with the rotating systems
    • F16F15/04—Suppression of vibrations of non-rotating, e.g. reciprocating systems; Suppression of vibrations of rotating systems by use of members not moving with the rotating systems using elastic means

Landscapes

  • Engineering & Computer Science (AREA)
  • General Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • Physics & Mathematics (AREA)
  • Acoustics & Sound (AREA)
  • Aviation & Aerospace Engineering (AREA)
  • Chemical & Material Sciences (AREA)
  • Combustion & Propulsion (AREA)
  • Ocean & Marine Engineering (AREA)
  • Vibration Prevention Devices (AREA)

Abstract

A method of manufacturing a dynamic damper installed in a mechanical device according to an embodiment of the present invention includes the steps of calculating a natural frequency (? 0 ) by measuring a mass (m 1 ) and an elastic modulus (k 1 ) of the mechanical device, mechanism vibration speed (v o) than the step of selecting a first weight (α) to produce a first vibration speed (v 1) the oscillation speed is small, the larger the frequency the more the natural frequencies, two frequencies of the (ω 2 ) From the equation of the first oscillation speed (v 1 ) and the second oscillation number (ω 2 ) represented by the mass and the elastic modulus, (m 2 ) and the elastic modulus (k 2 ) of the dynamic damper connecting portion.

Figure 112015099455866-pat00041

Description

BACKGROUND OF THE INVENTION 1. Field of the Invention [0001] The present invention relates to a method of manufacturing a dynamic damper,

The present invention relates to a method of manufacturing a dynamic damper installed in a structure.

The vibration and noise generated during the voyage of the ship cause the inconvenience of the crew and also affects the various structures on the ship, causing difficulties in safe navigation. In recent years, as vibration and noise regulations have been strengthened, interest in anti-vibration has been increasing more and more. Therefore, the necessity of development of anti-vibration device is becoming more important.

Particularly, in ships, structures such as bridge wing, radar mast, and foremast are considerable in weight and have a length longer than the width, so that the excitation source The possibility of resonance is very high. These vibrations affect the navigation equipment, which is a great obstacle to navigation.

In order to solve such a vibration problem, a dynamic damper may be installed. In the damper, a small mass-spring-damper system is attached to a large vibrating body vibrating at a certain frequency such as a mechanical structure, Method.

The principle of the dynamic damper is to adjust the dynamic damper composed of the mass-spring-damper system to the natural frequency (natural frequency, eigenfrequency) of the vibrating body vibrating at a certain frequency, Give.

SUMMARY OF THE INVENTION The present invention has been made in view of the above-mentioned technical problems, and it is an object of the present invention to provide a method of manufacturing a dynamic damper which is capable of manufacturing a dynamic damper installed in a structure or a mechanical device more efficiently and compactly.

A method of manufacturing a dynamic damper according to an embodiment of the present invention is provided in a mechanical device

Figure 112015099455866-pat00001
(Ω 0 ) by measuring a mass (m 1 ) and an elastic modulus (k 1 ) of the mechanical device, wherein the elastic modulus (k 1 ) of the mechanical device satisfies the vibration equation , Selecting a first weight (?) For calculating a first vibration speed (v 1 ) having a vibration speed lower than a vibration speed (v o ) of the mechanical device, calculating a second frequency (v 1 ) ( 2 ) to calculate the mass (?) of the dynamic vibration reducer ( 2 ) from the first oscillation speed (v 1 ) and the second oscillation number (? 2 ) expressed by mass and elastic modulus (m 2 ) and the elastic modulus (k 2 ) of the dynamic damper connecting portion.

The method of manufacturing a dynamic damper according to an embodiment of the present invention can prevent the excessive dynamic damper from being manufactured by mass design and elasticity, thereby manufacturing a more efficient and compact dynamic damper.

