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Chapter 14 Applications
Chapter
14
APPLICATIONS
14.1
INTRODUCTION
In the first chapter we have already mentioned that hydrostatic lubrication has been successfully applied in many branches of mechanical engineering, from large, slowly rotating machines to small and fast machines. In this chapter, a number of applications will be briefly described, beginning with the very important field of machine tools. Certain types of hydrostatic tilting pads used to build bearings for large machinery, such as telescopes, air preheaters, ore mills, debarking drums, and so on will then be examined. Lastly, after having mentioned a few applications of a different kind, a number of supply systems will be described, with particular reference to constant-flow systems making use of flow dividers or multiple pumps. 14.2 14.2.1
MACHINE TOOLS Spindles
Machine tool spindles form one of the most common fields of application of externally pressurized lubrication, since a high degree of stiffness and damping (i.e. precision characteristics) is required. Hydrostatic spindles may be supported by separate journal and thrust bearings, a s well as by a couple of opposed conical bearings; in certain cases other configurations may prove to be suitable: for instance, conical bearings may be substituted by spherical bearings, o r an opposed-pad bearing and a journal bearing may be com-
484
HYDROSTATC LUBRlCATlON
bined in a Yates configuration. Lubricant may be supplied directly, by means of multiple pumps, or a t a constant pressure (restrictor-compensated). The latter method is generally preferred because it is simpler. As a matter of fact, the compensating restrictors may be easily incorporated in the spindle housing; i t is therefore possible to build compact standardized units with only one inlet and one outlet port for lubricant: the supply system has merely to deliver lubricant at a given constant pressure and a t a temperature varying in a reasonably narrow range. Examples of spindles equipped with separate radial and axial bearings are to be found in Fig. 14.1 and Fig. 14.2. Figure 14.3.a shows how a combined journal and thrust bearing (see also section 8.7) may be used in a spindle, instead of conventional rolling bearings, Fig. 14.3.b. In this connection i t must be remembered that attempts have been made to produce ranges of hydrostatic bearings with outside and inside diameters following
V
6 Fig. 14.1 Hydrostatic spindle with journal and thrust bearings (compensating restrictors are not shown). (Reference 14.1).
U Fig. 14.2 Hydrostatic spindle with journal bearing and combined journal and opposed-pad thrust bearing. (Reference 14.1).
485
APPL /CATIONS
-a-
-b-
Fig 14.3 Hydrostatic spindle with a combined journal and thrust bearing (ref. 14.1).
the IS0 series for rolling bearings. In particular, the bearings depicted in Fig. 14.4 (ref. 14.21,mainly intended for use in machine tool spindles, follow the IS0 "0" series (their main dimensions are given in Table 14.1).After this first experimental range, another range was produced with similar dimensions and performance, but without the built-in seals, as shown in Fig. 14.5. -a-
-b-
Fig. 14.4 Standardized hydrostatic bearings: a- journal bearing; b- combined journal and thrust bearing (ref. 14.2).
In all the above units the journal bearings, of the multirecess type, with four recesses, are characterized by narrow lands: this has been done in order to obtain the greatest load capacity, while reducing the friction area and rise in temperature in the lubricant. The thickness of the film can be chosen in a small range of values, dependending on the stiffness and speed required, while the viscosity of the lubricant should be chosen, as usual, bearing minimum power consumption in mind. Journal bearings may also work without the inner ring.
HYDROSTATIC LUBRICATlON
T A B L E 14.1 Standardized hydrostatic bearing un d
D
(mm)
(mm)
(mm)
(mm)
50 60 70
80 95 110 125 140 150 170 180
60 70 80 90 105 115 130 140
75 90 100 110 130 140 160 170
80 90 100 110 120
-a-
dl
(see Fig. 14.4). B (mm)
Dl
Journal b. 68 80 88 98 110 118 126 140
-b-
Combined b. 75 85 95 106 115 125 136 152
-C -
Fig. 14.5 Standardized hydrostatic bearings: a- journal bearing; b- thrust bearing; c- combined journal and thrust bearing (ref. 2.2). Bearing units are usually fed at constant pressure and for this reason can be provided with laminar-flow restrictors, made up of a stack of special discs fitted in proper holes in the outer ring, very near the recesses of the bearing (Fig. 14.6.a). These discs are of two types: one is plain with a hole in its centre, whereas the other has a rectangular groove on both sides. Restriction is obtained i n the grooves, since they are shallow (however, not less than 80 pm). The total hydraulic resistance of the restrictor may be changed by varying the number of stacked discs. Another type of variable restrictor is shown in Fig. 14.6.b: in this case, a setscrew is used to adjust the hydraulic resistance. Bearings can also be feed a t a constant flow rate: this may be convenient especially for thrust bearings, in order to increase stiffness, at the cost of a slightly more complicated supply system.
487
APPLlCATlONS
t
-b-
-a-
t
1
2 3 2
3
4
c
.c
t
Fig. 14.6 Variable restrictors. a- Laminar-flow disc restrictor: 1-locking ring, 2-spacer disc, 3-restrictor disc, 4-bottom disc (ref. 14.2). b- Laminar-flow screw restrictor.
A typical application of the aforegoing standardized units is shown in Fig. 14.7. Another example is shown in Fig. 14.8: a spindle for a vertical grinding machine supported by two journal bearings and an opposed-pad thrust bearing (ref. 14.3). In the latter example a high degree of axial stiffness was required: for this reason it was decided to feed the thrust pads a t a constant flow rate, by means of a flow divider; the radial stiffness of the spindle, measured a t the nose, was found to be 180 N/pm under a 300 N load (the spindle diameter was 80 mm, the supply pressure 5 MPa), whereas axial stiffness was 500 N/pm under a 800 N load; maximum axial load was 14 KN, since the maximum supply pressure of the thrust bearings was limited to 8 MPa. Figure 14.9 shows a different type of spindle, used in a plane grinding machine, borne by a journal and an opposed-pad thrust bearing. Note that this type of combined bearing may be made to support large tilting moments using multirecess thrust pads (section 8.4) instead of the simpler annular-recess pads.
