FR2650352A1 - - Google Patents

Abstract

This spindle assembly is characterized in that it comprises a spindle 32 and a radial load bearing 33 which is unique and has a large bearing width intended to support the spindle 32 and in that in this radial load bearing 33 , a plurality of throttle nozzles 37 arranged in a circumferential manner and serving to introduce a compressed gas into a clearance between the spindle 32 and the bearing, is arranged near both ends thereof. at radial load 33.

Description

i LTA externally pressurized gas bearing spindle assembly

  The present invention relates to spindle bearing assemblies with

  externally pressurized gas in which the spindle is

  range, without contact with the bearing, thanks to the introduction of a gas

  compressed in a bearing clearance between the spindle and the bearing.

  An externally pressurized gas bearing can support a non-contacting pin between them, by introducing a compressed gas via throttling nozzles, into a set

  narrow bearing which is formed between the spindle and the bearing.

  Since this type of bearing is advantageous by comparison

  at any other level due to a small loss by friction.

  It is suitable for use as a bearing for a high speed spindle, for example a grinding machine, a drilling machine for drilling small diameter, machines for machine a fragile material or machinery

like.

  Figure 5 shows a conventional spindle assembly using externally pressurized gas bearings. This spindle assembly comprises a spindle 2 mounted in a housing 1. This spindle 2 is supported, by being out of contact with the bearing surfaces, by four bearings with radial load 5 to 6 and by two thrust bearings 8 and 9. facing each other and between which is sandwiched a collar 7

of pin 2.

  In bearings with radial load 3 to 6 and thrust bearings 8 and 9, throttling nozzles 11 and 12 are respectively provided.

  for introducing gas and communicating with a feed passage

  10 Due to the introduction of a compressed gas into the bearing play, the spindle 2 can be supported without contact in the radial direction, as well as in the direction of the loads. axial. Each bearing clearance communicates with

  the atmosphere via exhaust passages 13 to 16.

  In the embodiment shown, pin 2 is

  born directly by a rotor 17a and a stator 17b of a motor 17 mounted at the rear end of the spindle. It may be indirectly driven, via a belt or coupling, from another drive source provided outside the casing. If spindle 2 does not rotate well due to an exciter force due to a spindle defect or vibrations transmitted from the drive source, this can cause errors in the form of the machined surface, as well as a breakage of the tool. If the amplitude of such a runout increases further, the spindle may come into contact with the bearing surface, thereby damaging the bearing surface.

  This is why it has so far been considered important.

  as long as an externally pressurized gas bearing has a

  as high as possible in order to increase its natural frequency

  and thus make it possible to use it at a lower speed than

  than this natural frequency, and in order to keep the

runout of the spindle.

  In practice, in order to increase rigidity, a large number of bearings having a ratio have been disposed over a limited length of time.

  rigidity / bearing length which is important, as is

  Figure 5. If sufficient rigidity is obtained with a smaller number of bearings, the bearings are sometimes omitted in the

center of the spindle.

  FIG. 6 shows the relationship between the K / L ratio and the L / D ratio for a spindle assembly having a diameter of

  30 mm spindle, has eight nozzles, arranged in two rows and

  having a nozzle hole diameter of 0.3 mm, and provides a bearing clearance of 12 micrometers and a gas supply pressure (gauge pressure) of 5 kgf / cm2, K being the stiffness, L the length of the bearing

  at radial load and D the diameter of the spindle.

  As shown in Figure 6, the lower the L / D ratio, the higher the K / L ratio. That's why we have up to

  present used for a single pin several levels having a ratio

L / D port from 0.5 to 2.

  However, with such a spindle assembly of the prior art

  it is only within a certain limit that it is possible to increase

  Maintain the natural frequency by increasing rigidity. If this spindle assembly is used at a rotation speed greater than its frequency

  naturally, it is necessary to perform an in situ balancing

so much a good precision.

  If a tool or similar element is attached to the spindle head, it is necessary to restore in-situ balancing whenever there is a defect in balancing, albeit small, when mounting or dismounting the spindle. 'tool. Otherwise, it would be difficult to do

  turn the spindle at a speed higher than the natural frequency.

  Such in situ balancing work requires expert hands and a lot of work. This raises costs and hinders productivity. Moreover, it is inefficient that a user of the set-pin

  adjusts balancing in situ at the workplace.

  For the reasons described above, it has been very difficult to use a spindle assembly of the prior art at a speed

high rotation.

  This is why the present invention aims to provide a

  spindle assembly, comprising a gas bearing placed externally under pres-

  sion, which can prevent the spindle from turning round and

does not require in situ balancing.

  For this purpose, the subject of the invention is a spindle assembly, characterized in that it comprises a spindle and a bearing with radial load which is unique and has a large bearing width intended to support the spindle and in that, in this radial load bearing, there are provided, near both ends, a plurality of circumferentially disposed throttling nozzles for introducing a compressed gas into an existing gap between the brooch

and the bearing with radial load.

