WO2019200426A1

WO2019200426A1 - Mountings for vibrating machines and methods of isolating vibrations

Info

Publication number: WO2019200426A1
Application number: PCT/AU2019/050335
Authority: WO
Inventors: Stephen Colin Brown
Original assignee: S and V Brown Holdings Pty Ltd
Priority date: 2018-04-20
Filing date: 2019-04-15
Publication date: 2019-10-24

Abstract

Mountings for vibrating machines which vibrate at a pre-determined frequency are described, the mounting including: a sprung mass which is arranged to oscillate out of phase with the vibrating machine at about the pre-determined frequency. The sprung mass may be arranged to oscillate approximately 180 degrees out of phase with the vibrating machine.

Description

MOUNTINGS FOR VIBRATING MACHINES AND METHODS OF ISOLATING

VIBRATIONS

Technical Field

The present invention relates to mountings for vibrating machines and methods of isolating vibrations. The invention finds a particular application in isolating buildings from vibrations causes by industrial machinery such as vibrating screens.

Background to the Invention

Many types of industrial processes utilise vibrating machinery of some type.

For instance, in mineral processing operations it is common utilise vibrating screens known as banana screens for washing and/or grading the particle sizes of various bulk materials. These screens operate by rotating an out of balance mass to cause vibration of the screen. These machines are housed in industrial buildings.

Operating vibrating machinery causes vibration transmission into the building structure. This can lead to fatigue failures of building and support structures. In addition, there are impacts on human health due to physiological responses to floor vibration which bring health risks, decreased proficiency and operator discomfort. The problem is further complicated by installations of multiple systems running at similar speeds and resulting in beating and reinforced vibration effects

Various approached have been tried to alleviate the issues outlined above, each of which have their own drawbacks as follows:

Soft springs - by mounting machinery on soft springs, the softer the springs the less dynamic force is transmitted. However, this is rarely a practical fix as there is a practical limit to how soft the springs can be due to stability and impractical static deflections.

Air bag suspension - can provide softer springs without the high static deflections however to date these types of systems have not been a serviceable improvement as the risk of damage and deflation of the air bags is high.

Secondary inertia base with secondary set of soft springs- can further reduce transmitted force however is very costly and requires additional space. Also adds significant mass.

Structural stiffening - stiffening the building in which machinery is housed is very costly and difficult to retrofit.

Tuned Mass Dampers - when added to the support structure can reduce the resulting vibration response of the support structure, in practice the effectiveness of these devices is limited by available space, the need for accurate tuning, the typical sub optimal placement on the structure in reference to the locations of the force input, as well as high cost. Tuned mass dampers can be considered as a type of dynamic structural stiffening.

Conventional isolation systems have been shown to reduce dynamic force transmission by up to 97%. However, the residual transmission is often the cause of many problems such as described above.

There remains a need to provide an improved or alternative approach to mitigating against the impacts of vibrations caused by vibrating machinery.

Summary of the Invention

In a first aspect the present invention provides a mounting for a vibrating machine which vibrates at a pre-determined frequency, the mounting including: a sprung mass which is arranged to oscillate out of phase with the vibrating machine at about the pre-determined frequency.

The sprung mass may be arranged to oscillate approximately 180 degrees out of phase with the vibrating machine.

The sprung mass may include at least one weighted element which is associated with at least one spring.

The at least one spring may be a coil spring.

The spring may be mounted in use at an angle to the vertical direction.

At least a portion of the weighted element may be disposed within the coil spring.

The mounting may further include first and second engagement portions for engaging with a vibrating machine and a support structure.

In a second aspect the present invention provides a vibrating machine including a mounting according to the first aspect of the invention.

The vibrating machine may include at least two mountings according to the first aspect of the invention.

The at least two mountings may share the same sprung mass. In a third aspect the present invention provides method of at least partially isolating a support structure from vibrations caused by a vibrating machine which is arranged to vibrate at a pre-determined frequency including the step of: installing at least one sprung mass between the support structure and the vibrating machine, wherein the sprung mass is arranged to oscillate out of phase with the vibrating machine at about the pre-determined frequency.