Fig. 14.7 Hydrostatic spindle with a journal bearing and a journal and thrust bearing (ref. 2.2).
488
HYDROSTATIC LUBRICATlON
E Fig. 14.8 Hydrostatic spindle for a grinding machine; the thrust bearing is fed at a constant flow rate (ref. 14.3).
Fig. 14.9 Hydrostatic spindle with a journal and a double-effect thrust bearing. (Reference 14.1).
For tapered-bearing spindles the most common configuration seems to be that to be found in Fig. 14.10, although different types of spindles have been built, for instance with cones arranged as in Fig. 8.19.a. Standardized spindle units are cur-
A PPLICA TIQNS
489
Fig. 14.10 Hydrostatic spindle with conical bearings. Ring R is used to adjust film thickness. (Reference 14.1).
rently produced, which are interchangeable with rolling-bearing o r hydrodynamic units produced by the same firm; of course, the main spindle dimensions comform with international standards for machine tools (ref 14.4). An example of a standard spindle unit is shown in Fig. 14.11and the main relevant data are to be found in Table 14.2(ref. 14.4).
L
a
-
Fig. 14.11 Standard spindle unit with cylindrical housing for boring, turning or milling (ref. 14.4).
Selection of the main hydrostatic parameters (number of recesses and their dimensions, film thickness, lubricant viscosity and so on) is generally made caseby-case by the manufacturer, on the basis of the operating conditions for which the spindle is designed (mainly load and velocity) and also on the basis of particular requirements, concerning stiffness and damping. Comparing the data in Table 14.2 with data concerning the equivalent spindle units equipped with ball bearings (ref. 14.41,it should be noted that the hydrostatic units show greater radial stiffness (although units equipped with special roller bearings are much stiffer). It should be borne in mind, however, that the stiffness of
490
HYDROSTAX LUBRlCATION
Size
D
a
d
(mm)
(mm)
(mm)
120
350
40
130
450
50
200
550
70
330
650
90
550
8
150 180 230 300
850
110
11
380
1050
150
3 4 5 6
CR
tl
t2
nmax
(W
@m)
(rpm)
0.5
0.5
8500
750
0.5 0.5 0.6 0.8
0.5 0.5 0.6 0.8
1000
1
1
7000 5500 4000 3000 2000
the hydrostatic units is proportional to supply pressure, and may be considerably affected by large axial loads (see section 8.5.2). A distinguishing feature of hydrostatic spindles is their very good running accuracy: values of tl and t 2 are always smaller than 1 pm, whereas the values of similar ball-bearing spindles range from 2 to 4 pm for tl and from 1.5 to 2 pm for t 2 (these values may even double int he case of roller-bearing spindles). Figure 14.12 shows an opposed-cone multirecess bearing that may be used to build hydrostatic spindles (ref. 2.11, as in Fig. 14.13. Note that, in this case, only the right-hand bearing sustains axial loads, whereas the other is used as a pure radial bearing.
Fig. 14.12 Hydrostatic opposed-cone bearing.
APPLICATIONS
491
Fig. 14.13 Hydrostatic spindle with a pair of opposed-conebearings. 14.2.2
Steady rests
Mounting of long and heavy rotors (e.g. turbine rotors, steel mill rolls, calenders, etc.) on lathes or other machine tools often requires the use of steady rests in order t o relieve the headstock and tailstock spindles from excessive loads and to reduce the bending of the axis of the workpiece. On the other hand, conventional steadies are characterized by high friction, with the relevant wearing and heating of the rubbing surfaces: these problems can be completely eliminated by means of hydrostatic lubrication.
A steady for a heavy machine tool may easily be built with a couple of self-aligning shoes (ref. 14.5)of the type shown in section 14.3.Each shoe must be mounted on a radially adjustable support to allow exact positioning of the workpiece. In the application described in ref. 14.6,steadies for sustaining rubber-coated cylinders (up to 600 KN in weight) on a grinding machine have been built. The hydrostatic shoe is provided not only with a spherical seat allowing tilt in all directions, but also with a screw and nut assembly for easily adjusting the radial position of the shoe (Fig. 14.14).I t should be noted that the intermediate piece of the bearing is fitted in a hydraulic cylinder which is widened in the base piece; pressure in the cylinder is the same as in the recess: in this way the fillets of the screw and nut are loaded with only a fraction of the force acting on the bearing. In this case the cylinder to be machined does not lean directly on the shoe bearings since intermediate rings are fitted on the necks of the cylinder: the same steadies can hence be used with different workpieces without needing to change the shoes, but using different rings, all of which have the same external diameter.
492
HYDROSTATIC 1UBRlCATION
Fig. 14.14 Adjustable hydrostatic shoe bearing (ref. 14.6).
14.2.3
Feed drives
Modern high precision machine tools require feed drives with high feeding accuracy, freedom from backlash and low friction. For these reasons recirculatingball lead screws and nuts are widely used. Hydrostatic lead-screw nuts meet the same requirements and also have other advantages as compared to recirculatingball nuts. In particular, they are inherently free from backlash (without the need for mechanical preload) and from wear (which ensures continuity of performance) and have better damping properties. This last feature has a certain importance in machines with roller-bearing or hydrostatic guideways, since the intrinsic lack of damping in the feed direction of frictionless guides can lead to poor stability against chatter in the same direction (ref. 14.7). Moreover, construction of the lead screw should be simpler in the case of hydrostatic nuts, since a very high degree of surface hardness is not required. Nevertheless, hydrostatic nuts are much less used than recirculating ball nuts (at least in small and medium-size machines). The main reasons are, probably, the following: recirculating-ball nuts are well proven and perform satisfactorily; hydrostatic lubrication requires a high-pressure lubricant source; construction of hydrostatic nuts is much more difficult and critical than other types of hydrostatic bear-
493
APPLlCATlONS
ings. This last is also obviously true for recirculating-ball units, but does not constitute a drawback in this case, since they are easily available in the stock of specialized manufacturers. Since some firms have recently begun to produce a wide range of standardized screw and nut assemblies, this type of feed system is expected to spread in the future. Data concerning a range of hydrostatic screws are to be found in Table 14.3 (ref. 14.8). The nut constitutes a compact unit, with built-in restrictors and seals, a n inlet port and an outlet port, requiring only a n adequate but fairly simple supply system. Feeding accuracy depends mainly on the pitch error of the male screw, but owing to the levelling effect of hydrostatic lubrication the manufacturer claims that actual feeding inaccuracy is less than one third of the pitch fluctuations of the male screw. Lo
T A B L E 14.3 Hydrostatic screw and nuts (ref. 14.8).