  The single radial bearing is an externally pressurized gas bearing and, because of the presence of the throttle nozzles for the introduction of this gas, the spindle can be supported without touching the bearing surface. This remarkably increases the damping coefficient and prevents the spindle from

  not turn well during a high speed rotation.

  Thanks to the invention, since, on the basis of the calculation results, due account is taken of the damping coefficient which has hitherto been ignored, a radial load bearing which is unique and large bearing width and in which are arranged at one and the other of its two axial ends, a plurality of throttling nozzles'ervant to introduce gas and arranged in a circumferential manner. This increases in a remarkable way the

  ratio damping coefficient / rigidity of the bearing. This result

  itself the precision of the out-of-roundness of the spindle during a

high speed rotation.

  Moreover, if there is no need for false precision,

  round of the spindle which is greater than that of the whole of the

  than earlier, one can lower the accuracy of the adjustment of the spindle balancing defect in order to reduce operations such as

in situ balancing.

  Since only one radial bearing is provided, it becomes possible to reduce the number of parts and the number of very small holes, such as the throttle nozzles used to

  introduction of the gas, to simplify the arrangement of the feed-

  exhaust and reduce the number of manufacturing operations.

  By reducing the number of throttle nozzles provided for the introduction

  of the gas used to support the spindle, the part formed by the

  bearing consumes a smaller amount of compressed gas.

  Other features and advantages of the invention disclosed

  the following description, as non-limiting examples

  and with reference to the appended drawings, in which: FIG. 1 is a vertical sectional view of a first embodiment of a spindle assembly according to the invention, FIG. 2 is a vertical sectional view of a second embodiment, FIG. 3 is a vertical sectional view of a third embodiment,

  FIG. 4 is a diagram for explaining a vibrational model.

tory

  FIG. 5 is a vertical sectional view of an assembly

  pin, externally pressurized gas bearing, which belongs to the prior art, FIG. 6 is a graph showing the relationship between the length, diameter and rigidity of the radial load bearing and FIG. 7 is a graph representing the relationship between the damping coefficient per unit length of the bearing and

  the length and diameter of this bearing.

  Embodiments of 1-a present invention will be

  described with reference to Figures 1 to 4 and 7.

  By studying the influence of the stiffness and the damping coefficient on the displacement produced by an excitatory force due to a defect in a vibratory model, the inventors have found that it would be better to put the weight on the coefficient. damping only on rigidity. This is why the spindle assembly according to the invention comprises a portion constituting the bearing which

  has a high ratio of damping coefficient to stiffness.

  Consider the vibratory motion occurring in a spindle system of a spindle assembly, which system is simulated with the aid of a vibratory model which is represented in FIG. 4 and which comprises a point mass 21 of mass m, a spring 22 of constant

  elastic band K and a shock absorber 25 of damping coefficient

  The point mass 21 corresponds to the mass of the spindle or to its moment of inertia and the spring 22 and the vibration damper 23

  correspond respectively to the stiffness k and the damping coefficient

  parts of the landing. If an exciter force of amplitude F and angular frequency cJ acts on the point mass 21, the equation of motion of this point mass is expressed as follows: mx + C; + kx = F cos u t (1) x being the displacement of the point mass and t the time. By solving

  equation (1) so as to obtain the amplitude XF of the random mass

  it comes: XF = F / * (k - m.02) 2 + c2>) 2 (2) If the exciter force> is produced due to a spindle defect, the frequency this exciter force is equal to the number of turns of the spindle per unit of time. Its amplitude is given by: F = u, 02

  where u is the value of the spindle defect. As a result, the

  plitude Xu is given by: Xu = uu) 2 / (k - m,) 2) 2 + c2uJ2 (3) Since the natural frequency cO of the vibratory model is Vk / m, the amplitude of the outflow due to a defect of equilibrium is given by the following equation (4), obtained by introducing J = k- / m in equation (3): Xu = u / [c2 = u7 / J. 7 (4) Thus, if the natural frequency is equal to the speed of rotation, the outflow due to the defect balancing with the rigidity

of the part constituting the landing.

  On the other hand, if West is higher than the natural frequency VkIm, because k <m J2, the larger the value of k, the lower the value of (k - m & J2) 2, provided that the values of m and "are

  constant. In this way, the denominators of equations (2) and (3)

  nue. In other words, even if the exciter force has a frequency greater than the natural frequency, the value of k is

  large, the greater the degree of runout is important.

  As can be seen from the above description, if the

  rotational speed during actual use is greater than the

  natural frequency of the spindle-bearing system, it is impossible to prevent

  expensive spindle not to turn well round with the conventional structure in which the portion constituting the bearing has a rigidity

increased.

  With regard to the damping coefficient of the bearing, it follows from equation (4) that the higher the damping coefficient c is

  large, the degree of runout due to the lack of balancing is low.

  The same is true for the denominators of equations (2) and (3), since the damping coefficient c is multiplied by the excitation frequency cJ. This means that the spindle runout can be limited in a more efficient way by reducing rigidity, while increasing the damping coefficient, while the spindle turns

  a speed higher than the natural frequency.

