RU126782U1 - DYNAMIC VIBRATION DAMPER FOR ROTARY MACHINES - Google Patents

Abstract

1. Dynamic vibration damper containing an elastic rod and an inertial mass made with a Central hole for placement on the rod with the ability to move along it, the rod at one end is equipped with a fastening element for cantilever fastening of the dynamic vibration damper on a rotary machine, characterized in that for movement the inertial mass along the rod on the inner surface of the Central hole of the inertial mass and on the outer surface of the rod corresponding thread is made, while for The inertial mass relative to the axis of the shaft is gear and a gear motor, the gear consists of a driven gear and longitudinal grooves on the shaft, made by thread, the gear motor consists of a direct current motor and a pinion gear, and the motor is powered from a direct current source switched through a wireless control device, and the DC source, the wireless control device and the DC motor are placed in ubleniyah made in the case of inertial massy.2. The dynamic vibration damper according to claim 1, characterized in that the elastic rod comprises a nipple fastening element on which a thread is cut. The dynamic vibration damper according to claim 1, characterized in that the recesses made in the inertial mass body are equipped with plugs for securing the direct current source, wireless control device and direct current electric motor located therein. The dynamic vibration damper according to claim 3, characterized in that the inertial �

Description

The utility model relates to mechanical engineering and is designed to absorb negative mechanical vibrations on elements of dynamically loaded machines.

Most of the processes occurring in nature and technology are oscillatory. Such processes include a very wide range of the most diverse phenomena from vibration of a sounding string to earthquakes. The term vibration is most often used where vibrations have a relatively small amplitude and not too low frequency. At present, the elimination of negative vibrations is of particular importance in connection with the growth of capacities of rotary machines and in accordance with the requirement to increase their durability and reliability, to ensure stability and controllability of mechanical systems.

Due to the possibility of a significant reduction in the cost of servicing equipment and preventing accidents, many enterprises for the extraction and transportation of oil and gas are switching to servicing and repairing machinery and equipment in actual condition. The basis of such a transition is vibration diagnostics, which allows at any time to have complete information about the state of the operating equipment. Vibrodiagnostics allows you to:

- organize incoming quality control of components (for example, rolling bearings) and machines (for example, pumps, electric motors);

- assess the condition of the equipment;

- prevent the progressive wear of the units, eliminate their breakdowns and premature failure and ensure control over the quality of repair work

The vibration signal in most cases is very complex and carries a large amount of information about the technical condition of the object under examination. Correctly organized measurements of vibrational parameters make it possible to identify most defects even at the initial stage of their appearance. Modern devices produce signal processing with sufficient quality, but, like any diagnostics, this type of signal is somewhat probabilistic. Moreover, as the amount of information about each particular unit has been accumulated, the probability of the correct detection of a defect will constantly increase.

In addition to the vibration signal directly taken from the unit, it is important to control other parameters characterizing its operation, for example, temperature, load, and pumping modes (for pumps). In particular, in the diagnosis of rolling bearings, cavitation processes occurring in pumps can become a factor that reduces the reliability of the conclusion.

Vibrational processes are always associated with resonances. For working units, oscillations occur under the action of a constant driving force, and taking into account that in rotary machines this is rotational movement, the force in projections onto the selected directions (usually vertical, transverse and axial) has a harmonic character.

Each mechanical system has a so-called natural frequency, with which the driving force coincides, a resonance phenomenon occurs, that is, an unlimited increase in the oscillation amplitude, although for real systems the amplitude is, of course, limited. An increase in the amplitude most often leads to negative consequences, and if the unit, due to its design features, has resonance in the work area, an increased level of vibration can lead to destruction of both individual work units and the unit as a whole. In order to avoid such consequences in mechanical systems, in particular in rotary machines, detuning from resonance and damping are carried out. Damping reduces the intensity of vibration at and near the resonant frequency. Detuning from resonance is done by changing the rigidity of the support system or mass, which requires structural changes and, as a rule, is unacceptable. A decrease in stiffness leads to a shift of the resonance to the low-frequency region, and an increase to a high-frequency region.

The pump unit is a rather complex mechanical system for which there are a large number of resonant frequencies of different directions. And although at the design stage, detuning from resonances is provided, in reality, resonant phenomena can often be observed on operated units.

