1. INTRODUCTION Market demands of new machine tools, oriented towards productivity improvements, can be satisfied by innovative designs capable of working simultaneously at higher speed and power ranges. A typical currently available top-level spindle can provide 40 kW at 40–50 k rpm, but some manufacturing sectors ask for100 k rpm spindles capable of working at the same (or higher) power. At such high speeds even the most performing ball bearings are not suitable for their limitations due to noise and vibrations phenomena, wear of balls and cage, frequent uneconomical maintenance interventions (and consequent long lay-offs from production),local structural deformations induced by temperature rising due to high friction levels. Even conventional hydrostatic bearings do not allow the reaching of such high performances mainly because of the effects of temperature rising, resulting from energy dissipations generated by fluid viscosity. 2. TYPES OF BEARING 2.1 ROLLING CONTACT BEARING A rolling contact bearing consist of four parts – inner and outer races, a rolling element like ball, roller or needle and a cage which hold the rolling elements together and spaces them evenly around the periphery. Depending upon the type of rolling element, bearing are classified as ball
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1. INTRODUCTION
Market demands of new machine tools, oriented towards productivity
improvements, can be satisfied by innovative designs capable of working
simultaneously at higher speed and power ranges. A typical currently available top-
level spindle can provide 40 kW at 40–50 k rpm, but some manufacturing sectors ask
for100 k rpm spindles capable of working at the same (or higher) power. At such high
speeds even the most performing ball bearings are not suitable for their limitations
due to noise and vibrations phenomena, wear of balls and cage, frequent
uneconomical maintenance interventions (and consequent long lay-offs from
production),local structural deformations induced by temperature rising due to high
friction levels. Even conventional hydrostatic bearings do not allow the reaching of
such high performances mainly because of the effects of temperature rising, resulting
from energy dissipations generated by fluid viscosity.
2. TYPES OF BEARING
2.1 ROLLING CONTACT BEARING
A rolling contact bearing consist of four parts – inner and outer races, a rolling
element like ball, roller or needle and a cage which hold the rolling elements together and
spaces them evenly around the periphery. Depending upon the type of rolling element,
bearing are classified as ball bearing, cylindrical roller bearing, taper roller bearing and
needle bearing. Depending upon the direction of the load, the bearing are classified as
roller bearing and thrust bearing. There is however, no clear distinction between this two
groups. Certain types of radial bearing also take thrust load and some type of the thrust
bearing also take radial load.
Rolling contact bearing having a low starting friction as comparing to the sliding
contact bearing so it is also called antifriction bearing. But it is also more noisy , and low
resistance to shock loading
(1)
2.2 HYDRODYNAMIC BEARING
In a hydrodynamic lubricated bearing there is a thick film between the journal and
the bearing. A little consideration will show that when the bearing is supplied with
sufficient lubricant, a pressure is buildup in clearance space when journal is rotating about
its axis that eceentric with the bearing axis. So the load can be supported by this pressure
without any actual contact between the journal and bearing. Load carrying ability of this
bearing arise simply because of a viscous fluid resist being pushed around.
2.3 HYDROSTATIC BEARING
Hydrostatics bearing is defined as a system of lubrication in which the load
supporting fluid film, separating the two surface, is created by an external source, like
pump, supplying sufficient fluid under the pressure. Since the lubricant is supply the
under pressure, this type of bearing called externally pressurized bearing. In this type of
bearing as the pump start high pressure fluid is admitted in the clearance space, forcing
the surface of bearing and journal to separate out compared to the hydrostatic bearing ,
hydrodynamic bearing are in simple construction, easy to maintain and lower initial as
well as maintenance cost. Hydrostatic bearing is costly it offer the advantage :
(1) High load carrying capacity even low speed,
(2) No starting friction and,
(3) No rubbing action at any operating speed or load.
This types of bearing are used on the vertical turbo-generators, centrifuges and ball mills
2.4 FLOATING BEARING
In the full-floating journal bearing we located a floating sleeve between the
journal and bearing surface. the floating sleeve can be operated for a wide range speed for
a given shaft . In high speed element like turbines and compressor sometime we getting
overheating of the journal bearing which is also encounter in high speed operation. So for
this we increasing the clearance but increasing of clearance get into decreasing load
carrying capacity. But the full-floating journal bearing means increasing oil flow without
increasing clearance. A floating sleeve provides the two channels through which the oil
may flow.
(2)
Fig:1 Difference between conventional bearing and full-floating bearing
2.5 SQUEEZE FILM JOURNAL BEARING
There is a certain case, the bearings are oscillate or rotate so slowly squeeze film
bearing gives a satisfactory result. If load is uniform or varying in magnitude while acting
in constant direction, this becomes a thin film or possibly a zero film problem. But if load
reverse its direction, the squeeze film may developed sufficient capacity to carry the
dynamic load without between the journal and bearing. such bearing are known as the
squeeze film journal bearing.