1 is a conceptual view showing a mechanical device with a dynamic damper.
2 is a vibration model for designing a dynamic damper applied to Fig.
FIG. 3 is a graph showing the amplitude and the frequency according to the vibration damping device manufacturing method using the vibration model of FIG. 2. FIG.
4 is a flowchart illustrating a method for designing a dynamic damper in the method for manufacturing dynamic damper according to an embodiment of the present invention.
FIG. 5 is a graph showing the frequency and amplitude of the dynamic damper manufactured according to FIG.

Hereinafter, exemplary embodiments of the present invention will be described in detail with reference to the accompanying drawings, which will be readily apparent to those skilled in the art to which the present invention pertains. The present invention may be embodied in many different forms and is not limited to the embodiments described herein.

In order to clearly illustrate the present invention, parts not related to the description are omitted, and the same or similar components are denoted by the same reference numerals throughout the specification.

In addition, since the sizes and thicknesses of the respective components shown in the drawings are arbitrarily shown for convenience of explanation, the present invention is not necessarily limited to those shown in the drawings.

Also, throughout the specification, when an element is referred to as "including" an element, it is understood that the element may include other elements as well, without departing from the other elements unless specifically stated otherwise.

1 is a conceptual view showing a mechanical device with a dynamic damper.

1, the dynamic damper 20 includes a vibrating portion 22 having a mass and a connecting portion 24 connecting the vibrating portion 22 with a structure or a mechanical device . The connection portion 24 is made of a material having elasticity and may be a spring. The shape of the vibrating portion 22 is not limited and the mass is designed according to the mechanical device 10. [

Vibration is generated in the vertical direction by the vibrating force (excitation force, excitation force) 15 of the mechanical device 10. [ At this time, when the frequency of the exciting force 15 becomes equal to the frequency of the mechanical device 10, resonance occurs and excessive vibration occurs in the mechanical device 10, It may not be able to perform its functions or may impair stability.

The excitation force of the dynamic damper 20 attached to the mechanical device 10 is a vibration force of the mechanical device 10. [ Therefore, the vibration of the dynamic damper 20 is delayed and oscillated more than the vibration of the mechanical device 10. [ Therefore, the vibratory force generated by the dynamic damper 20 is generated in a 180 占 phase opposite to the excitation force 15 applied to the mechanical device 10, and all of the opposite forces act on the mechanical device 10 The vibration of the mechanical device 10 is greatly reduced. Therefore, it is very important to design the mass (m 2 ) and the elastic modulus (k 2 ) of the dynamic damper 20 in accordance with the mechanical device 10.

2 is a vibration model for designing a dynamic damper applied to Fig.

Referring to Fig. 2, a vibration model showing a mechanical device 10 to which the dynamic damper 20 is attached is disclosed.

The equation of motion of the mechanical device 10 and the dynamic damper 20 is expressed by the following equation (1).

[Equation 1]

Figure 112015099455866-pat00002

Figure 112015099455866-pat00003

Here, m 1 , k 1 and x 1 mean the mass, elastic modulus and amplitude of the mechanical device 10, and m 2 , k 2 and x 2 mean the mass, elastic modulus and amplitude of the dynamic damper 20 do.

Figure 112015099455866-pat00004
Means an excitation force. At this time, the amplitude can be obtained by the following equation (2).

&Quot; (2) "

Figure 112015099455866-pat00005

Figure 112015099455866-pat00006

Therefore, the condition that the amplitude x 1 of the mechanical device 10 near the resonance point of the vibration system before the damper is attached when the damper is attached theoretically in Equation (2) is 0 is expressed by Equation (3) .

&Quot; (3) "

Figure 112015099455866-pat00007

As can be seen from Equation (3), theoretically, if the amplitude (x 1 ) of the mechanical device 10 becomes zero near the resonance point, the ratio of the mass m 1 of the mechanical device to the elastic modulus k 1 , The ratio of mass (m 2 ) to elastic modulus (k 2 ) should be designed to be the same.