Brief Description of the Drawings

Embodiments of the present invention will now be described, by way of example only, with reference to the accompanying drawings, in which:

Figure 1 is a perspective view of a mounting according to a first embodiment of the invention;

Figure 2 is a cross sectional view of the mounting of figure 1;

Figure 3 is a side view of a banana screen fitted with mountings according to figure 1;

Figure 4 is a perspective view of the banana screen of figure 3 attached to a support structure in a building;

Figure 5 is a schematic representation of a prior art mounting system;

Figure 6 is a schematic representation of a mounting system according to the operating principles of the invention;

Figure 7 shows the frequency response of a traditional prior art isolation system.

Figure 8 shows the frequency response of a system according to the operating principles of the invention;

Figure 9 is a perspective view of a mounting according to a second embodiment of the invention;

Figure 10 is a cross sectional view of the mounting of figure 9;

Figure 11 shows the frequency response of the screen of figure 3 fitted with mountings according to figure 9;

Figure 12 is a cross sectional view of a mounting according to a third embodiment of the invention; and

Figure 13 is a schematic representation of a vibrating machine fitted with mountings, two of which share the same sprung mass.

Detailed Description of Preferred Embodiments

Referring to figures 1 and 2, a mounting 10 is shown which is suitable for use with a vibrating machine such as a banana screen. Mounting 10 includes engagement portions ion the form of upper plate 12 and a lower plate 14 for engaging with a vibrating machine and a support structure as will be later described. A pair of springs 20 are provided between the plates 12, 14. In addition, a sprung mass is provided in the form of a weighted element in the form of metal fabricated weight assembly 30 which is associated with springs 21, 22 and 23, 24 by way of lugs 31 and 32 (see figure 2). As best seen in figure 2, lug 31 is sandwiched between springs 21 and 22 and lug 32 is sandwiched between springs 23 and 24.

Embodiment 1

Referring to figure 3, banana screen 100 is shown fitted with mountings 10. Four mountings 10 are used (two are visible in figure 3), each mounting 10 being located at a comer region of the screen 100.

Referring to figure 4, when installed for use, screen 100 is mounted on a support structure 200 which is typically inside a building and is formed from the floor of the building floor, joists, building foundations or other structural building elements. The four mountings 10 sit between the screen 100 and the support structure 199 to isolate the building from vibrations caused by operation of screen 100.

As will be later described, the physical characteristics of the springs and the weight assembly are selected so that in use, the weight assembly oscillates

approximately 180 degrees out of phase with the banana screen.

Before describing specific embodiments in further detail the principles underlying the operation of embodiments of the invention will be explained.

Operating Principles

The principals by which embodiments of the invention achieve significant further reduction of dynamic force into the support structure of a vibrating machine can be described by approximation as a two degree of freedom system as follows: • A spring support system is provided which has two parts, one passive and one reactive.

• The passive part of the system transmits force from the spring in phase with the main mass displacement (xl) (see figure 5 and figure 6)

• The reactive part of the system transmits force nominally equal to the passive part but essentially 180 degrees out of phase with the main mass displacement, ie equal but in the opposite direction.

• The sum of the passive and reactive parts essentially cancel each other out

• Residual unbalanced moments can be balanced by forming the system to be symmetrical or concentric.

For the reactive part of the system, the relationship between the forcing frequency, spring stiffness and the optimal size of the reactive mass can be described from the theory described below. The optimal size of the reactive mass being that which produces an equal and opposite force to the passive part of the system.

Referring to figure 5, a conventional passive spring support system is shown schematically in a single degree of freedom system with two passive springs of the same specification kl and kl . In figure 6, the system of figure 5 has been modified according to an embodiment of the invention to include a sprung mass ml which is mounted between two springs of stiffness ku and kl, (the system is then modelled as a two degree of freedom system)

Main Mass ml is being forced at frequency /. The response of ml to the input force is essentially mass controlled as it is being forced above the fundamental natural bounce mode and causes the mass ml to displace xl about a mean position.

Mode 1 fundamental bounce mode (suspension mode), ml and mr in phase:

Mode 2: reactive mass resonance, ml displacement out of phase with mr displacement

It is required that the forcing frequency / sufficiently above mode 2 for mr to be nominally 180 degrees out of phase with ml. Under these conditions the solution can be simplified to static equilibrium. For ease of formulation let the combined stiffness of ku and kl = kl, (this is not required for the system to function however it is more convenient in terms of passive spring replacement if the reactive system has the same total static stiffness. ie 1 //cl = 1 /ku + 1 /kl

Approximation to calculate the natural frequency of mode 1 and mode 2:

Assumption mr « ml (typical reactive mass is less than 10% of main mass) Mode 1 wn = (2/el/ml)^A.5 rad/s . (1)

Mode 2 wn = (( ku + kl)/mr)^A.5 rad/s . (2)

Force Balance:

We want FI = Fr so that the force from the passive spring is approximately equal (and opposite) to the lower reactive system spring. So want to target mr such that: klxl = klxr . (3)

for the forcing frequency /.

