RU2532235C2 - Vibration transporting machine - Google Patents

Abstract

FIELD: machine building.

SUBSTANCE: vibration transporting machine includes a working element connected via an elastic connection to a jet part carrying a resonant one-directional vibration device, and shock absorbers of low stiffness; besides, the resonant one-directional vibration device is made in the form of at least a pair of similar parametric vibration exciters mounted on the jet part of the machine and installed so that rotors of inertia elements can be rotated in opposite directions in vertical planes and brought into rotation from independent electric motors, and resonant frequency of the resonant one-directional vibration device is determined as follows: ω=λ₁+λ₂, λ₁=ν·ω, 0<ν<1, where ω - averaged value of partial angular speeds of rotors, λ₁ - effective proper frequency of swinging pendulums of rotors of inertia elements, $${\lambda }_2=\sqrt{C/M_{reduced}}$$

- partial proper frequency of the working element, which corresponds to an opposite-phase shape of one-directional free vibrations, M_reduced=M₁M₂/(M₁+M₂) - reduced mass, C - stiffness of an elastic connection, M₁ - mass of the working element, M₂ - total mass of the machine jet part.

EFFECT: achievement of high stability of a resonant operating mode of a machine at high quality of its vibrating system; as a result, creation of energy-saving vibration transporting machines.

2 dwg

Description

The invention relates to vibration technology and can be used in all industries for the transportation of various bulk and bulk materials in dustproof or airtight design and, if necessary, the simultaneous processing of bulk cargo (classification, dosing, mixing, drying, compaction).

The general name vibrotransport machine (VTM) refers to a group of machines: vibration conveyors, vibration feeders, vibrating screens, screen feeders, etc.

For VTM drive, inertial (unbalanced), eccentric, electromagnetic, hydraulic and pneumatic exciters are used.

Currently, the vast majority of VTMs operate in the forced oscillation mode with a resonant tuning. The resonant operation modes of VTMs, which are the most energetically efficient, are practically unrealizable due to their low stability under the usual resonance of forced oscillations.

Resonance machines with an electromagnetic drive of small capacity have also gained some popularity in industry. With the permissible size of the electromagnetic vibration exciter, the required amplitude of oscillations of the working body can be obtained only in the near-resonance mode, which is characterized by low stability. At the same time, there is a significant change in the amplitude of oscillations with a change in material on the working body, which leads to a decrease in the intensity of the process [Vibrations in technology in 6 volumes. T.4. Vibration processes. M .: Mechanical Engineering, 1982. S. 257].

In modern designs of resonant VTMs, self-synchronizing inertial (unbalanced) vibration exciters and motor vibrators are widely used, simplifying the layout and increasing the reliability of the design [I. Blekhman What can vibration do? About "vibrational mechanics" and vibrational technology. M .: Nauka, 1988. P.135].

In the article [Antipov V.I., Palashova I.V. Dynamics of a resonant transport-technological machine // Bulletin of the Nizhny Novgorod University. N.I. Lobachevsky. 2010.3 (1), pp. 141-147] a two-mass transporting machine with one parametric (resonant) vibration exciter is proposed. The disadvantage of this machine is that the driving force is constant in magnitude, but continuously changes in direction. As a result, additional “spurious” oscillations of the working body in the transverse direction are excited. To eliminate “spurious" oscillations, it is necessary that the driving force act in a straight line, periodically changing in magnitude and direction. This force can be obtained by using two identical parametric (resonant) vibration exciters rotating in opposite directions [RF Patent No. 2441714, MKI V06V 1/10. The method of excitation of resonant mechanical vibrations / Antipov V.I., Antipova R.I., publ. 02/10/2012. Bull. No. 4].

As a prototype, a two-mass VTM with a resonance tuning is adopted, containing a working body supported by elastic bonds on the reactive mass (on the reactive part of the machine), which is mounted on the base through shock absorbers of low stiffness, and a vibration exciter mounted on the reactive mass in the form of two paired self-synchronizing inertial (unbalanced) vibrators with directed driving force [Spivakovsky A.O., Goncharevich I.F. Vibratory conveyors feeders and accessories. M .: Mechanical Engineering, 1972, p.17, Fig. 6, c. Vibration in technology. Handbook in 6 volumes. T.4. Vibration processes. M.: Engineering, 1981, p. 308, Fig. 4, c].

