Dependency of bearing noise properties on surfaces lubrication/Pavirsitepimo itaka guoliu keliamam triuksmui.
Augustaitis, V.K. ; Bucinskas, V.
1. Introduction
Noise of technical equipment usually assumed to be negative effect.
Noise affects personnel and brings its negative effect of operating
machinery and equipment. Technical equipment with rotational movement
has different types of bearings, which are known as technical noise
generators.
Modern materials in bearings create some specific problems with
noise intensity and this paper touches some aspects of noise generation
in steel and ceramic bearings. To make research of such task it is
necessary to built a model of noise generation and to find parameters of
noise generation mechanism.
Noise of bearings is created by vibrating solid bodies, which is
transmitting further to air. It is known [1] that sound filters solid
body vibrations due to its own physical properties of wave media.
Vibration of bearing components has different nature, but for
operable ones the main part of noise in 2 5 Hz area belongs to friction
noise [2-8].
2. Aim and formulation of research
This research is intended to find dependency between bearing
material hardness, lubrication of bearing and vibration characteristics
as well as acoustic noise.
The research presented consists from several parts. The first step
is to accept model of friction noise generation in a bearing, based on
tribologic properties of bearing surfaces and other sources of vibration
are neglected. This statement is used only in case of experimental
setup; in real machinery it can't be used directly. Noise
generation hypothesis is presented in Fig. 1. There are two cases of
vibration generation - when two surfaces are sliding (as shown in Fig.
1, a), and direction of surfaces movement is opposite in tangential
direction or when the surfaces are rolling and direction of surfaces
movement is coincident in tangential direction.
In this paper only rolling mode is assumed.
[FIGURE 1 OMITTED]
Surface of contact area are random peaked and only statistically
evaluated surface pitch [R.sub.t] and average roughness [R.sub.z] are
available to evaluate in this model. So, piece of some surface
projection, showed on Fig. 1, b can be accessed directly by another
surface to support force, or support surface is modified by lubricant
film, which is represented by two positions of oil surface as "thin
oil" and "thick oil".
These surfaces are interacting with peaks, "sticking out"
from lubricant film with their own [R.sub.z] and [R.sub.t].
This creates different conditions in excitation of vibrations when
the surfaces are moving in coinciding tangentiall directions (rolling)
or in opposite tangentiall directions (sliding).
Then for a single surface cross section is possible to write the
following dependencies
[f.sub.sl] = [[omega].sub.1][R.sub.1] + [[omega].sub.2][R.sub.2] /
[R.sub.t] (1)
where [f.sub.sl] is average statistical frequency in sliding;
[[omega].sub.1], [[omega].sub.2] are angular speed of surfaces;
[R.sub.1], [R.sub.2] are radii of rotation for these surfaces; [R.sub.t]
is equivalent pitch for a roughness.
In case of rolling average theoretical frequency can be expressed
as
[f.sub.sl] = [[omega].sub.1][R.sub.1] / [R.sub.t] =
[[omega].sub.2][R.sub.2] / [R.sub.t] (2)
Dependencies (1) and (2) are very rough estimation of kinematical
excitation in contact of friction noise, which separates modes of
movement of contact surfaces cross-sections. Real frequencies are
generated from real surface relief, so the number of frequencies in
huge, but center values follows this model [8 - 9].
This model has also more modifications, but the main ideas can be
accepted also in slightly simplified form [10 - 11].
3. Modelling of contact
In order to look on contact area behavior, FEM model of one side
bearing contact, which allows to evaluate distribution of stresses in
contact and shows active area in noise generation was created.
As the basis for this model a bearing No. 208 was taken, but the
model can be used in other applications. Contact model was created using
SolidWorks software and FEM analysis was performed on CosmosWorks 2010.
FEM model (Fig. 2) was created using one fragment from bearing No. 6208.
The fragment of outer ring was taken with angular size, which
corresponds to another neighboring ball contact places. For contact
modeling here was chosen contact type "surface -surface",
while in the model little gap was created and bonding place was unknown.
Because of ball shape, restrictions in both horizontal axis directions
were fixed by creation axis, only vertical ball movement was allowed.
Outer ring fragment was fixed in all 3 axes direction on the outer
surface. In order to run such model, special stabilizing spring was
used.
All elements were taken as 1st order tetrahedron, contact was
assumed to be Coloumb's. Load of the ball was applied in the area
of estimated contact area of other ring, size - 80% of bearing radial
load, according the shape of free body diagram.
[FIGURE 2 OMITTED]
Solution of this task is presented in Fig. 4, where inner surface
of the bearing and ball contact surface are shown. Configuration of
contact on ring running surface and ball are slightly shifted from the
axis of symmetry due to global friction in the contact.
[FIGURE 3 OMITTED]
The analysis of stresses and displacements shows slight change in
stresses in the area contact proves model of friction noise, because
great gradient in stresses would create another scenario in vibration
excitation.