63
36.5
108
173 205 167 207
134 166 124 164
D
20
2.79 3.72 2.26 3.39
181 241 147 220
1.733 2.300 0.755 1.133
494
HYDROSTATIC L UBRICATlON
As already noted, the manufacture of hydrostatic nuts is somewhat difficult, either because of the relatively inaccessible position of the recesses, or because a small pitch difference in relation to the male screw can lead to a considerable loss of loading capacity (see section 7.3). Both problems can be easily overcome by means of a clever technique consisting in coating the inner surface of the nut with a thick layer of plastic, which is cast while the lead screw is held in position; recesses are obtained by means of patterns temporarily fixed to the flanks of the screw with an adhesive (note that hydrostatic nuts are in general of the multirecess type rather than of the continuous recesses type). The gap is obtained because of the shrinkage of cast plastic (ref. 14.9).
In large machine tools it may be preferable to substitute the screw-nut feed drive with rack and worm systems, which permit runs of practically any length, with a high degree of stiffness; furthermore, stiffness proves t o be independent from run length and the position of the slide. These systems can also obviously be assisted with hydrostatic lubrication. An example is that of the so-called hydrostatic "Johnson drive" (ref. 14.9)shown in Fig. 14.15.In this case, a short worm drives a long rack firmly fixed to the slide. The
Fig. 14.15 Hydrostatic Johnson drive (Ingersoll). 1-Slide, 2-rack, 3-pump pressure, 4-capillaries, 5-cells, 6-worm, 7-external gear teeth, 8-oil supply for forward flanks, 9-bed.
A PPLICA TIONS
495
worm is supported by means of hydrostatic thrust bearings; its circumference is toothed and is in mesh with a pinion driven by the feed gear. Recesses are hollowed in the flanks of the rack. A simple distributing device is needed to deliver lubricant only to the recesses covered by the worm, hence avoiding a considerable waste of power. In other applications (see, for instance, ref. 7.1 o r ref. 14.11) the rack is fixed to the bed, whereas the worm is supported by the slide, together with the relevant feed gear, which drives it by means of a toothed gear, fitted near the worm on the same shaft. Lubricant is supplied through ducts drilled in the worm; recesses may be hollowed in the flanks of the rack (as in Fig. 14.16)as well a s in the flanks of the male screw.
Fig. 14.16 Hydrostatic rack and worm;diarnete-270 mm,pitch=60 mm (INNSE).
In this case, too, a distributor is needed in order to cut off the high-pressure supply of lubricant to the ducts not ending on the flanks of the rack. When speed is high (speeds up to 750 rpm can be used) the centrifugal force may empty inactive ducts and that may cause aeration of the lubricant: hence the supply distributor should incorporate a pre-filling device whose task is to pump lubricant at low pres-
496
HYDROSTATIC LUBRICA TlON
sure into the inactive ducts, just before they become active again (a similar device is described in ref. 14.12). Hydrostatic worms are generally built with a pitch of between 36 and 60 mm and a n outside diameter of between 150 and 300 mm; load capacity may vary between 50 and 180 KN. The rack may be of virtually any length since it is built in sections (for instance, 1000 mm in length) that are bonded and firmly bolted to the slide bed after having been adjusted in relation to one another and measured to verify the pitch error (ref. 7.1). Accuracy may be about 70i-80 pm on a length of 25 m. Owing to this accuracy and to the very high degree of stiffness this feed system can also be used for monitoring the position of the slide during normal operation (by means of electronic compensation the relevant error can be further reduced to a very small value). 14.2.4
Guideways and rotating tables
Hydrostatic lubrication proves to be particularly suitable for guideways of modern high precision machine tools (especially those equipped with numerical control), because of their intrinsic characteristics: very low friction (and proportional to speed); freedom from stick-slip; freedom from wear (which means constancy of performance for an indefinite time); thickness of the oil film independent of the sliding speed (whereas for lubricated plain bearings it increases with speed); high damping capacity fin directions perpendicular to guide); levelling ability: the fairly high film thickness (commonly a few hundredths of a millimeter) allows the hydrostatic lubrication to compensate, a t least partially, for small geometric inaccuracies and deformations of the guides; possibility of building guides of virtually any length (which is difficult with roller guides). On the other hand, it should be noted that the virtual elimination of friction can enhance the effects of the flexibility of other parts of the machine and in particular of the feed drive. For instance, consider the experimental diagrams in Fig. 14.17 (see ref. 14.13 for further details): they refer to a milling machine and show that the displacement due to loading in the direction of the guides (mainly due to the flexibility of the ball screw and nut and of the relevant thrust bearing) is greatly reduced by the friction of the sliding ways. Diagrams in Fig. 11.18 (ref. 14.71, obtained with a similar experimental rig, show the reduction in damping connected with the use of frictionless guides (either
497
APPLlCATlONS
0
20000
kN
10000
load P
Fig. 14.17 Influence of hydrostatic ways on static stiffness, compared with sliding ways.
a- Hydrostatic system in action; b- without the hydrostatic system.
plain or ball screws were used as feed drives, without leading to any notably different behaviour). Problems of this kind are easily eliminated by means of simple clamping devices when feed rate is null, whereas in other cases they may be solved by stiffening the feed drive (for instance, in heavy machines, by selecting a worm and rack feed drive instead of screw and nut), by eliminating any backlash and increasing damp-
I
0
I
200
I
I
400
I
V
*
mmlmin
Fig. 14.18 Influence of feed rate V on maximum vibration amplitude A, (at resonance frequency) along feed direction for: a- sliding guideways; b- hydrostatic guideways; c- roller guideways.