The prototype has the following disadvantages.

- The large centrifugal forces created by the rotation of the unbalances load the vibration exciter bearings. This reduces their resource, entails a large unproductive energy consumption to overcome resistance to rotation of the shaft.

- The use of inertial (unbalanced) vibration exciters in low-frequency modes is irrational, since in this case it is necessary to significantly increase the mass of unbalances to obtain a sufficiently large amplitude of the driving force.

- In the case of resonance tuning, in order to overcome the region of intense oscillations, it is necessary to have an engine whose power is 5-6 times higher than the power needed to operate in the resonance mode.

- In the resonant mode of operation of the VTM, the energy of the vibration exciter is consumed not only to overcome the friction forces, but also to overcome the inertia forces of the masses of the system.

These shortcomings are eliminated by the proposed solution. The problem of creating a fundamentally new energy-saving multivibrator VTM at multiple Raman parametric resonance with the expansion of functional and operational capabilities is being solved.

The technical result is the achievement of high stability of the resonant mode of operation of the machine with a high quality factor of its oscillatory system, which opens up great opportunities for creating energy-saving VTM with vibration isolation and dynamic balancing of oscillating masses with a synergistic effect, i.e. mutual stimulation of mass oscillations of the system.

- парциальная собственная частота рабочего органа, соответствующая противофазной форме однонаправленных свободных колебаний, M_пр=M₁M₂/(M₁+M₂) - приведенная масса, C - жесткость упругой связи, M₁ - масса рабочего органа, M₂ - общая масса реактивной части машины.This technical result is achieved by the fact that in a vibrating transporting machine, including a working body, connected by an elastic connection with a reactive part, carrying a means for communicating resonant unidirectional vibrations, and shock absorbers of low stiffness, a means for communicating resonant unidirectional vibrations is made in the form of machines mounted on the reactive part at least a pair of identical parametric vibration exciters installed with the possibility of rotation of the rotors of inertial elements in the opposite false directions in vertical planes and driven by independent electric motors, and the resonant frequency of the means for communicating resonant unidirectional oscillations is determined from the relations ω = λ ₁ + λ ₂ , λ ₁ = v · ω, 0 <v <1, where ω is the averaged the value of the partial angular velocities of the rotors, λ ₁ is the effective natural frequency of the swinging pendulums of the rotors of inertial elements, $${\lambda }_2=\sqrt{C/M_{PR}}$$

- partial natural frequency of the working body, corresponding to the antiphase form of unidirectional free vibrations, M _CR = M ₁ M ₂ / (M ₁ + M ₂ ) - reduced mass, C - stiffness of the elastic bond, M ₁ - mass of the working body, M ₂ - total mass of the reactive part of the machine.

The proposed TMV simultaneously manifests the effects of both pendulum and rotor self-synchronization.

A diagram of a vibrating transporting machine is shown in FIGS. 1, 2. Two masses of the machine M ₁ and M ₂ , one of which is a working transporting body 1, and the other is reactive 2, are interconnected by a main elastic bond 3 of rigidity C (FIG. 1) . A means for communicating unidirectional resonant oscillations in the form of at least a pair of identical self-synchronizing parametric vibration exciters 4 with the possibility of rotation in opposite directions is fixed on the reactive mass. The working transporting body 1 rests on the base 5 through shock absorbers 6 of low stiffness (auxiliary elastic bonds). The planes of rotation of the rotors are located in vertical planes.

For operating vibrations, resonance vibrations are taken at a frequency λ ₂ corresponding to the translational antiphase form of free vibrations of masses M ₁ and M ₂ in the direction of the ox axis.

The partial natural frequency of the working body is determined by the formula:

, $${\lambda }_2=\sqrt{C/M_{PR}}$$

,

where M _CR = M ₁ M ₂ / (M ₁ + M ₂ ) is the reduced mass of the system.