Consider the forces acting on the reactive mass mr. The total of the spring forces acting on the reactive mass are balanced by the acceleration of the reactive mass (a).

mr * a = kl xr + ku(xr + x 1) . (4)

Now express acceleration of reactive mass a in terms of reactive mass displacement xr and frequency / (Hz):

a = xr * ( 2p/)^L2 ... . (5)

Substitute for a in eqn 4:

Subs xl /xr from eqn 3:

For the special case where l//cl = l//cu + l//cZ

kl = ku = 2/cl

/cl = /cZ/2

4 kl

mr =

(2 p/)²

mr (7)

(p/)²

or

2fcl

r .(8)

(p/)²

It is important to note that the amplitude of the main mass does not affect the size of the optimal reactive mass and that the optimal reactive mass is a function only of the stiffness of the springs and forcing frequency.

It is also important to note that although the above derivation is shown for vertical vibration of the main mass, the theory still holds when the orientation of the single and two degree of freedom system is orientated in the direction of the main mass vibration. Figure 7 shows a sample frequency response function of the force transmitted or reacted by the support structure for a traditional passive spring isolation system. The input force to the main mass has been made as a function of the square of the frequency to simulate the force input from counter rotating eccentric mass shakers, typically used for many screening processes. At frequencies above the suspension modes this results in a constant displacement of the main mass independent of frequency and hence a constant force transmission into the support structure, nominally 6kN in this example.

Figure 8 shows the same plot for the same model and load conditions but with the system according to the principles of operation of the invention with a target maximum isolation frequency of l5.5Hz, ie the pre-determined or intended running speed of the main vibrating mass (or screen or feeder). Here it is clear that the net transmitted force into the support structure is zero at l5.5Hz and remains much reduced from the passive system over quite a large range of forcing frequencies, (eg in this example the system provides between 90% and 100% reduction in transmitted force to the support structure for forcing frequencies over the range l5Hz to l6Hz). It is important to note that the relative insensitivity of the system means that accurate or fine tuning is generally not required unless force transmission reductions of above say 90% are required. Note also that the percentage force reductions quoted here are the additional reductions over and above that provided by the passive system.

Worked Example 1

Referring now again to figures 3 and 4, the specifics of a design for the mountings 10 for banana screen 100 have been determined according to the theory set out above.

The screen 100 has a weight of 7.5Tonnes, vibrated at a nominal lOmm peak to peak stroke at l5.5Hz in the vertical plane (this requires a nominal 356kN eccentric mass shaker at 15.5HZ). The minimum allowable suspension mode of the screen was 2.0Hz, lower suspension modes tends to be impractical.

The calculated optimal reactive mass and one set of possible spring constants is shown in the table 1 :

Table 1 Nominal Spring constants

From equation 6 optimal Reactive Mass 508kg,

(From equation 1 mode 1 approximately 2.0lHz and from equation 2 mode 2 approximately 11.8Hz)

Under these conditions the transmitted force range would be reduced from nominally l8kN pk to pk at l5.5Hz to zero in theory. This system was analyzed using the Finite Element Method. This system produced zero vertical force transmission at l5.5Hz

Embodiment 2

Referring to figure 9 and 10, a second embodiment of the invention is shown. This embodiment differs from embodiment 1 in that the reactive sprung components are mounted at an angle of 45 degrees to the vertical plane. This configuration may be more practical when the main mass is being vibrated at 45 degrees to the vertical. The concept here is to align the reactive part of the invention with the direction of main mass vibration. The concept is therefore not limited to 45 degrees and could be arranged at any angle to the vertical depending on the direction of main mass vibration.

It should be noted that reduction of the vertical component of the transmitted force is usually of highest importance due to the response of the support structure, ie lateral vibration transmission does not usually cause the fatigue and human response issues noted above.