  FIG. 7 represents the relationship between the C / L ratio (damping coefficient per unit length) and the L / D ratio, assuming that the dimensions of the bearing are the same as in FIG. this figure 7 that the C / L ratio increases with the L / D ratio. The rate of increase decreases, however, once the L / D ratio has become greater than. 3. In accordance with the invention, the above results are applied to a pin having a finite length and a radial load bearing is used which is unique and has a large bearing width. In addition, a plurality of throttle nozzles for the introduction of gas are arranged in a circumferential manner, along two rows located in the vicinity of

  one and the other of the two ends of the bearing.

  FIG. 1 represents a first embodiment, according to the invention, in which a pin 32, mounted in a housing 31, is

  supported without contact by a radial load bearing 33 to gas extinguished

  pressure-bearing, sleeve-shaped and single, and by two thrust bearings 35 and 36 to gas placed externally under pressure, which are arranged face to face and between which is sandwiched a

flange 34 of pin 32.

  A plurality of throttle nozzles 37 for introducing gas are circumferentially disposed in two rows in proximity to each other at both ends of the bearing.

  radial load 33. Compressed gas is introduced through a feed passage

  compressed gas 39, formed in the housing 31, and by the nozzles

  throttling 37 and throttling nozzles 38 serving as the introduc-

  In the thrust bearings 35 and 36 h, the gas is placed under pressure in order to support the pin 32 without contact. In the embodiment shown, the pin 32 is driven directly by a rotor 40a and a stator 40b of a motor 40 and

  the bearing clearance communicates with the atmosphere via pas-

exhaust tips 41 to 43.

  Since the radial load bearing 33 is in the form of a single sleeve having a large bearing width

  and that it is located close to both of its two extremes.

  As throttle nozzles 37 are used for the introduction of gas, the ratio of damping coefficient to rigidity of the bearing increases remarkably. As a result, the damping coefficient is much higher than in the conventional spindle assembly. This prevents the spindle from turning round during a rotation

at high speed.

  Since the radial load bearing 33 is in the form of a single sleeve, this bearing sleeve must be longer than conventional bearings. This is why it is sometimes very difficult to provide one of the thrust bearings at one end of the radial load bearing 33 so that it is in one piece with this bearing radial load and drill holes. driving holes for the throttle nozzles in an axial direction, as is the case for

  the example of the prior art which is shown in FIG.

  In order to avoid this situation in the first mode of

  of FIG. 1, the thrust bearing 36 is distinct from the h bearing

  33. However, although such a structure is easy to manufacture, the radial load bearing 33 is shorter by a length corresponding to the length of the thrust bearings. Moreover, being

  given that the mass of a thrust plate and the rotor of the motor

  linked by an even greater length relative to the end of the bearing b radial load 33, it becomes difficult to limit the runout

pin in conic mode.

  This problem is solved in the second embodiment, shown in Figure 2, wherein one of the abutment bearings 36a is provided at one end of the radial load bearing 33 so as to be in one piece with thereof, and in which driving holes 38a for the throttle nozzles 38 for the introduction of gas are

  formed in this thrust bearing 36a in a radial direction.

  FIG. 3 represents the third embodiment in which an air turbine is used as a training source for the

  pin 32, so that the assembly can be used for a spraying

  electrostatic precipitator or similar apparatus. The same elements as those of the first embodiment shown in FIG. 1 are designated by

  the same references and their description are omitted.

  In FIG. 3, the pin 32 is supported by a radially loaded, sheath-shaped, externally pressurized gas bearing 33 mounted in a housing 31, and on its rear end is provided, made of a piece with it, a large-diameter collar 51 on the outer peripheral surface of which are provided

  turbine blades 52 arranged in a circumferential manner and intended

  born to receive compressed air. The spindle 32 is rotated at a high speed by blowing, on these turbine blades 52, air introduced by an air supply connector 53 formed in the

housing 31.

  Compressed gas is supplied from the air supply connector 53 to the throttle nozzles 38 for supplying gas and provided on either side of the large diameter flange 51 so as to support pin 32 without contact, in the direction of axial thrust, with the compressed gas which acts on one and the other

sides of the flange 51.

  Since an air turbine is used as a drive source for pin 32 in the third embodiment, the size of the complete assembly can be reduced in comparison with the

case where an engine is used.

  In the embodiment shown in FIG. 3, the gas is sent to the air turbine and to the bearing via the single supply connection 53 which is common to them. However, we can foresee

separate connections.

  On the other hand, in the case of FIG.

  radial, sleeve-shaped and single, is mounted in the housing 31.

  However, the sleeve 33 can be divided into several cylindrical parts.

  which can be introduced into the casing with their

  my adjacent abutting one another by force-fitting or retracting interlocking to form a load bearing

radial which is unique.

1u

Claims (1)

CLAIM   Spindle assembly, characterized in that it comprises a spindle (32) and a radial load bearing (33) which is unique and has a large bearing width for supporting the spindle (32) and in that bearing with radial load (33), it is arranged, close to   both ends, several nozzles of   ment (37) arranged in a circumferential manner and serving to introduce   a compressed gas in an existing gap between the spindle (32) and the bearing with radial load (33).