498
HYDROSTATIC LUBRICATION
ing (for instance, introducing hydrostatic lubrication in the feed drive) or by means of external dampers (ref. 14.13). A number of different examples of layout for slideway guides are presented in Fig. 14.19; type ‘c’ and ‘d‘use an opposed-pad design: this is necessary when great stiffness and damping are required for a large range of loading conditions. The lower pads are in this case much smaller than the upper ones, in order to compensate for the weight of the slide.
-a-
-b-
-c-
-d-
Fig. 14.19 Sample layouts of hydrostatic guideways.
A compromise, often used in rotary tables, may consist i n substituting the preloading effect of the hydrostatic recesses on the underside of the guide with a spring force applied by means of rolling bearings (in practice, this is a trick for increasing the weight of the slide without increasing its mass). At least two recesses must be used on each guide to absorb torque, but a larger number of smaller recesses (each fed independently) provide greater compensating ability for the geometric inaccuracies of guideways; moreover, since the load is more evenly distributed on the guides, better results should also be obtained from the point of view of elastic distortion. Recesses may be either of the conventional fully-hollowed type, or be reduced to narrow grooves, as in the guides i n Fig. 14.20. From the point of view of static load capacity both designs perform i n the same way, but the narrow-groove recesses have greater damping ability and a larger bearing area in the absence of lubrication (hence, they are less prone to damage in the event of failure of the supply system). On the other hand, friction is also much higher and this type of recess proves to be adequate only for low-sliding velocities. The ability of hydrostatic lubrication to even out inaccuracies due to manufacturing errors or deformations caused by external forces is limited by the thickness of the lubricant film. Especially in the case of very large and heavily loaded slides
APPLlCA TlONS
499
Fig. 14.20 Rototraversing table equipped with hydrostatic lubrication of the guides (INNSE). In a the thrust bearing of the rotary table is shown; the pinions of the feed drive are also visible, as well as four clamps that may be used to fix the angular position of the table and the laminar-flow restrictors. In b the same table is shown from another angle: the linear guideways are visible, as well as two clamps.
500
HYDROSTATIC LUBRICATION
and rotary tables (such as the rotary table of a large vertical lathe), elastic deformation might even force the designer to select an excessively thick film to avoid metalto-metal contact. A solution may be to build the guideway with self-aligning tilting pads, as will be shown in section 14.3. The geometric inaccuracies of the slideways (e.g. waviness) might be completely compensated by controlling recess pressure: the principle is outlined i n Fig. 14.21 (see also ref. 14.10). Pressure in each recess is controlled by a valve, piloted by a regulator which compares a reference signal with the signal produced by a transducer. This last is, for instance, a pneumatic sensor which monitors the position of the slide in relation to a reference straight edge, or a photoelectric sensor, which uses a laser beam a s a reference "guide".
1
2
Fig. 14.21 Scheme of compensating bearing control. 1-Guide, 2-reference guide, 3-distance transducer, 4-regulator, 5-set value, 6-controlled valve, 7-supply pressure.
An example of hydrostatic lubrication applied to guideways is presented in Fig. 14.20,in which details are shown of a hydrostatic rototraversing table: one of a wide range of such equipment, suitable for indexing and contour milling (ref. 14.14)with a load capacity varying from 400 to 5000 KN. A similar range of rotary and rototraversing tables is also suitable for turning operations, with a turning speed of up to 2565 rpm, depending on the diameter of the table (2.5+10m).
The rotary table in Fig. 14.20 has a circular thrust bearing (with a mean diameter of 1400 mm) made up of 12 pads, all fed independently through a set of laminarflow restrictors. These are made by cutting small-diameter (111.5mm) pipes to the appropriate length and are also visible in the photographs. With a supply pressure of 6 MPa, the table can bear loads of up to 600 KN. The radial forces are sustained by a tapered roller bearing, which also exerts a preloading force (150KN) on the hydrostatic thrust bearing, in order to increase its stiffness. The photographs also show
APPL ICATlONS
501
clamps that are able to hold the table firmly in any position, without affecting the film thickness of the hydrostatic bearings. Rotary motion is obtained by means of two controlled-preload pinions meshing with a helical crown gear, whereas a ball screw is used for linear axis transmission (the largest members of the same family of tables use hydrostatic worms and racks for axial feed drive). Hydrostatic lubrication is often also applied to the guides of ram-type milling arms (Fig. 14.22). The design of the guides is of course different from that of the guides of horizontal tables: in this case the ram is supported by two rows of eight recesses (two for each side). The recesses in the lower end of the guide are generally larger since they must support higher loads (in other applications there are three rows of recesses, two of which are set at the lower end of the guide). The supply system is made up of a set of multiple pumps (each pump directly feeds one recess),
Fig. 14.22.a- Hydrostatically lubricated milling arm (Pensotti).
502
HYDROSTATC LUBRlCAT/ON
Fig. 14.22.b- Hydrostatically lubricated milling arm:detail showing hydrostatic pads.
which are fed at constant pressure (-2.5MPa) by a larger pump. In this case, too, the recesses of the pads (which are made of bronze) are reduced to narrow grooves.
It is interesting that hydrostatic lubrication has also been used to compensate for the deflection of the ram due to the cutting force. The geometric adaptive control system described in ref. 14.15 measures the displacement of the milling head by means of a laser gun fixed to the milling arm, which emits a laser beam parallel to the undeformed axis of the ram, and a photoelectric scanner attached to the milling head. The signal produced by the measuring equipment is taken a s its input by a control unit which varies accordingly the speed of a servo-motor driving a further set of pumps. The flow produced by these compensating pumps is directed towards the appropriate recesses and added to the normal flow in order to produce a displacement of the milling head, realigning it with the laser beam.