It can be noted that a third of diagnosed and repaired machines with a capacity of 3-5 MW show resonance to one degree or another at the front bearing support with a frequency multiple of the second reverse. To identify resonances, a technique is used to record the speed characteristics of the rotor of the electric motor on the coast, as shown in figure 1. Here the main signs of resonance are clearly visible: a change in amplitude and phase. Moreover, we can conclude that the unit operates precisely at the resonant frequency, since at the time of shutdown there is a sharp drop in amplitude by 10 times with a simultaneous phase rotation of more than 100 °.

This situation is typical for a sufficiently large number of pumping units of this class.

The most common designs of dynamic vibration dampers (antivibrators) are rod structures, consisting of loads mounted on the rod. (See, for example, SU 1490344 from 06/30/1989, F16F 15/00 or SU 1357626 from 12/07/1987, F16F 15/10). In particular, the dynamic vibration damper disclosed in the copyright certificate SU1490344 contains an elastic rod connected to the protected object with an inertial mass mounted on it. In order to increase the efficiency of damping oscillations, it is equipped with a sleeve placed coaxially and installed on the free end of the sleeve, and the inertial mass is mounted on the sleeve with the possibility of axial movement. In addition, the sleeve is mounted on the rod with the possibility of axial movement and fixation. A dynamic vibration damper has two spaced resonant vibration frequencies, one of which is determined by the distance from the point of incorporation of the elastic rod to the center of mass, and the second by the distance from the point of incorporation of the elastic rod to the point of attachment of the sleeve on it, thus damping the vibrations of the protected object at two different frequencies. The resonant frequencies of the vibration damper are adjusted when the sleeve is moved relative to the elastic rod and the mass relative to the sleeve. The disadvantage of this device is that during operation requires constant adjustment, performed manually.

In the design of rod dynamic vibration dampers (anti-vibrators), the rod is an elastic element. The dynamic vibration damper is adjusted by moving the load along the rod. As a result of this, the distance l from the center of mass to the attachment point changes, and hence the anti-vibration stiffness determined by the formula:

,

,

where E is the elastic modulus of the rod material;

I is the moment of inertia of the rod section;

l is the distance from the center of mass to the attachment point.

The main disadvantage of using the considered dynamic vibration dampers is that they turn out to be effective in a relatively narrow frequency range due to the small mass values of antivibrators allowed in mechanical engineering. In this regard, their use gives a great effect only for machines with a fairly narrow range of speed control. The above designs are usually installed directly on the machine body and are designed to eliminate body vibrations.

A way out of this situation is the use, for example, of dynamic vibration dampers with adjustable parameters.

The closest analogue of the claimed technical solution is a dynamic vibration damper, disclosed in the description of the patent of the Russian Federation for utility model RU 109814 U1, published on 10.27.2011 (IPC F16F 7/104). A dynamic vibration damper includes an elastic rod and an inertial mass made with the possibility of moving the inertial mass along the rod. The rod is a console, which consists of a nipple part into which the thread is cut, and a hole of different diameters is drilled in the center along the rod to accommodate the drive mechanism and cable wiring. Also, a rod with an external thread is fixed along the rod with the possibility of rotation around its axis, and a counter-hole with an internal thread for the rod is made in the inertial mass. To rotate the rod, an electric motor is included in the design of the device. To transmit torque to the rod, a planetary gearbox is installed.

The effectiveness of the dynamic vibration damper described above can be significantly impaired by its “detuning” due to deviations from the calculated values of stiffness or mass. The anisotropy of the device’s properties, associated with the presence of a longitudinal rod mounted along the rod on one side, also causes a negative effect, which causes, in particular, a difference in natural frequencies in the directions orthogonal to the axis of the anti-vibrator.

The utility model is aimed at solving the problem of creating a more effective dynamic vibration damper for rotary machines.

The technical result of the utility model is to increase the vibration damping efficiency of rotary machines by increasing the isotropic properties of the device while providing automatic and remote control of the vibration damping process.

To solve this problem, a dynamic vibration damper is declared, containing an elastic rod and an inertial mass made with a central hole for placement on the rod with the ability to move along it, the rod at one end is equipped with a fastener for cantilever fastening of the dynamic vibration damper on a rotary machine. A corresponding thread is made to move the inertial mass along the rod on the inner surface of the central hole of the inertial mass and on the outer surface of the rod, while gearing and a gear motor are used to rotate the inertial mass relative to the axis of the rod, the gearing consists of a driven gear and longitudinal grooves on the rod made by thread, the gear motor consists of a DC motor and pinion gear, and the motor is powered from a source tinuous current through the switched wireless control device, wherein the DC power source, a wireless control apparatus and the DC motor are arranged in recesses made in the housing of the inertial mass.