For free vibrations, the relation holds:

A ₁ / A ₂ = M ₂ / M ₁ = l ₁ / l ₂ ,

where A ₁ , A ₂ are the amplitudes of the oscillating masses M ₁ and M ₂ , l ₁ , l ₂ are the coordinates of the center of mass of the system.

It is assumed that the form of stationary resonant oscillations coincides with the form of free vibrations (Widler's postulate).

Figure 2 shows a diagram of an inertial element (IE) of a parametric vibration exciter.

The rotor of this system is recruited from separate identical balanced discs 7 with drum parts. In each disk, a pair of open racetracks 8 of a circular profile is formed, which are located symmetrically with respect to two mutually perpendicular diameters, and their centers are offset from the axis of rotation of the rotor in diametrically opposite directions by equal distances AB = l. On treadmills are placed the same balanced rolling bodies (pendulums) 9 with the possibility of break-in. The disks are connected in a single design so that the break-in axes in adjacent disks are rotated around the rotor axis of rotation by the same angle γ = π / s, where s is the number of disks (an even number). The IE rotor contains N = 2s rolling elements.

The number of disks is set depending on the power of the machine and the required amplitude of the oscillations of the working body. The rotors of two identical IEs in assembled form are rigidly mounted on the cantilever ends of the shaft of an asynchronous type electric motor or a DC motor with the same orientation of the treadmills. In this case, the vibration exciter is a motor-vibrator and is included in the oscillatory system of the vibrator. In other cases, a vibration exciter is preferred that receives rotation from a separate electric motor removed from the oscillatory system of the machine.

Further, we will assume that s = 2, and the parametric oscillator is a vibrator motor. The coordinate system o'x'y'z 'with the beginning in the center of mass of the rotor-pendulum system (without rolling elements) of each of the vibration exciters moves progressively relative to the stationary oxyz, and the z axis is parallel to the axes of rotation of the rotors. A movable system Ax'y'z 'is connected with the center of each disk of the rotor-pendulum system, the axes of which are parallel to the corresponding axes of the system o'x'y'z' (Fig. 2).

. Здесь N - число беговых дорожек одного ротора вибровозбудителя, ω⁽ⁱ⁾ - парциальная частота вращения i-го вибровозбудителя, i=1,2. За обобщенные координаты принимают углы $${\varphi }_k^{(i)}$$

, определяющие положения тел качения одного ротора i-го вибровозбудителя, и перемещения x₁, x₂ масс M₁, M₂. Предполагается, что тела качения другого ротора определяются теми же углами.The vertical plane oxy is taken as the main plane, relative to which the IE rotors make a plane motion. Unidirectional vibrations of the active and reactive masses in the direction of the ox axis are considered. The necessary shape of the trajectory is provided by the directed stiffness of the main elastic system, for example, a low-sheare plane spring elastic system. The orientation of the centers of curvature of the treadmills (break-in axes) is determined by the angles $${\psi }_k^{(i)}={\omega }_t^{(i)}+2\pi k/N,k=\mathrm{1,2}...N,N=\mathrm{four}$$

. Here N is the number of treadmills of one rotor of the vibration exciter, ω ⁽ⁱ⁾ is the partial frequency of rotation of the i-th vibration exciter, i = 1.2. For the generalized coordinates take the angles $${\varphi }_k^{(i)}$$

determining the position of the rolling bodies of one rotor of the i-th vibration exciter, and the movement x ₁ , x ₂ masses M ₁ , M ₂ . It is assumed that the rolling bodies of the other rotor are determined by the same angles.

- безразмерный коэффициент, m - масса тела качения, J_B - момент инерции тела качения относительно оси обкатки, l=AB, ρ_c=BC_k. Вторую подсистему осцилляторов образуют упругосвязанные массы M₁ и M₂.With a uniform rotation of the vibration exciters, each of them forms a subsystem of 2N identical rolling oscillators (pendulums) in the field of centrifugal inertia forces with rolling axes at the center of curvature of the treadmills and the same partial natural frequencies λi = v · ω _i in the coordinate system rotating with the rotor. Here $$v=\sqrt{m{\rho }_{c}l/J_B}$$

is the dimensionless coefficient, m is the mass of the rolling body, J _B is the moment of inertia of the rolling body relative to the running axis, l = AB, ρ _c = BC _k . The second subsystem of oscillators is formed by elastically coupled masses M ₁ and M ₂ .