A P P l ICA TlONS
503
A particular application of externally pressurized lubrication to the ram guide of a gear-shaping machine is described in ref. 14.10. A cross section of the guide is shown in Fig. 14.23: the ram is shaped like a spur gear with every third tooth removed. The accuracy of the internal bore of the sleeve is obtained by casting with a plastic material (this technique is briefly described in section 14.2.3).
Fig. 14.23 Hydrostatic ram guide of a gear-shaping machine (Liebherr). 14.3
LARGE TILTING PADS
Hydrodynamic bearings for very large rotating machine-members have been equipped for many years now with high-pressure hydrostatic pockets, used as jacking devices at starting (hydrostatic lifts). More recently, i t has been found to be expedient to retain the hydrostatic effect in normal running and then to substitute the hydrodynamic bearings completely with hydrostatic (or hybrid) bearings, in the case of slowly rotating machines in particular, or when irregularities in load or speed are expected. One problem connected with this type of bearing in certain machines (such as ore mills) is that the elastic deformation of the runner, due to the pressure of the lubricant, may greatly reduce the effectiveness of hydrostatic lubrication (Fig. 11.24.a). This problem may be overcome by foregoing the "optimum" design, obtained by assuming rigid surfaces and uniform film thickness, and displacing the recesses from the centre of the bearing (Fig. 11.24.b); separate pads may even be used instead of a multirecess bearing (ref. 14.16).
A further improvement in design, able to eliminate most of the problems connected with elastic deformation, machining tolerance, thermal expansion and so
504
HYDROSTATIC LUBRICA TlON
Fig. 14.24 Trunnion deformation due to bearing pressure: a- bearing as designed; b- improved concept; e- most effective concept (ref. 14.16).
on, consists in supporting the large journal by means of a set of self-aligning hydrostatic shoes, as shown in Fig. 14.25 (ref. 14.17). Each shoe is split up into two parts: the upper part rests on a spherical seat and hence can tilt in all directions. The underside of the upper part is shaped like a piston which fits into a cylinder in the base: since the piston area, on which the recess pressure acts, is slightly smaller than the effective area of the pad, the load on the spherical seat is quite low during normal operation.
Fig. 14.25 Arrangement of tilting-pad hydrostatic bearings (ref. 14.17).
505
APPLICATIONS
The spherical rest of each inner shoe (slave shoe) is pushed against the runner by a further piston on which, thanks to a hydraulic connection, the recess pressure of the relevant outer shoe (master shoe) acts. Clearly, if the sum of the two piston areas equals the effective pad area, the slave shoe must necessarily have the same film thickness, and thus the same recess pressure, a s the relevant master shoe (each pad is fed by the same flow rate). Thus when the load direction is vertical all four shoes have the same film thickness and recess pressure, regardless of the deviation of the runner from the ideal circular shape. When the load deviates from the vertical direction the two shoes on each side have an equal part of the load component falling along the line between the two shoes (ref. 14.17). The shape of the recess is also of particular interest. It is known that when a cylindrical pad with a simple recess (as in Fig. 5.30) is tilted from the concentric configuration the pressure field on the land surface is altered and produces a moment that tends to realign the pad; however, this self-aligning capacity is too small to ensure the stability of the shoe in all conditions and i t is hence necessary to use multirecess pads. In Fig. 14.26 the main recess is surrounded by four auxiliary recesses, situated in the corners of the pad, which are fed with the lubricant which passes from the central recess over the bearing lands and through small drilled ducts (this is a compromise aimed a t avoiding dependence upon the direction of rotation: for the greatest stability the auxiliary recesses on the trailing side should only be supplied over the lands). The hydrostatic system described in ref. 14.17 supported a large tube mill for crushing ores: each bearing runner had a diameter of 2700 mm, the maximum
I -.
I
I
L Fig. 14.26 Improved recess pattern (ref. 14.17).
-
W
506
HYDROSTAT C LUBRlCAT/ON
load was 3500 KN and the velocity was 0.24 reds. Each shoe was 640 mm long and 500 mm wide and was fed at 25 Ym with a lubricant whose viscosity was 0.1 Ns/m2 at 50°C. Film thickness in normal operation was 0.1410.15 111111.
A range of hydrostatic shoes based on the foregoing working principles is currently produced by the same firm (ref. 14.5): a sketch of them is to be found in Fig. 14.27 and their main dimensions are given in table 14.4. The recess pattern is similar t o that shown in Fig. 14.26, but the main recess is now annular in shape, in order t o increase the bearing area a t rest (in the absence of hydrostatic lubrication) virtually without affecting bearing performance during normal operation. Hydrostatic shoes may be used to support horizontal as well as vertical rotating equipment. In the first case the rotating drum may lean on the shoes by means of trunnions (Fig. 14.28.a) or by means of girth rings (Fig. 14.28.b). The latter arrangement, which is often inapplicable with rolling bearings due to their size limits, permits large feed openings and a simplified (and less expensive) design. Each -b-
-a-
Fig. 14.27 Bearing shoes: a- master shoe; b- slave shoe. T A B L E 14.4 H
410 SO0 600
530 640 756
(mm)
Master
Slave
300
180+190 2601270 2951305 320+330 4201430
425
507
APPLlCA TlONS
ring (or trunnion) is supported by two master shoes, to each of which one or two slave shoes may be added to boost the load-carrying capacity (Fig. 14.29). The suggested ring diameter D varies between 500 and 5400 mm, with a load capacity F ranging from 480 to 12000 KN, depending on pad size and the total number of pads. -a-
-b-
Fig. 14.28 Horizontal rotating arrangements: a- trunnion arrangement; b- girth ring arrangement.
D
IMaster
shoe: O S l a v e shoe
Fig. 14.29 Shoe arrangements for horizontal rotating cylinders.