(3)
3. THE COAXIAL FLOATING SLEEVE
A feasible approach to high-speed / high-power machine tool consists in placing
free sleeve between the shaft and the housing, thus splitting the total relative speed in two
contributions Comparative tests on different coaxial ball bearing configurations have
proven this solution being not practicable because, depending on tested layout, either of
the two coaxial bearings is not driven into rotation. It follows that the speed ratio between
the two ball bearings has to be forced and controlled by external pneumatic/mechanical
devices.
On the contrary, the proposed coaxial hydrostatic configuration keeps all the
advantages of an hydrostatic bearing: contactless capability, no wear phenomena, no
maintenance required, the viscosity of the lifting fluid being the only source of friction.
Moreover, if properly dimensioned, this bearing configuration allows to divide the total
relative velocity into two almost identical contributions, whose ratio is self-controlled by
the balance of the viscous torques acting on the internal and external cylindrical surfaces
of the sleeve. This way, through the simplification of the technical difficulties related to
high speed design, the target of 100 krpm can be achieved with the current available
know how.
The insertion of the sleeve allows saving up to half of the friction power
generated by the same bearing without the bush. This happens because of the quadratic
relationship existing between the friction power in an hydrostatic clearance and relative
speed of its two opposite sliding surfaces; thus, once given the shaft speed, the reduction
of friction power is maximum when the speed of the floating sleeve becomes half the one
of the shaft. Of course, under the same conditions, friction powers decrease when
increasing the clearance height; nevertheless this height cannot become too high as this
solution would also increase the flow rate, and consequently the pumping power.
Friction cannot be further decreased by reducing the viscosity of the fluid, as the current
limit of the best industrial oils is about 4–5 cSt.
Damping performances are also improved, while stiffness characteristics are
expected to be reduced. This last shortcoming can be overcome by introducing automatic
bipartition valves on the feeding line of the pads of the external bearing, in order to
provide infinite stiffness on the external clearance: this way the total stiff-ness is the same
as the one of a single-clearance hydrostatic bearing. Anyway the realized test bench was
(4)
not provided with such devices, as in a real spindle the main stiffness problems originates
from the deformability of both shaft and cantilever tool head; eccentricities of the
bearings are negligible if compared to the for ementioned deformations.
Disadvantages and costs coming from the need of a feeding system (pumps,
filters, tanks, etc.) are expected to be widely compensated by economical advantages of
higher production rates, by avoiding to interrupt production for maintenance, and new
market opportunities due to high speed capability.
Previous experimental activities on a pneumostatic test bench proved the floating-
sleeve bearing to be capable of splitting the global relative speed gap into two
contributions of comparable sizes. These tests also pointed out that an accurate definition
of the clearance heights is essential for having the bearing working properly, that is
obtaining the desired speed ratio between the sleeve and the shaft. Therefore, when the
rotating speed increases up to some 100 krpm, the deformability of the rotating
components has to be accounted for, as elastic radial expansions modify the nominal
heights of the clearances by becoming comparable with the heights themselves
An incorrect design of the clearances or the neglecting of centrifugal deformations
may prevent from reaching the requested performances. The purpose of this paper is to
provide hydrostatic designers with a new bearing configuration suitable for very high
speed spindles, describing a method enabling them to correctly predict the behaviour and
the performances of the bearing with the best engineering precision: this requires
accounting for the radial expansions of the sleeve.
(5)
4. THE RIGID MODEL
The geometry of the hydrostatic bearing is sketched in Fig. 2: geometry of the
pads on the shaft and on the external housing is shown in Fig. 3. The model is based on
the following hypothesis:
(a) the viscosity of the lifting fluid does not depend on temperature,
(b) fluid viscosity has a constant value inside each of the two clearances,
(c) there is no slip of the fluid at the walls,
(d) the lifting fluid is uncompressible; corrections on this hypothesis are required when
using air or gas
(e) laminar flow: velocity profile across the clearances is linear and shear stressis
constant,
FIG 2. scheme of the bearing with floating element
FIG 3. CAD model of prototype bearing.
(6)
(f) local effects of curvature on fluid flows are neglected as the gap height is
negligiblewith respect to its radius,
(g) all eccentricities e are zero, so hydrodynamic phenomena are ignored.
In order to write the expression for the rotating speed 2 of the floating sleeve, the
viscous torques T1 and T2 acting on the opposite surface of the floating element are forced
to be equal, once the system is working in steady conditions. Torques can be expressed
by intergrating the viscous shear stress over the whole bearingfriction areas S:
S = NP.Af = NP.(AP –3/4AR)
being NP the number of pads of the bearing, Af the pad friction area, AP the area of a
single pad and AR the recess area
T1
=S1 R1.
1.dS,
T2
=S2 R2.
2.dS,
According to hypothesis (e), considering the expression of viscous shear stress , and then
writing the relative tangential velocity as .R, the shear stress becomes:
= ( µ) / h = (µ.R)/h
So the expressions of the viscous torques can be rewritten as follows:
T1= (µ1 (R1)2 S1 (1 - 2))/h1
T2= (µ2 ( R2)2 S2 2)/h2
Now, imposing T1 = T2, the angular speed ratio 1/2 can be written as a function of the