GB2235259A

GB2235259A - Spindle assembly with externally pressurised gas bearing

Info

Publication number: GB2235259A
Application number: GB9016234A
Authority: GB
Inventors: Shizuka Yamazaki; Yoshio Fujikawa; Haruhisa Harada; Koji Nakamura
Original assignee: NTN Corp; NTN Toyo Bearing Co Ltd
Current assignee: NTN Corp
Priority date: 1989-07-28
Filing date: 1990-07-24
Publication date: 1991-02-27
Anticipated expiration: 2010-07-24
Also published as: GB9016234D0; FR2650352A1; FR2650352B1; GB2235259B

Abstract

A spindle assembly has a spindle 32 supported by a single journal bearing 33 having a large bearing width. The journal bearing is formed near both ends thereof with a plurality of circumferentially arranged gas supply restrictor nozzles 37. The spindle is supported by supplying compressed gas to the bearing clearance around the spindle. This arrangement prevents the spindle from running out of true. A pair of externally pressurised gas thrust bearings 35, 36 may also be provided. <IMAGE>

Description

SPINDLE ASSEMBLY WITH EXTERNALLY PRESSURIZED GAS BEARING This invention relates to a spindle assembly with an externally pressurized gas bearing in which the spindle is supported without making contact with the bearing by supplying compressed gas into a bearing clearance between the spindle and the bearing.

An externally pressurized gas bearing can support a spindle with no contact between them by supplying compressed gas through restrictor nozzles into a narrow bearing clearance formed between the spindle and the bearing.

Because this type of bearing is advantageous over any other bearing in view of less frictional loss and longer life, it is suited for use as a bearing for a high-speed spindle of e.g. a grinder, a drill for forming a smalldiameter hole, machines for working a brittle material or the like.

Fig. 5 shows a conventional spindle assembly using externally pressurized gas bearings. It has a spindle 2 mounted in a housing 1. It is supported, out of contact with the bearing surfaces, by four journal bearings 3 - 6 and a pair of thrust bearings 8, 9 opposed to each other with a collar 7 of the spindle 2 sandwiched therebetween.

The journal bearings 3 - 6 and the thrust bearings 8 and 9 are formed with gas supply restrictor nozzles 11 and 12, respectively, communicating with a compressed air supply passage 10 formed in the housing 1. By supplying compressed gas into the bearing clearances, the spindle 2 can be supported in a non-contact manner in radial as well as thrust directions. Each bearing clearance opens to the atmosphere through exhaust passages 13 - 16.

In the illustrated embodiment, the spindle 2 is driven 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 by other driving source provided outside the housing through a belt or a coupling.

If the spindle 2 runs out of true by an exciting force due to imbalance of the spindle or by vibrations transmitted from the driving source, errors in the shape of the finished surface as well as breakage of the tool might be caused. If the amplitude of such runout further increases, the spindle might get into contact with the bearing surface, thus damaging it.

Therefore, it has heretofore been considered important that an externally pressurized gas bearing has as high a stiffness as possible in order to increase its natural frequency and thus to make it possible to use it at a lower speed than its natural frequency and in order to keep the runout of the spindle to minimum.

In practice, in order to increase the stiffness, a large number of bearings having a large stiffness/bearing length ratio were arranged within a limited length as shown in Fig. 5. If a sufficient stiffness is achieved with a smaller number of bearings, the bearings at the central portion of the spindle are sometimes omitted.

Fig. 6 shows the relationship between the K/L ratio and the L/D ratio with a spindle assembly having a shaft diameter of 30 mm, having eight nozzles arranged in two rows and having a nozzle hole diameter of 0.3 mm and having a bearing clearance of 12 microns and a gas supply pressure (gauge pressure) of 5 kgf/cm2, wherein K is the stiffness, L is the length of the journal bearing and D is the shaft diameter.

As is apparent from Fig. 6, the lower the L/D ratio, the higher the K/L ratio. Thus a plurality of bearings having an L/D ratio of 0.5 - 2 have heretofore been used for one spindle.

But with such a prior art spindle assembly it was possible but within some limit to increase the natural frequency by increasing the stiffness. If it is used at a revolving speed higher than the natural frequency, it is necessary to achieve a field balance with good accuracy.

If a tool or the like is attached to the head of the spindle, it is necessary to regain field balance every time even a small imbalance occurs when the tool is attached or detached. Otherwise, it would be difficult to rotate the spindle at a higher speed than the natural frequency.

Such field balancing work requires skilled hands and much labor. This will push up cost and hamper productivity.

Also, it is inefficient for an operator of the spindle assembly to adjust field balance at the job site.

For the above-described reasons, it was very difficult to use a prior art spindle assembly at a high revolving speed.

It is an object of this invention to provide a spindle assembly with an externally pressurized gas bearing which can restrain the spindle from running out of true and which do not require field balancing.