In the case of vertical equipment three master shoes are obviously required in order t o obtain a statically determined load distribution; to each master shoe a slave unit can be added, thereby doubling the load capacity. A typical arrangement is shown in Fig. 14.30, in which two alternatives are also proposed for the radial guidance of the runner: a rolling bearing mounted on the shaft, or a set of hydrostatic guiding pads (see below). For six pad arrangements the load capacity ranges from 1900 t o 14000 KN (depending on the pad size) and correspondingly the minimum pitch diameter D varies from 800 to 2400 mm. Besides the hydrostatic shoes described above, the same firm produces a range of smaller tilting pads of simplified design (see Fig. 14.31 and table 14.5). These still retain a self-aligning capacity, since they have a spherical seat and multiple recesses, but are not equipped with hydraulic cylinders. They are mainly proposed (ref. 14.5) as guiding pads for the axial location of a girth ring (Fig. 14.32) or for the
508
HYDROSTATIC LUBRICATION
D
r=
ALT I
I
ALT II
Fig. 14.30 Shoe bearing arrangement for vertical rotating equipment. radial guiding of platforms (Fig. 14.30). Compact assemblies are also available consisting in a master shoe bearing with two guiding pads (in an opposed-pad configuration) mounted on the fixed part of the shoe (Fig. 14.32.b). Another type of tilting pad (ref. 14.18) can be used to build spherical thrust bearings with a very large diameter. In practical terms, it consists of a circular recess pad laid on a spherical rest whose position can be adjusted by means of a wedge. In Fig. 14.33 a set of twenty pads is used to build a large bearing (with a mean diameter of 5000 mm) for a large parabolic antenna. The dimensions of the bearing and angle a depend on the value of the axial and radial components of the load: bearings with an external diameter of up to 8000 mm can be built.
APPLlCATlONS
509
Fig. 14.31 Guiding pad.
TABLE 14.5 Dimensions of guiding pads (ref. 14.5).
Fig. 14.32 Axial guiding pads: a- separate axial guidance; b- axial guidance integral with a master shoe.
A further type of tilting shoe is shown in Fig. 14.34(ref. 14.19):it can tilt around the cylindrical rib on the underside and align itself thanks to the multiple recesses (two or four) which are fed independently through capillary restrictors. These pads can also be used to sustain radial loads as well as the axial thrust of a large rotating platform. In the latter case, bearings with diameters exceeding 5000 mm may be built, which sustain thrusts greater than 5000 KN and rotating a t more than 20 rpm. These pads prove to be particularly suitable for building rotary tables for large machine tools (e.g. for vertical lathes): an example is given in Fig. 14.35.A different application is described in ref. 14.20,concerning the supporting ring of a 3.5 m telescope.
510
HYDROSTATIC LUBRICATION
P I
Fig. 14.33 Spherical pad arrangement,
Fig. 14.34 "Hydro-tilt"shoe bearing (ref. 14.19).
APPLlCA TlONS
51 1
Fig. 14.35 "Hydro-tilt" shoe arrangement.
Lastly, Fig. 14.36 shows a large-size spherical bearing (ref. 2.2); it has three recesses fed a t a constant flow rate. Bearings like this can sustain heavy loads (up to 10,000 KN) and in general their rotating speed is low. For instance, the bearing depicted in Fig. 14.36 was made to support the rotor of an air preheater weighing 800 KN and rotating at 2 rpm.
14.4
OTHER APPLICATIONS
Apart from those quoted in the foregoing sections, hydrostatic lubrication has a number of different applications. For instance, let us consider the pump in Fig. 14.37 (ref. 14.21): the pistons (1) lean on the tilted plate (2) by means of the spherical pads (3) which are hydrostatically borne by the same circulating fluid. Another special application is quoted in ref. 14.22, that is the lower journal bearing of the main pump of the Super-PhBnix nuclear power plant. This bearing has a
51 2
HYDROSTATIC LUBRICATlON
L.L.1
r't'i
Fig. 14.36 Large spherical bearing.
Fig. 14.37 Piston pump.
diameter of 0.85 m and a width of 0.3 m; it has twelwe recesses carved in the shaft. In this case, too, the lubricant is the fluid circulating in the plant, i.e. liquid sodium. Hydrostatic lubrication has been successfully used in a number of testing rigs. An example is shown in Fig. 14.38:an experimental rig for testing rolling bearings (ref. 14.23).The bearing being tested (1) is made to rotate by a motor (2)by means of a
APPLICATIONS
513
Fig. 14.38 Experimental rig for rolling bearings.
belt drive (3) and is loaded by a jack (4) through the hydrostatic bearing (5);this last leans on a cell (61, which measures load Fa, by means of a spherical seat. Friction moment M R is measured by means of dynamometer (7)and the angular speed n by means of thacheometer (8). 14.5 14.5.1
HYDRAULIC CIRCUITS
Simple layout
A typical supply system for hydrostatic spindles, such as the one in Fig. 2.24, is shown in Fig. 14.39 (ref. 2.2). The bearings are fed at a constant pressure, which is usually in a range between 3 and 7 MPa. A gear pump supplies lubricant at a rate which is 30% greater than the calculated value: the surplus flows back to the reservoir through the pressure regulating valve. Lubricant is pushed through two filters, the first of which is coarser (15 pm), while the other is narrower (5-10pm). A pressure switch prevents the spindle from running until the pressure reaches the established value and stops it when pressure falls: in the last case, an oil accumulator
51 4
HYDROSTATIC LUBRICATION
Fig. 14.39 Supply system for a hydrostatic spindle: 1-oil tank; 2-pump; 3-motor; 4-pressure regulating valve; 5-pressure filter; 6-pressure switch; 7-check valve; 8-piston accumulator; 9-pressure gauge; 10-cooler; 11-thermostaticsystem; 12-heater.