4. Methodology of research
This experimental research was performed in Braunschweig technical
university (Germany) for ceramic bearings and in Hannover Leibnitz
technical university (Germany) was made extensive steel bearing
research.
Initially contact area of the bearing was measured for roughness.
Profile of contact, which is shown in Fig. 4, was evaluated for the main
parameters and statistically proved values or [R.sub.t] and [R.sub.z]
were defined. These values were basic in the definition of desired noise
frequency range. It is necessary to take into account, that
statistically proved data on surface of contact differs within certain
tolerance and phase of vibration from touching roughness profile is also
unknown. The frequency of resulting vibration from outer ring of the
bearing was recorded and further analyzed. Spectrum of vibration
accelerations is a result of such research and the main information
supplier.
In case of nonlubricated bearing, ceramic bearing with very similar
surface parameters was tested in another test rig. This test rig was
intended to record acoustic pressure. Duration of the test in this test
rig was very short in order to avoid heating. Partial axial load was
applied, because installation of ceramic bearing was not tightening
enough the outer ring and there was internal gap. Loads on ceramic
bearing and steel bearing were made proportional to their maximum load
(80% of it) and axial load was assumed not influential.
[FIGURE 4 OMITTED]
Bearings (steel and ceramic) were built into special setups and
rotated with corresponding load and rotational speed. Because of
different design of setup and bearing size, rotational frequency was
taken so, that linear rotational velocity of the rolling bodies should
be the same - 600 rpm for steel bearing and 1000 rpm for ceramic
bearings.
5. Results
Results of performed research are shown below graphically. Fig. 5
shows steel bearing vibration acceleration signal in time. This sample
was lubricated by industrial lubricant 9.
[FIGURE 5 OMITTED]
Vibration signal output from the steel bearing was recorded; signal
discretion is 0.00002 s, which corresponds 50 kHz of sampling rate. From
such signal frequency spectrum was calculated using FFT. As it is seen
from vibrational spectrum (Fig. 6), the main frequency peak is about
3000 Hz and 7000 Hz as the second harmonics and low V subharmonic on
1500 Hz. Higher frequency range in spectrum has low accuracy because of
sampling rate. These frequencies do not fit to ball pass or inner
bearing ring revolution rate.
[FIGURE 6 OMITTED]
In case of lubricant t68 (Fig. 7), the spectrum was shifted to
higher side and the values of vibration amplitudes are significantly
higher, what means lower film thickness, as stated in initial model. The
main frequency range shifted to 2600 Hz, correspondingly the second
harmonics to 7200 Hz. During these tests lubricant film thickness was
not measured, but the frequency change for more than 10% shows that the
number of sticking roughness peaks correspondingly increased about the
same number.
[FIGURE 7 OMITTED]
In case of ceramic bearing analysis (the measuring of outer ring
vibration was made using laser sensor) another software was used. As it
is possible to see in processed spectrum, ceramic bearings create much
higher values of vibrations (Fig. 8, a) and correspondingly higher
acoustic pressure. Higher level of vibration amplitudes is caused due to
higher hardness of bearing material. Absence of lubricant in contact
enables to touch much more peaks in contact area and energy of
elementary impact between roughness peaks makes much higher.
[FIGURE 8 OMITTED]
Result of such investigation evidently show that more accurate
model for vibration frequency evaluation is necessary, because the
presence of prescribed frequency in the range of 3000 Hz and below has
coinciding sub harmonics and higher harmonics.
6. Conclusions
After analysis of different types of bearings, it is possible to
state, that material of a bearing and lubrication influence on bearing
vibrations and bearing noise is significant. Harder materials for
bearings, such as ceramics, without the use of lubricant increase noise
values and vibration energy and only soft media between ceramic parts
can make effect of vibration energy decrease. Correspondingly, in case
of soft media, like thick oil film, increase resistance moment in the
bearing and cause higher heating. In case of oil film breakage, dry
friction due to dry surface increase energy of vibration (expressed in
acceleration amplitude shift to higher frequency range) and higher
values of local stresses (contact in surface peaks) develop heat and
surface damage. As particular conclusion it is possible to state:
1. Difference in dynamic characteristics of bearing vibration in
the same surface interaction velocity with bearings from different
material is defined by surface hardness and particularly rolling/sliding
values.
2. Lowering of bearing vibration and noise with different lubricant
properties are directly influenced by lubricant film efficient thickness
and surface "sticking roughness", load in the bearing can be
expressed in the terms of efficient film thickness;
3. Rolling and momentum sliding mode of rolling elements inside
bearing is characterizing by vibration spectrum frequencies
significantly.
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Received January 21, 2011
Accepted May 30, 2011
V. K. Augustaitis, Vilnius Gediminas Technical University,
Basanaviciaus 28, 03224 Vilnius, Lithuania, E-mail:
vytautas.augustaitis@vgtu. lt
V. Bucinskas, Vilnius Gediminas Technical University, Basanaviciaus
28, 03224 Vilnius, Lithuania, E-mail: vytautas.bucinskas@vgtu.lt