That is, since the frequency ω is defined as a ratio of the mass m to the elastic modulus k, it is customary to design the natural frequency ω 0 of the mechanical device 10 to match the natural frequency of the dynamic damper.

FIG. 3 is a graph showing the amplitude and the frequency according to the vibration damping device manufacturing method using the vibration model of FIG. 2. FIG.

2 and 3, there is shown a graph when the natural frequency ( 0 ) of the mechanical device 10 coincides with the frequency of the dynamic damper 20 as described above.

When the frequency of the dynamic damper 20 is designed to be equal to the natural frequency ω 0 of the mechanical device 10, when the frequency theoretically oscillates at the natural frequency ω 0 of the mechanical device 10, 0 < / RTI >

That is, when the ratio of the frequency of the dynamic damper 20 to the natural frequency (? 0 ) before attachment of the damper of the mechanical device 10 is 1, the amplitude x becomes theoretically zero.

When the mechanical device 10 is operated, vibration due to the excitation force 15 is generated in the mechanical device 10 as mentioned above. The vibration frequency omega of the vibration caused by the exciting force 15 can be formed to be smaller than the natural frequency omega o of the mechanical device 10 in the case where the power device having an operation speed varying such as an automobile or a ship is changed. At this time, if the frequency of the excitation force is formed to be equal to the natural frequency? 1 of the mechanical device 10 including the dynamic damper 20, the amplitude x of the excitation force 15 is greatly increased . In this case, too, there is a problem that the mass (m 2 ) of the dynamic damper 20 may excessively increase to obtain a sufficiently small amplitude.

4 is a flowchart illustrating a method of manufacturing a dynamic damper according to an embodiment of the present invention.

4, a method of manufacturing the dynamic damper 20 including the vibrating portion 22 and the connecting portion 24 provided in the mechanical device 10 is disclosed. The dynamic damper manufactured in accordance with the present invention is applicable to all vibration systems capable of being represented by a one degree of freedom vibration system.

First, the mass (m 1 ) and the elastic modulus (k 1 ) of the mechanical device are measured to calculate a natural frequency (ω 0 ) (S 10).

The natural frequency of the mechanical device 10 is

Figure 112015099455866-pat00008
. Here, the mass m 1 of the machine 10 and the elastic modulus k 1 of the machine can be measured or analyzed. Every object with mass and elasticity has a natural frequency, and therefore all the machines 10 have their natural frequency ( 0 ). Therefore, the dynamic damper 20 according to an embodiment of the present invention can be used for all the mechanical devices 10, and each mechanical device 10 should be designed for each.

The dynamic damper 20 having a large mass m 2 has a large value because the ratio between the mass m 2 of the dynamic damper 20 and the elastic modulus k 2 is meaningful There is a great risk of excessive design of the dynamic damper 20. If the mass (m 2 ) is too small, the dynamic damper 20 can not prevent the mechanical device 10 from vibrating unlike the theory. Therefore, it is important to design the dynamic damper 20 having a suitable mass (m 2 ) while preventing vibration of the mechanical device 10 more efficiently.

Next, a step of selecting a first weight? For calculating a first vibration velocity v 1 having a vibration velocity lower than a vibration velocity v o of the mechanical device 10 is selected (S20).

When the dynamic damper 20 is attached to the mechanical device 10, two natural frequencies are generated in the mechanical device 10 to which the dynamic damper 20 is attached. For convenience original mechanism 10 natural frequency (ω 0) than the first frequency (ω 1) for small frequency frequency La, and the more the greater the frequency the frequency original mechanism 10 natural frequency (ω 0) 2 is called the frequency (ω 2 ).