1
2
Fig. 14.40 Supply system for the hydrostatic bearing of an air preheater: 1-pump; 2-motor; 3-pressure filter; 4-pressure switch; 5-check valve; 6-flow divider; 7-piston accumulator; 8-pressure-limiting valve; 9-cooler.
515
APPLICA TlONS
keeps on feeding the bearings for the time needed for the spindle to come to a complete stop. A thermostatic system keeps the temperature of the lubricant close to the design value. Sometimes a further pump may also be needed (generally inserted upstream from the cooler) to bring the lubricant back from the spindle to the reservoir. Flow dividers
14.5.2
Figure 14.40 shows the supply circuit for the three-recess preheater bearing in Fig. 14.36.The flow rate produced by the main pump is divided into three equal streams by means of a flow divider made up of three equal gear pumps mounted on
8
1
1
13
9
\
5
6 4
3
2
17
15
16
14
1
Fig. 14.41 Supply system for the hydrostatic bearing of an ore mill: 1-oil tank; 2-pump; 3-pressure switch; 4-pressure filter; 5-check valve; 6-pressure limiting valve; 7-pressure gauge; 8-piston accumulator; 9-nitrogen gas bottle; 10-flow divider; 11-shoe bearing; 12-circulation pump for cooling circuit 13-throttle valve; 14-oil cooler; 15-water flow control valve; 16-temperature-sensing device.
51 6
HYDROSTATIC LUBRICATlON
a common shaft. To ensure continuous operation a second pump is ready to be started up automatically when supply pressure drops below a safe value. A further spare pump is available for replacement, to permit maintenance operations. In the case of a n electric mains failure a diesel generator can provide power for the motors of the pumps. The last emergency device is a set of oil accumulators which can supply lubricant to the bearings for a short time. The hydraulic circuit for the bearing arrangement in Fig. 14.25 is shown in Fig. 14.41. The flow rate produced by the main pump is divided into four equal streams by means of a flow divider. To ensure continuous operation a second pump is ready to be started up automatically in case of failure of the other one and a set of piston accumulators (driven by pressurized nitrogen bottles) can feed oil to the bearings for a certain time in case of power failure, allowing the runner to stop without damaging the bearings. 14.5.3
Multiple pumps
The constant-flow supply circuit of the guideway presented in Fig. 14.22.b is shown in Fig. 14.42. The pre-feeding pump (1)delivers lubricant at a pressure of 25 bar to two multiple pumps (2). Each pump can feed ten recesses independently, each with a 0.33 m3/s flow rate, a t a pressure of 40 bar.
Fig. 14.42 Supply system, with multiple pumps, of a hydrostatic slide.
A PPLlCATlONS
517
Figure 14.43 shows the supply circuit of the pad arrangement in Fig. 14.33. Each pad is directly fed a t a constant flow rate; that is, a set of five multiple pumps is used and each pump delivers four equal streams which are supplied to four pads situated a t 90 degrees from each other. Thanks to the layout mentioned, emergency operation of bearing system is possible even if a pump fails.
Ic
Fig. 14.43 Supply system for the bearing system of a large-beam antenna: M-motor; P-multiple pump; 1-20 pads.
REFERENCES
- Herzstuck Leistungsfahiger Werkzeugmaschinen; FAG publ. 02-113DA (1985);68 pp. Hallstedt G.;Standardized Hydrostatic Bearing Units; Instn. Mech. Engrs., C48 (1971),420-430. Lewinschal L.; Contributo dei Cuscinetti Zdrostatici allXumento d i Produttivitb delle Rettificatrici; La Rivista dei Cuscinetti/SKF, 196 (19781,24-27. FAG Spindeleinheiten fur das Bohren-Drehen-Frasen; FAG publ. 02-1OW3 DA (1985);12 pp. Hydrostatic Shoe Bearing Arrangements; SKF Publication 3873 E (19881,28 PP. Bi l dt sh C., Htillnor G.; Problema Risolto con 1'Adozione di Pattini Idrostatici;La Rivista dei CuscinettYSKF, 181 (1974),18-20. Polseck M.,Vavra Z.; The influence of different types of guideways on the static and dynamic behaviour of feed drives; Proc. 8th Int. MTDR Conf. (19671, pt. 2, p. 1127-1138.
14.1 Die Arbeitsspindel und Hire Lagerung 14.2 14.3 14.4 14.6 14.6 14.7
518
HYDRCSTATIC LUBRICATlON
14.8 Catalog B1025E; Nachi Corp., Japan, 1984; 4 p. 14.9 Weck M.; Handbook of Machine Tools, Volume 2 (Construction and Mathe-
matical Analysis); J. Wiley & Sons, 1980; 296 pp.
14.10 Rohs H. G.; Die Hydostatische Bewegungspaarung i m Werkzeugmaschinenbau; Konstrudion, 22 (1970); 321-329. 14.11 Andreolli C.; Eliminazione dell'Attrito e dei Giochi nelle Macchine Utensili;
Controlli NumericiIMacchine a CN/Robot Industriali, anno XI1 (19791, n. 7, p. 32-45. 14.12 Appoggetti P.; Perfezionamento negli Accoppiamenti Vite-Cremagliera a Sostentamento Zdrostatico; Patent IT 51829 N69; Bollettino Tecnico RTM n. 9, 1969; p. 47-51. 14.13 Umbach R., Haferkorn W.; Some Examples and Problems in Zmplementation of Mwlern Design Features on Large Size Machine Tools; 10th Int. MTDR cod., Manchester, 1969; paper 34; 30 pp. 14.14 Rototraversing Tables for Indexing, Milling and Turning; INNSE Publication DMU/27 (1985),4pp. 14.15 Weck M.; Handbook of Machine Tools, Volume 3 (Automation and Controls); J . Wiley & Sons, 1980; 451 pp. 14.16 Rippel T., Hunt J. B.; Design and Operational Experience of 102-Znch Diameter Hydrostatic Journal Bearings for Large Size Tumbling Mills; Instn. Mech. Engrs., C16 (1971), 76-100. 14.17 Arsenius H. C., Goran R.; The Design and Operational Experience of a SelfAdjusting Hydrostatic Shoe Bearing for Large Size Runners; Instn. Mech. EWS., C303 (19731,361-367. 14.18 Supporti idrostatici FAG; FAG Publication 44109 IB (19711, 8 pp. 14.19 Andreolli C.; Guida Circolare Idrostatica Assiale per Tavola Portapezzo Rotante; Patent IT 2353CA, 1975; 15 pp. 14.20 Andreolli C.; Sopporto Zdrostatico per 1'Asse Azimutale del 3.5 m New Technology Telescope (NTT) dell %SO; Convegno AIM-AMME (Tribologia-Attrito, Usura e Lubrificazione), Sorrento, 1987; p. 421-430. 14.21 Giordano M., Boudet M.; Thermohydrodynamic Flow of a Piezoviscous Fluid Between Two Parallel Discs; J. Mech. Eng., 1980. 14.22 F r h e J., Nicolas D., Deguerce B., Berthe D., Godet M.; Lubrification Hydrodynamique; Edition Eyrolles, Paris, 1990; 488 pp. 14.23 Martin F. J.; Prove Funzionali e di Qualificazione nello Sviluppo dei Cuscinetti Volventi; La Rivista dei CuscinettiBKF, 224 (1986),28-36.