In accordance with the present invention, there is provided a spindle assembly comprising a spindle and a single journal bearing having a large bearing width for supporting the spindle, the journal bearing being formed near both ends thereof with a plurality of circumferentially arranged gas supply restrictor nozzles to supply compressed gas into clearance between the spindle and the journal bearing.

The spindle is supported by a single externally pressurized gas journal bearing having a large bearing width. The journal bearing is formed near its both ends with a plurality of gas supply restrictor nozzles so that the spindle can be supported without touching the bearing surface. This remarkably increases the damping coefficient and restrains the spindle from running out of true during high-speed rotation.

According to this invention, by taking into due consideration the damping coefficient which has heretofore been ignored, based on the results of the calculation, a single journal bearing having a large bearing width is used and it is formed with a plurality of gas supply restrictor nozzles arranged circumferentially at both axial ends thereof. This remarkably increases the damping coefficient/stiffness ratio of the bearing. This in turn improves the runout accuracy of the spindle during highspeed revolution.

Further, if no higher runout accuracy of the spindle than that of a prior art assembly is necessary, the accuracy in adjusting the imbalance of the spindle may be lowered to reduce the steps such as field balancing.

Because there is provided only one journal bearing, it becomes possible to reduce the number of parts and the number of tiny holes such as the gas supply restrictor nozzles, to simplify the supply/exhaust passage arrangement and to reduce the number of manufacturing steps. By reducing the number of gas supply restrictor nozzles for supporting the spindle, the bearing portion consumes less amount of compressed gas.

Other features and objects of the present invention will become apparent from the accompanying drawings, in which: Fig. 1 is a vertical sectional view of the first embodiment of the spindle assembly according to this invention; Fig. 2 is a vertical sectional view of the second embodiment; Fig. 3 is a vertical sectional view of the third embodiment; Fig. 4 is a view for explaining a vibration model; Fig. 5 is a vertical sectional view of a prior art spindle assembly with an externally pressurized gas bearing; Fig. 6 is a graph showing the relation among the length, diameter and stiffness of the journal bearing; and Fig. 7 is a graph showing the relation among the damping coefficient per unit length of the bearing and the length and diameter of the bearing.

Now the embodiments of this invention will be described with reference to Figs. 1 - 4 and 7.

By studying the influence of the stiffness and the damping coefficient on the displacement produced by an exciting force due to imbalance in a vibration model, the inventors found that more weight should be put on the damping coefficient than the stiffness. Thus the spindle assembly according to this invention has bearing portions having a large ratio of damping coefficient to stiffness.

Let us consider the vibration in a spindle-bearing system of a spindle assembly, which is simulated by a vibration model shown in Fig. 4 which comprises a point mass 21 having a mass m, a spring 22 having a spring constant K and a dash pot 23 having a damping coefficient C. The point mass 21 corresponds to the mass of spindle or its moment of inertia, and the spring 22 and the dash pot 23 correspond to the stiffness k and the damping coefficient of the bearing portions, respectively. If an exciting force having an amplitude F and an angular frequency & acts on the point mass 21, the equation of motion of the point mass is expressed as follows: m1 x c C X k x = F#23 # t (1) wherein x is the displacement of point mass and t is the time.Solving the equation (1) to obtain the amplitude XF of the point mass,

If the exciting force is generated due to imbalance of the spindle, the frequency of the exciting force is equal to the number of revolutions of the spindle. Its amplitude is given by F = wherein u is the amount of imbalance of the spindle. Thus the amplitude Xu is given by:

Because the natural frequency This of the vibration model is

tte amplitude of the runout due to imbalance is given by the following equation (4), provided

in the equation (3):

Thus, if the natural frequency is equal to the revolving speed, the runout due to imbalance will increase with the stiffness of the bearing portion.

Also, if ct is larger than the natural frequency

because k < mum 2, the larger the k value, the smaller the value of (k - my 2)2, provided the m and W values are constant. Thus, the denominators in the equations (2) and (3) will decrease. In other words, even if the exciting force has a larger frequency than the natural frequency, the larger the k value, the larger the degree of runout.

As will be apparent from the foregoing description, if the revolving speed during actual use is higher than the natural frequency of the spindle-bearing system, it is impossible to restrain the spindle from running out of true with the conventional construction in which the bearing portion has an increased stiffness.

As for the damping coefficient of the bearing, it will be apparent from the equation (4) that the larger the damping coefficient c, the smaller the degree of runout due to imbalance. The same is true with the denominators in the equations (2) and (3) Lecause the damping coefficient c is multiplied by the exciting frequency W . This means that the runout of the spindle can be more effectively restrained by reducing the stiffness while increasing the damping coefficient while the spindle is rotating at a higher speed than the natural frequency.