Is smaller than the natural frequency (? 0 ) of the mechanical device (10) in the case of the first frequency (? 1 ). However, if the amplitude x 1 of the first frequency? 1 is also smaller than the amplitude of the natural frequency? 0 of the mechanical device 10, the frequency? Even if it becomes equal to the first frequency (? 1 ), no problem arises because the amplitude due to the resonance is not so large as to damage the mechanical device (10). Therefore, a very important factor for adjusting the amplitude (x 1) of a first frequency (ω 1), the user is the natural frequency (ω a first frequency amplitude (x 1) a mechanical system (10) of the (ω 1) 0 ) and the amplitude. Therefore, the first weight? To be described later is selected. Normally, the first weight (α) should be designed by referring to the allowable vibration standard of the equipment (international standard or classification guide).

Next, a step of selecting a second weight? For calculating a second frequency? 2 having a frequency greater than the natural frequency? 0 is selected (S30).

The second frequency? 2 means a larger frequency as compared to the natural frequency? 0 of the mechanical device 10. [ That is, than the natural frequency (ω 0) of the exciting force (15) the frequency (ω) of the can is formed to be smaller than the natural frequency (ω 0) of the mechanism 10, the mechanism 10 acting on the mechanism (10) And does not reach the second large frequency ( 2 ).

Therefore, when selecting the second frequency (ω 2), amplitude (x 2) than the second frequency (ω 2) is not a very important factor to appropriately larger adjustment than the natural frequency (ω 0) of the mechanism (10) .

Next, the mass of the first vibration speed (v 1) and the second frequency (ω 2), the in-reducer 20, the vibration 22 from the equation, expressed in the mass and the elastic coefficient (m 2) and the and a step for calculating the elastic modulus (k 2) of that damper (20) connecting portion (24) (S40).

FIG. 5 is a graph showing the frequency and amplitude of the dynamic damper manufactured according to FIG.

4 and 5, a method of manufacturing the dynamic damper 20 according to an embodiment of the present invention includes the steps of: measuring a natural frequency ω 0 of the conventional mechanical device 10 and a frequency of the dynamic damper 20 Unlike the manufacturing method in which the dynamic damper 20 is attached, the amplitude x 1 at the first frequency ω 1 generated in the mechanical device 10 to which the dynamic damper 20 is attached is smaller than the amplitude x 1 at the natural frequency ω 0 before the damper Is designed to be 30% of the amplitude (x 0 ), and the second frequency (ω 2 ) is designed to avoid 110% of the natural frequency (ω 0 ) before the damper attachment.

As explained above, the frequency (ω) by the exciting force (15) is more than 0 ω 0 ≪ / RTI > In this case, since the frequency ω by the excitation force 15 is generated to be smaller than the natural frequency ω 0 of the mechanical device 10, the natural frequency ω 0 and the oscillation speed v 0 of the mechanical device 10, The first weight? And the second weight? Are selected on the basis of the first weight? And the second weight?.

When the mechanical device 10 has no attenuation, the amplitude becomes infinite during resonance when the frequency? By the excitation force 15 is equal to the first frequency? 1 and the second frequency? 2 . However, this is the theoretical case, and in practice, the structure attenuation according to the structure of the machine 10 itself

Figure 112015099455866-pat00009
), The amplitude (x) is generated in a certain size.

At this time, the design criteria can be seen in two ways. First, is it possible to move the second frequency (? 2 ) to not less than the natural frequency (? 0 ) of the mechanical device 10 where resonance may occur. Second, is it possible to reduce the amplitude (x 1 ) of the first frequency? 1 , which is formed to be smaller than the natural frequency? 0 of the mechanical device 10, to the extent that the mechanical device 10 can absorb it.

Through these two design standards, the dynamic damper 20 is excessively designed, so that the weight can be prevented from becoming larger than necessary and increasing in size. According to the conventional design method, the ratio of the mass (m 2 ) of the dynamic damper 20 to the elastic modulus (k 2 ) is fixed, but the designing method according to the embodiment of the present invention can freely design have.

Structural damping of the machine 10 (

Figure 112015099455866-pat00010
) Is expressed by using a matrix as shown in Equation (4).