In figures 9 and 10, like reference numerals to figures 1 and 2 are used in the series 200 onwards. Passive springs 220 are provided at either end of the mounting. Weight assembly 230 is captured between pairs of springs 221, 222 and 223, 224. As best seen in figure 10, part of weight assembly 230 extends inside springs 221, 222, 223 & 224 to save space. Worked Example 2

The specifics of a design for the mountings 200 for banana screen 100 have been determined according to the theory set out above.

The screen has a weight of 7.5Tonnes, vibrated at a nominal l5mm peak to peak stroke at l8Hz and 45deg to the horizontal, (this requires a nominal 670kN eccentric mass shaker). The minimum allowable suspension mode of the screen was 2.0Hz, lower suspension modes tend to be impractical. The optimal reactive mass and one set of possible spring constants is shown in the table 2

Table 2 Nominal Spring constants

From equation 6 optimal Reactive Mass 376kg.

(From equation 1 mode 1 approximately 2.0lHz and from equation 2 mode 2 approximately l2.8Hz)

Under these conditions the transmitted force range would be reduced from 18kN pk to pk at l8Hz to zero in theory. This system was analyzed using the Finite Element Method. Referring to figure 11, this system produced zero vertical force transmission at l8Hz and reduced the horizontal force component from l2kN to 3kN. (Note that the coil springs used in the Finite Element Model matched the axial stiffnesses in Table 2 but were only within 10% of the target lateral stiffnesses shown in Table 2, hence the horizontal force was not reduced to zero at l8Hz).

Embodiment 3

Referring now to figure 12, a further embodiment of a mounting 300 is shown. This embodiment differs from that shown in figure 1 and 2 in that the reactive sprung mass in the form of weight assembly 330 and springs 321, 322 are located co-axially inside passive spring 320. This embodiment saves space and may reduce bending moments applied to the upper and lower plates 12, 14.

Embodiment 4

Referring now to figure 13, another embodiment is shown in which a vibrating machine 100 is fitted with four mountings 400. Two of mountings 400 are visible in the foreground and share the same sprung mass which is provided in the form of a frame.

In another variation, the frame could extend around the base of the machine 100 and act as a shared sprung mass for all four mountings 400. These variations may be preferable in some circumstances to versions in which each mounting has its own sprung mass.

It can be seen that embodiments of the invention have at least one of the following advantages:

• reduced dynamic force transmission significantly over conventional mass controlled (passive) spring suspension systems.

• The total of the transmitted dynamic force can be optimized so that the reactive part of the system completely cancels the dynamic force transmitted by the passive part, ie essentially zero net transmitted dynamic force as a result of the sum of the passive and reactive parts of the isolation system.

• The system can be used on any vibrating machine, eg vibrating screens and feeders), in order to reduce (and essentially eliminate) transmitted dynamic force at the driving frequency of the machine.

Any reference to prior art contained herein is not to be taken as an admission that the information is common general knowledge, unless otherwise indicated.

Finally, it is to be appreciated that various alterations or additions may be made to the parts previously described without departing from the spirit or ambit of the present invention.

Claims

CLAIMS:

1. A mounting for a vibrating machine which vibrates at a pre-determined

frequency, the mounting including:

a sprung mass which is arranged to oscillate out of phase with the vibrating machine at about the pre-determined frequency.

2. A mounting according to claim 1 wherein the sprung mass is arranged to oscillate approximately 180 degrees out of phase with the vibrating machine.

3. A mounting according to either of claim 1 or claim 2 wherein the sprung mass includes at least one weighted element which is associated with at least one spring.

4. A mounting according to claim 3 wherein the at least one spring is a coil spring.

5. A mounting according to claim 4 wherein the spring is mounted in use at an angle to the vertical direction.

6. A mounting according to claim 4 wherein at least a portion of the weighted element is disposed within the coil spring.

7. A mounting according to any preceding claim further including first and

second engagement portions for engaging with a vibrating machine and a support structure.

8. A vibrating machine including a mounting according to any one of claims 1 to 7.

9. A vibrating machine according to claim 8 including at least two mountings according to any one of claims 1 to 7.

10. A vibrating machine according to claim 7 wherein the at least two mountings share the same sprung mass.

11. A method of at least partially isolating a support structure from vibrations caused by a vibrating machine which is arranged to vibrate at a pre-determined frequency including the step of:

installing at least one sprung mass between the support structure and the vibrating machine, wherein the sprung mass is arranged to oscillate out of phase with the vibrating machine at about the pre-determined frequency.