Fig. 7 shows the relationship between the C/L ratio (damping coefficient per unit length) and the L/D ratio, provided the dimensions of the bearing are the same as in Fig. 6. It will be apparent from this figure that the C/L ratio increases with the L/D ratio. The increase rate, however, drops after the L/D ratio has exceeded three.

According to this invention, the above results are applied to a spindle having a finite length and a single journal bearing having a large bearing width is used. Also, a plurality of gas supply restrictor nozzles are arranged circumferentially in two rows in the vicinity of both ends of the bearing.

Fig. 1 shows the first embodiment of this invention in which a spindle 32 mounted in a housing 31 is supported in a non-contact manner by a single sleeve-shaped externally pressurized gas journal bearing 33 and a pair of externally pressurized gas thrust bearings 35 and 36 arranged opposite to each other with a collar 34 of the spindle 32 sandwiched therebetween.

A plurality of gas supply restrictor nozzles 37 are arranged circumferentially in two rows near both ends of the journal bearing 33. Compressed gas is supplied through a compressed gas supply passage 39 formed in the housing 31 and through the gas supply restrictor nozzles 37 and gas supply restrictor nozzles 38 formed in the externally pressurized gas thrust bearings 35 and 36 to support the spindle 32 in a non-contact manner.

In the illustrated embodiment, the spindle 32 is directly driven by a rotor 40a and a stator 40b of a motor 40 and the bearing clearance communicates with the atmosphere through exhaust passages 41 to 43.

Because the journal bearing 33 is in the form of a single sleeve having a large bearing width and is formed near its both ends with gas supply restrictor nozzles 37, the damping coefficient/stiffness ratio of the bearing increases remarkably. Thus the damping coefficient is far higher than in the conventional bearing assembly. This makes it possible to restrain the spindle from running out of true during high-speed rotation.

Because the journal bearing 33 is in the form of a single sleeve, the bearing sleeve has to be longer than in the conventional bearings. Therefore, it is sometimes very difficult to provide one of the thrust bearings at one end of the journal bearing 33 so as to be integral with the journal bearing and to drill starting holes for the gas supply restrictor nozzles in an axial direction, as with the prior art example shown in Fig. 5.

In the first embodiment of Fig. 1, in order to avoid this, the thrust bearing 36 is separate from the journal bearing 33. But, while such a structure is easy to manufacture, the journal bearing 33 will be shorter by the length of the thrust bearings. Also, because the mass of a thrust plate and the motor rotor projects from the end of the journal bearing 33 for a larger length, it will become difficult to restrain the conical-mode runout of the spindle.

This problem is solved in the second embodiment shown in Fig. 2 in which one of the thrust bearings 36a is provided at one end of the journal bearing 33 so as to be integral therewith and starting holes 38a for the gas supply restrictor nozzles 38 are formed in the thrust bearing 36a in a radial direction.

Fig. 3 shows the third embodiment in which an air turbine is used as the driving source for the spindle 32 so that the assembly can be used for an electrostatic sprayer or the like. The same parts as the first embodiment shown in Fig. 1 are designated by the same numerals and their description is omitted.

In Fig. 3, the spindle 32 is supported by a sleeveshaped externally pressurized journal bearing 33 mounted in a housing 31 and is integrally formed on the rear end thereof with a large-diameter collar 51 which is formed on its outer peripheral surface with circumferentially arranged turbine blades 52 for receiving pressurized air. The spindle 32 is rotated at a high speed by blowing air through an air supply port 53 formed in the housing 31 against the turbine blades 52.

Compressed gas is supplied from the air supply port 53 to gas supply restrictor nozzles 38 provided at both sides of the large-diameter collar 51 to support the spindle 32 in a non-contact manner with respect to the thrust direction by the compressed gas which acts on both sides of the collar 51.

Since an air turbine is used as the driving source for the spindle 32 in the third embodiment, the size of the entire assembly can be reduced compared with when a motor is used.

In the embodiment shown in Fig. 3, gas is supplied to the air turbine and to the bearing through the single common supply port 53. But separate ports may be provided.

Further, in Fig. 3, a single sleeve-shaped journal bearing is mounted in the housing 31. But the sleeve 33 may be divided into a plurality of cylindrical parts, which may be inserted in the housing with their adjacent end faces abutting each other by press-fitting or shrink-fitting to form a single journal bearing.

Claims

WHAT IS CLAIMED IS:

1. A spindle assembly comprising a spindle and a single journal bearing having a large bearing width for supporting said spindle, said journal bearing being formed near both ends thereof with a plurality of circumferentially arranged gas supply restrictor nozzles to supply compressed gas into clearance between said spindle and said journal bearing.

2. A spindle assembly substantially as hereinbefore described with reference to and as shown in Figures 1 to 4 of the accompanying drawings.