самовозбуждается многократный комбинационный параметрический резонанс. Здесь ε=v²Nµ/2, µ=mρ_c/(M₂l), σ=2; $${\tilde{n}}_0=n_0/{\lambda }_2$$

, $$\tilde{n}=n/{\lambda }_2$$

- относительные коэффициенты линейного демпфирования соответственно осцилляторов качения и масс M₁, M₂. Вследствие качаний маятников ИЭ их общий центр масс описывает окружность в плоскости o'x'y'. Угловая скорость его вращения Ω₂ по этой окружности равна приблизительно частоте λ₂ свободных колебаний рабочего органа. Иначе говоря, автоматически образуется неуравновешенность ИЭ «невидимый дебаланс». Поскольку Ω₂≈λ₂, то неуравновешенная центробежная сила инерции будет возбуждать резонансные колебания масс M₁ и M₂, а колебания этих масс вызывают колебания осцилляторов качения (маятников). В результате реализуется синергетический эффект, т.е. взаимное стимулирование колебаний элементов системы. В резонансной зоне выполняется соотношение ω=Ω₁+Ω₂, где Ω₁≈v·ω, Ω₂≈λ₂ - некратные частоты генерации осцилляторов качения и рабочего органа соответственно. При этом осцилляторы качения самосинхронизируются по Гюйгенсу, автоматически компенсируя погрешности изготовления роторно-маятниковой системы.Consider the case of one parametric vibration exciter. Let the IE rotors rotate with a constant angular velocity ω. The uniform rotation of the rotors leads to a periodic change in time of the inert properties of the vibrational system of the machine with a period of 2π / ω. When setting ω = λ ₁ + λ ₂ , where λ ₁ = v · ω, v = 0.25, and the threshold condition is satisfied $$\epsilon >\sigma 8v{\tilde{n}}_0\tilde{n}/(\mathrm{one}-v)$$

multiple Raman parametric resonance self-excited. Here ε = v ² Nµ / 2, µ = mρ _c / (M ₂ l), σ = 2; $${\tilde{n}}_0=n_0/{\lambda }_2$$

, $$\tilde{n}=n/{\lambda }_2$$

- the relative coefficients of linear damping, respectively, of the rolling oscillators and masses M ₁ , M ₂ . Due to the swings of the IE pendulums, their common center of mass describes a circle in the o'x'y 'plane. The angular velocity of its rotation Ω ₂ along this circle is approximately equal to the frequency λ _{2 of} free vibrations of the working body. In other words, the imbalance of IE “invisible unbalance” is automatically formed. Since Ω ₂ ≈λ ₂ , the unbalanced centrifugal inertia force will excite the resonant oscillations of the masses M ₁ and M ₂ , and the oscillations of these masses cause oscillations of the rolling oscillators (pendulums). As a result, a synergistic effect is realized, i.e. mutual stimulation of oscillations of system elements. In the resonance zone, the relation ω = Ω ₁ + Ω ₂ holds, where Ω ₁ ≈v · ω, Ω ₂ ≈λ ₂ are the multiple frequencies of the generation of the rolling oscillators and the working body, respectively. In this case, the rolling oscillators self-synchronize according to Huygens, automatically compensating for the manufacturing errors of the rotor-pendulum system.

In the zone of multiple Raman parametric resonance, the TMV acquires new properties.

When parametric vibration exciters work together with close partial angular velocities ω ⁽¹⁾ , ω ⁽²⁾ , they self-synchronize as unbalanced rotors. As a result, a synchronous mode is established in which the common center of mass of the swinging pendulums of each of the vibration exciters “invisible unbalance” rotates around their axes with the same angular velocity Ω ₂ ≈λ ₂ in opposite directions, informing the working body of linear oscillations, while the angular velocities of the rotors themselves can to be different. When the “invisible unbalances” rotate, a total driving force arises in the direction of the ox axis, periodically changing with a frequency of Ω ₂ ≈λ ₂ ≈0.75ω, where ω = (ω ⁽¹⁾ + ω ⁽²⁾ ) / 2 are the average values of the partial angular velocities vibration exciters. When setting ω = λ ₁ + λ ₂ , λ ₁ = v · ω, v = 0.5, rectilinear vibrations of the working body with a frequency of Ω ₂ ≈λ ₂ ≈0.5ω are excited.