&Quot; (4) "

Figure 112015099455866-pat00011

When the frequency of the mechanical device 10 to which the dynamic damper 20 is attached is obtained from the equation (4), the first frequency? 1 and the second frequency? 2 are generated as shown in equation (5).

&Quot; (5) "

Figure 112015099455866-pat00012

Figure 112015099455866-pat00013

here,

Figure 112015099455866-pat00014
to be.

The amplitude x 1 of the mechanical device 10 is obtained when the frequency ω by the excitation force 15 is equal to the first frequency ω 1 and ω = ω 1 .

&Quot; (6) "

Figure 112015099455866-pat00015

Here, substituting S for the preceding matrix is as follows.

Figure 112015099455866-pat00016

here,

Figure 112015099455866-pat00017

Figure 112015099455866-pat00018
to be.

Therefore, the inverse matrix of S,

Figure 112015099455866-pat00019

here,

Figure 112015099455866-pat00020

therefore,

Figure 112015099455866-pat00021

Figure 112015099455866-pat00022

to be.

Then, the tuning parameters? And? Are defined as in Equation (7).

Figure 112015099455866-pat00023

? represents the ratio of the natural frequency (? 0 ) of the mechanical device 10 to the second frequency (? 2 ). the second frequency omega 2 is larger than the natural frequency omega 0 of the mechanical device 10 so that the frequency omega due to the excitation force 15 can be applied to the dynamic damper 20 Can not reach the second frequency (? 2 ) of the system which is a mechanical device (10), so that the risk of resonance can be avoided. Therefore, the range of? Can be selected in a range exceeding 1.

Since the second frequency ω 2 includes the mass and the elastic modulus, a second relational expression can be obtained which is an equation of the mass m 2 of the dynamic damper 20 and the elastic modulus k 2 .

If the product of the first frequency (ω 1) amplitude (x 1) to be represented by a vibration speed (v 1). When measuring vibration, the vibration velocity is used because the vibration velocity is more convenient to measure than the amplitude. The vibration speed means the vibration level.

A first frequency (ω 1) and the amplitude (x 1) it is all represented by a mass and an elastic coefficient, the first equation of the mass (m 2) and the elastic modulus (k 2) of that damper (20) through their product A relational expression can be obtained.

When? is smaller than 1 , it means that the vibration speed v 1 of the first frequency is smaller than the vibration speed v 0 of the natural frequency (? 0 ) of the mechanical device (10). That is, even if the resonance occurs due to the frequency (?) Of the excitation force 15 being equal to the first frequency (? 1 ), the mechanical device (10) can be operated without a large damage.

As a result, the user can select α and β, and calculate the mass (m 2 ) and the elastic modulus (k 2 ) of the dynamic damper 20 by combining the first and second relational expressions.

For example, the dynamic damper 20 is designed in the mechanical device 10 having the following conditions. The initial condition of the machine 10 is

Figure 112015099455866-pat00024
= 0.04, m 1 = 4537.9 kg, k 1 = 6493510 N / m, ω 0 = 37.83 rad / s (= 6.02 Hz), and v 0 = 100 mm / s. At this time, when the second frequency ω 2 is shifted by 10% on the basis of the natural frequency ω 0 of the mechanical device 10 and the amplitude x 1 of the first frequency ω 1 is reduced by 30% . That is, in the case where? = 1.1 and? = 0.3, the dynamic damper 20 is designed as follows.

Figure 112015099455866-pat00025

3, the first frequency ω 1 and the second frequency ω 2 generated by installing the dynamic damper 20 are located on both sides of the natural frequency ω 0 . At this time, the second frequency (ω 2) has a natural frequency (ω 0) of the mechanism 10, the natural frequency is positioned to 110% frequencies greater than (ω 0) of the mechanism 10, as discussed above And the magnitude of the amplitude x 1 in the case of the first frequency ω 1 which can be equal to the frequency ω of the excitation force 15 to be applied to the mechanical device 10, (ω 0 ) of 30% or less of the maximum amplitude (x), even if resonance occurs, there is no great damage to the mechanical device 10.