The elastic rod of the dynamic vibration damper contains a nipple fastening element on which the thread is cut.

The recesses made in the inertial mass body are equipped with plugs for fixing the direct current source, wireless control device and direct current electric motor placed in them.

The inertial mass body of the claimed device is made in the form of a massive cylinder with a central hole and is additionally equipped with at least one protective cover that covers structural elements located in the housing openings, namely, a direct current source, a wireless control device, and a direct current electric motor, fixed with plugs. The protective cover is made in the form of a ring with a shoulder and completely covers the end surface of the cylindrical body of inertial mass.

Structural elements of a dynamic vibration damper, namely: a direct current source, a wireless control device, and a direct current electric motor, fixed with plugs, are placed in the openings of the inertial mass casing under the condition of their uniform distribution and maximizing the isotropy of the properties of the massive inertial mass casing.

Figure 2 shows a dynamic vibration damper disassembled.

Figure 3 shows the dynamic vibration damper assembly.

Dynamic vibration damper contains an elastic rod 1 and an inertial mass 2 of a cylindrical shape made with a central hole for placement on the rod with the possibility of movement. The body of the inertial mass contains several recesses to accommodate the DC source 3, which is fixed in it by a plug 4. Under the plug 5 there is a DC motor 10 connected to the pinion 6, which functions as a gear motor. In the next recess, the wireless control device 7 is mounted, fixed by a plug 11. The drive gear 6 interacts with the driven gear 9, mounted on the axis 8, mounted on the body of the inertial mass 2. To rotate the inertial mass 1, gearing of the driven gear 9 and the longitudinal grooves on the shaft is used 1, made according to the thread applied to the shaft 1. That is, the protrusions of the thread applied to the shaft 1 are cut by longitudinal grooves to a depth and in increments corresponding to the teeth of the gear 9. All elements of the mechanism and conveying the inertial mass of the dynamic vibration reducer rod are closed by mechanical damage and contamination of the protective cover 12.

The claimed utility model is designed to reduce the level of vibration at the working speed of rotary machines. To install this mechanism does not require any changes in the design of the protected unit, which, undoubtedly, is the advantage of this type of vibration adjustment.

To adjust the natural frequency of the dynamic vibration damper, the inertial mass 2 moves along the elastic rod 1 using remote control. The movement occurs along the thread by rotation of the inertial mass 2. The rotation of the inertial mass 2 relative to the axis of the rod 1 is carried out by the interaction of the gearing and the gear motor. The engine is powered from a direct current source 3, switched through a wireless control device 7.

The principle of operation is based on the fact that forced oscillations at a frequency greater than the resonant frequency occur in antiphase with respect to the driving force, i.e. if the resonant frequency of the dynamic vibration damper is slightly less than the oscillation frequency of the protected object, then the oscillations will be damped.

In the case of an ideal antivibrator that works without friction, the system has the property that, at a speed numerically equal to the partial frequency of the antivibrator, the amplitude of movement of the body vanishes, which allows it to be used at fixed speeds.

However, there is always damping in the most dynamic vibration damper, which makes it fundamentally impossible to zero the movements of the body of the rotary machine. Provided that such damping is small, approximate estimates can be obtained,

,

,

β is the operating frequency,

µ is the ratio of the mass of the vibration damper to the mass of the protected object,

E is the resulting imbalance,

A is the amplitude of the vibrations,

which allow you to understand the main features of the dynamic vibration damper:

1) the vibration level of the housing is almost directly proportional to the damping forces in the dynamic vibration damper and inversely proportional to their mass;

2) the movement of the dynamic vibration dampers themselves is determined mainly by their mass.

The effect of using a dynamic vibration damper is determined when comparing the vibration of systems without anti-vibration and anti-vibration. The efficiency L (in dB) is taken as

,

,

where A ₀ is the amplitude of vibrations without a dynamic vibration damper,

A - with a dynamic vibration damper.

The effectiveness of the dynamic vibration damper is maximum at an operating speed that matches the natural frequencies of the system. And it is essentially broken at their "detuning". There is also the fact that the higher the efficiency of tuning a dynamic vibration damper, the sharper it falls during detuning.

The calculation of the dynamic dynamic vibration damper consists in determining its structural dimensions and mass, which would satisfy the condition of reducing vibration to an acceptable level.

As a design scheme, a weightless cantilever beam is used, oriented strictly vertically, with a point mass at the upper end. Then you can compose the equation of motion of such a system, and determine the natural frequency, for which the dynamic vibration damper will work.