When parametric vibration exciters work together, the VTM setting does not formally change, but the value of ω should mean the average angular velocity of the vibration exciters.

Currently, interest has increased in the problem of reducing the oscillation frequency of TMV. This is due to the fact that the intensification of some technological processes is associated with the use of low-frequency mechanical vibrations, for example, for feeding fine materials, nanopowders.

The generation of low-frequency oscillations can be carried out by setting ω = λ ₁ + λ ₂ , where λ ₁ = v · ω, v = 0.5. In this case, rectilinear oscillations with a frequency of Ω ₂ ≈λ ₂ ≈0.5ω self-excite. The decrease in the frequency of oscillations of the working body of the VTM at a constant amplitude reduces the speed of transportation by approximately the same amount as the frequency decreases. Therefore, to maintain the required speed, the amplitude of the oscillations should be increased, which is not a problem for a resonant machine.

The proposed resonant vibration transporting machine has important advantages that give it new qualities.

. Снижается также шум машины.1. Reproduction of stable resonant oscillations of the working body of large amplitude and low frequency, and the frequency reduction is provided without the use of additional frequency converters. A halving of the oscillation frequency leads to a halving of the inertia forces. The load on the supporting structures is reduced by the same amount. In addition, the rigidity of an expensive elastic system is reduced by four times $$(\mathrm{one}/{[{\lambda }_2^{\ast }/{\lambda }_2)}^2]=\mathrm{four,}{\lambda }_2^{\ast }/{\lambda }_2=\mathrm{one}/2)$$

. Car noise is also reduced.

2. The performance of the VTM easily changes on the go from zero to maximum by changing the amplitude of the oscillations. In this case, the amplitude of the oscillations can be regulated over a wide range due to the smooth change in the frequency of rotation of the vibration exciters. To stop the supply of material, it is enough to remove the machine from the resonance zone (without turning off the vibration exciters). This makes it possible to effectively use resonant VTM in automated lines, as well as in cases where a metered or controlled supply of material is required.

3. Resonance VTM is characterized by low energy consumption. The installation power of the vibro drive is reduced by more than two times in comparison with the inertial or kinematic vibration exciters used in industry.

4. The TMV's ability to self-synchronize and self-organize resonant oscillations, i.e. the machine includes elements of technical intelligence.

5. The proposed TMV has the advantages of machines with inertial (unbalanced) and electromagnetic vibration exciters and does not have their drawbacks.

The analysis shows that the proposed solution meets the criteria of "novelty", "inventive step" and "industrial applicability".

Claims (1)

Vibration transporting machine, comprising a working body, connected by an elastic connection with a reactive part, carrying means for communicating resonant unidirectional vibrations, and shock absorbers of low stiffness, characterized in that the means for communicating resonant unidirectional vibrations is made in the form of at least mounted on the reactive part of the machine , pairs of identical parametric vibration exciters installed with the possibility of rotation of the rotors of inertial elements in opposite directions in vert cial planes and driven in rotation by independent motors, and the resonant frequency means for communicating the resonance unidirectional vibrations is determined from the ratios of ω = λ ₁ + λ _2, λ ₁ = ν · ω, 0 <ν <1, where ω - the average value of the partial angular rotor speeds, λ ₁ is the effective natural frequency of the swinging pendulums of the rotors of inertial elements,

- partial natural frequency of the working body, corresponding to the antiphase form of unidirectional free vibrations, M _CR = M ₁ M ₂ / (M ₁ + M ₂ ) - reduced mass, C - stiffness of the elastic bond, M ₁ - mass of the working body, M ₂ - total mass of the reactive part of the machine.

Images