While the invention has been shown and described with reference to certain preferred embodiments thereof, it will be understood by those skilled in the art that various changes and modifications may be made therein without departing from the spirit and scope of the following claims. Those who are engaged in the technology field will understand easily.

10: Mechanism 15: Excitation force
20: Dynamic damper 22:
24: Connection
m 1 : Machine mass m 2 : Copper damper mass
k 1 : mechanical device elasticity k 2 : dynamic damper elasticity
ω: the frequency of the excitation force ω 0 : the natural frequency of the mechanical device
ω 1 : first frequency ω 2 : second frequency
?: the ratio of the natural frequency to the first frequency (first weight)
β: ratio of the natural frequency to the second frequency (second weight)
x: Amplitude of mechanical device with dynamic damper
x 0 : amplitude of the mechanical device x 1 : amplitude of the first frequency
x 2 : second frequency amplitude

Claims (2)

1. A method of manufacturing a dynamic damper provided in a mechanical device and satisfying the vibration equation of [Equation 1], including a vibrating portion and a connecting portion,
Measuring a mass (m 1 ) and an elastic modulus (k 1 ) of the mechanical device to calculate a natural frequency (? 0 );
Selecting a first weight (alpha) for calculating a first oscillation speed (v 1 ) having a vibration speed lower than a vibration speed (v o ) of the mechanical device;
Selecting a second weight (?) For calculating a second frequency (? 2 ) having a frequency greater than the natural frequency; And
(M 2 ) of the dynamic vibration reducer oscillation portion and the elastic modulus (k 2 ) of the dynamic damper coupling portion from the first vibration velocity (v 1 ) and the second vibration frequency (ω 2 ) represented by mass and elastic modulus , ≪ / RTI >
The first weight < RTI ID = 0.0 >
Figure 112017001593831-pat00043
And is selected within the range of 0 to less than 1,
The second weight < RTI ID = 0.0 >
Figure 112017001593831-pat00044
, And is selected within a range of more than 1 and less than 3.
[Formula 1]
Figure 112017001593831-pat00026

Here, the vibration speed of the mechanical device
Figure 112017001593831-pat00027
, The first oscillation speed
Figure 112017001593831-pat00028
,
Figure 112017001593831-pat00045
Is the mechanical damping coefficient The method according to claim 1,
A first relation expression for the mass (m 2 ) of the vibration portion of the dynamic damper and the elastic modulus of the joint (k 2 ) is obtained through the following [Formula 2]
[Formula 2]
Figure 112017001593831-pat00029

Figure 112017001593831-pat00030
,
Figure 112017001593831-pat00031

Figure 112017001593831-pat00032
,
Figure 112017001593831-pat00033
,
Figure 112017001593831-pat00034

A second relation expression for the mass (m 2 ) of the vibration portion of the dynamic damper and the elastic modulus of the connecting portion (k 2 ) is obtained through the following [Formula 3]
[Formula 3]
Figure 112017001593831-pat00035

Figure 112017001593831-pat00036
,
(M 2 ) and the elastic modulus of the connecting portion (k 2 ) of the dynamic damper are obtained by simultaneously taking the first relationship and the second relation,
Here, the natural frequency (ω 0 ), the natural frequency amplitude (x 0 ), the machine mass (m 1 ), the machine elastic modulus (k 1 ), the machine structure damping coefficient
Figure 112017001593831-pat00037
) Is substituted for a preset value, respectively.
KR1020150143692A 2015-10-14 2015-10-14 The manufacturing method of a dynamic vibration absorber KR101764233B1 (en)

Priority Applications (1)

Application Number Priority Date Filing Date Title
KR1020150143692A KR101764233B1 (en) 2015-10-14 2015-10-14 The manufacturing method of a dynamic vibration absorber