We use the Lagrange equations in the form:

,

,

where T is kinetic, and U is the potential energy of the system, which in turn looks like:

where α is the angle of deviation of the rod from the vertical,

k is the stiffness of the rod,

m is the mass of the point load,

l is the length of the rod,

g is the acceleration of gravity.

After differentiation, we obtain the Lagrange equation:

,

,

It can be seen that the natural frequency of a given system is defined as:

Now all the parameters of the dynamic vibration damper are determined theoretically, however, in reality, it is necessary to adjust the real device, and this can be done in two ways, firstly by changing the height of the load position, secondly, you can use a rod with different stiffness in the directions, in the aggregate, this will give great opportunities for settings.

There are also empirical formulas for choosing the diameter of the rod and its length depending on the frequency of oscillations, which must be eliminated.

,

,

m is the mass in kg; d is the diameter in cm; k is the frequency coefficient, k = f / 50, f is the oscillation frequency of Hz.

,

,

L is the length of the rod.

For d = 4 cm, f = 100 Hz, k = 2, therefore, m = 2 * (4/2) ² = 8 kg

Then

The maximum restoring force for such a damper will be:

F = 0.1 d ³ [σ _{_-1} ] / L, where [σ _-1 ] is the allowable stress during an alternating cycle for steel 45 is approximately equal to 1000 kg / cm ² then in our case we get

F = (0.1 * 2 ³ * 1000) / 29 = 220 kg.

The obtained calculations show that in order to damp the vibration on the bearing pedestals of the rotary machine in the axial direction at a frequency of 100 Hz, a dynamic absorber with a length of 350 mm, a diameter of 40 mm, and an inertial mass of 8 kg is required.

Since the calculations are based on a model design scheme with a number of assumptions, the shaft length is 350 mm, which is somewhat overestimated, which is subsequently used to adjust the dynamic vibration damper.

In the process of tuning, it is necessary to achieve a minimum value of the amplitude of vibration. In our case, the most optimal is the change in the active length of the rod, i.e. the distance from the base to the center of gravity of the load by moving the inertial mass along the axis of the rod.

To install this mechanism is not required to make changes to the design of the unit. This approach to vibration adjustment is most effective because it allows you to approach vibration in a comprehensive way and take into account the fact that vibration is a continuous volume field distributed throughout the bearing housing in all directions and to predict in advance how the vibration will be redistributed in projections to standard directions (vertical, transverse, axial) is almost impossible.

In the process of tuning, continuous monitoring (time response) of the following vibration parameters is carried out:

1. The amplitudes and phases of the harmonic components of the vibration velocity in the desired direction on the bearing support.

2. The amplitudes and phases of the harmonic components of the vibration velocity in parallel to the axis of the desired direction on the inertial mass.

3. General vibration levels in 3 orthogonal directions on the bearing support.

The setup is reduced to minimizing the amplitude of the vibration velocity of the resonant harmonic on the bearing pedestal in the required direction with the simultaneous antiphase state of this harmonic of the bearing support and inertial mass.

The second condition for tuning is to reduce the overall level of vibration velocity in the desired direction.

When these two conditions are met, the effect of the dynamic force of the dynamic vibration damper is considered positive.

Claims (4)

1. Dynamic vibration damper containing an elastic rod and an inertial mass made with a Central hole for placement on the rod with the ability to move along it, the rod at one end is equipped with a fastening element for cantilever fastening of the dynamic vibration damper on a rotary machine, characterized in that for movement the inertial mass along the rod on the inner surface of the Central hole of the inertial mass and on the outer surface of the rod made the corresponding thread, while for The inertial mass relative to the axis of the shaft is gear and a gear motor, the gear consists of a driven gear and longitudinal grooves on the shaft, made by thread, the gear motor consists of a direct current motor and a pinion gear, and the motor is powered from a direct current source switched through a wireless control device, and the DC source, the wireless control device and the DC motor are placed in ubleniyah made in the body of inertial mass. 2. The dynamic vibration damper according to claim 1, characterized in that the elastic rod comprises a nipple fastening element on which a thread is cut. 3. The dynamic vibration damper according to claim 1, characterized in that the recesses made in the inertial mass housing are provided with plugs for securing the direct current source, wireless control device and direct current electric motor located therein. 4. The dynamic vibration damper according to claim 3, characterized in that the inertial mass, made in the form of a cylinder with a central hole, is additionally provided with at least one protective cover at the end.

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