Applications Claiming Priority (1)

Application Number Priority Date Filing Date Title
KR1020150143692A KR101764233B1 (en) 2015-10-14 2015-10-14 The manufacturing method of a dynamic vibration absorber

Publications (2)

Publication Number Publication Date
KR20170043961A KR20170043961A (en) 2017-04-24
KR101764233B1 true KR101764233B1 (en) 2017-08-03

Family

ID=58704380

Family Applications (1)

Application Number Title Priority Date Filing Date
KR1020150143692A KR101764233B1 (en) 2015-10-14 2015-10-14 The manufacturing method of a dynamic vibration absorber

Country Status (1)

Country Link
KR (1) KR101764233B1 (en)

Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
KR101965800B1 (en) 2018-12-01 2019-07-26 주식회사 스마텍이앤씨 A seismic vibration-damped response reduction cabinet in which a tuning mass damping device is installed

Citations (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JP2003206979A (en) 2002-01-09 2003-07-25 Ishikawajima Harima Heavy Ind Co Ltd Vibration control method

Patent Citations (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JP2003206979A (en) 2002-01-09 2003-07-25 Ishikawajima Harima Heavy Ind Co Ltd Vibration control method

Non-Patent Citations (3)

* Cited by examiner, † Cited by third party
Title
고우식 외 4명. 전자기력에 의한 능동 동흡진기의 진동제어 성능에 관한 연구. 한국소음진동공학회. 1996.05
김사수 외 6명. 선박용 동흡진기 시스템에 관한 연구. 1995.08
송동호외 2명. 감쇠동흡진기의 최적설계. 2011.11

Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
KR101965800B1 (en) 2018-12-01 2019-07-26 주식회사 스마텍이앤씨 A seismic vibration-damped response reduction cabinet in which a tuning mass damping device is installed

Also Published As

Publication number Publication date
KR20170043961A (en) 2017-04-24

Similar Documents

Publication Publication Date Title
JP6669648B2 (en) Self-tuning mass damper and system including the same
CN110094452B (en) Broadband vibration suppression device utilizing acoustic black hole characteristics
Bakre et al. Optimum multiple tuned mass dampers for base-excited damped main system
Hosek et al. A tunable torsional vibration absorber: the centrifugal delayed resonator
US9500247B2 (en) Pounding tune mass damper with viscoelastic material
KR20120068119A (en) High damping horizontal dynamic absorber
Kecik et al. Non-linear dynamics and optimization of a harvester–absorber system
KR101764233B1 (en) The manufacturing method of a dynamic vibration absorber
JP2006077812A (en) Multiple dynamic vibration absorber designing method
KR101844386B1 (en) A vibration absorber
Marano et al. Stochastic optimum design of linear tuned mass dampers for seismic protection of high towers
US5586512A (en) Ship's hull vibration damper
CN103742580A (en) Method for adjusting intrinsic frequency of vibration isolation system of inertial navigation assembly
JPS58146744A (en) Vibration absorber
JP2007064353A (en) Swing damping device
KR20130073645A (en) High damping vertical dynamic absorber
KR101803742B1 (en) Dynamic Vibration Absorber
Hendrowati et al. Optimizing the value of reduction and generating energy on mechanism of cantilever piezoelectric vibration absorber (CPVA)
KR101301349B1 (en) Dynamic Absorber
CN108121166B (en) Active vibration absorber and micropositioner
JPS6217442A (en) Vibration damping structure
JP2004332847A (en) Damping device
CN215370780U (en) Space folding beam local resonance vibration damper and steering tie rod with built-in local resonance vibration damper
CN113076592B (en) Steering tie rod with built-in shock absorber and shock absorption performance parameter estimation method
KR102200322B1 (en) Damping device of beam structure for reducing shock having shape of the modified section

Legal Events

Date Code Title Description
A201 Request for examination
E902 Notification of reason for refusal
E701 Decision to grant or registration of patent right