14
APPLICATIONS
14.1
INTRODUCTION
In the first chapter we have already mentioned that hydrostatic lubrication has been successfully applied in many branches of mechanical engineering, from large, slowly rotating machines to small and fast machines. In this chapter, a number of applications will be briefly described, beginning with the very important field of machine tools. Certain types of hydrostatic tilting pads used to build bearings for large machinery, such as telescopes, air preheaters, ore mills, debarking drums, and so on will then be examined. Lastly, after having mentioned a few applications of a different kind, a number of supply systems will be described, with particular reference to constant-flow systems making use of flow dividers or multiple pumps. 14.2 14.2.1
MACHINE TOOLS Spindles
Machine tool spindles form one of the most common fields of application of externally pressurized lubrication, since a high degree of stiffness and damping (i.e. precision characteristics) is required. Hydrostatic spindles may be supported by separate journal and thrust bearings, a s well as by a couple of opposed conical bearings; in certain cases other configurations may prove to be suitable: for instance, conical bearings may be substituted by spherical bearings, o r an opposed-pad bearing and a journal bearing may be com-
484
HYDROSTATC LUBRlCATlON
bined in a Yates configuration. Lubricant may be supplied directly, by means of multiple pumps, or a t a constant pressure (restrictor-compensated). The latter method is generally preferred because it is simpler. As a matter of fact, the compensating restrictors may be easily incorporated in the spindle housing; i t is therefore possible to build compact standardized units with only one inlet and one outlet port for lubricant: the supply system has merely to deliver lubricant at a given constant pressure and a t a temperature varying in a reasonably narrow range. Examples of spindles equipped with separate radial and axial bearings are to be found in Fig. 14.1 and Fig. 14.2. Figure 14.3.a shows how a combined journal and thrust bearing (see also section 8.7) may be used in a spindle, instead of conventional rolling bearings, Fig. 14.3.b. In this connection i t must be remembered that attempts have been made to produce ranges of hydrostatic bearings with outside and inside diameters following
V
6 Fig. 14.1 Hydrostatic spindle with journal and thrust bearings (compensating restrictors are not shown). (Reference 14.1).
U Fig. 14.2 Hydrostatic spindle with journal bearing and combined journal and opposed-pad thrust bearing. (Reference 14.1).
485
APPL /CATIONS
-a-
-b-
Fig 14.3 Hydrostatic spindle with a combined journal and thrust bearing (ref. 14.1).
the IS0 series for rolling bearings. In particular, the bearings depicted in Fig. 14.4 (ref. 14.21,mainly intended for use in machine tool spindles, follow the IS0 "0" series (their main dimensions are given in Table 14.1).After this first experimental range, another range was produced with similar dimensions and performance, but without the built-in seals, as shown in Fig. 14.5. -a-
-b-
Fig. 14.4 Standardized hydrostatic bearings: a- journal bearing; b- combined journal and thrust bearing (ref. 14.2).
In all the above units the journal bearings, of the multirecess type, with four recesses, are characterized by narrow lands: this has been done in order to obtain the greatest load capacity, while reducing the friction area and rise in temperature in the lubricant. The thickness of the film can be chosen in a small range of values, dependending on the stiffness and speed required, while the viscosity of the lubricant should be chosen, as usual, bearing minimum power consumption in mind. Journal bearings may also work without the inner ring.
HYDROSTATIC LUBRICATlON
T A B L E 14.1 Standardized hydrostatic bearing un d
D
(mm)
(mm)
(mm)
(mm)
50 60 70
80 95 110 125 140 150 170 180
60 70 80 90 105 115 130 140
75 90 100 110 130 140 160 170
80 90 100 110 120
-a-
dl
(see Fig. 14.4). B (mm)
Dl
Journal b. 68 80 88 98 110 118 126 140
-b-
Combined b. 75 85 95 106 115 125 136 152
-C -
Fig. 14.5 Standardized hydrostatic bearings: a- journal bearing; b- thrust bearing; c- combined journal and thrust bearing (ref. 2.2). Bearing units are usually fed at constant pressure and for this reason can be provided with laminar-flow restrictors, made up of a stack of special discs fitted in proper holes in the outer ring, very near the recesses of the bearing (Fig. 14.6.a). These discs are of two types: one is plain with a hole in its centre, whereas the other has a rectangular groove on both sides. Restriction is obtained i n the grooves, since they are shallow (however, not less than 80 pm). The total hydraulic resistance of the restrictor may be changed by varying the number of stacked discs. Another type of variable restrictor is shown in Fig. 14.6.b: in this case, a setscrew is used to adjust the hydraulic resistance. Bearings can also be feed a t a constant flow rate: this may be convenient especially for thrust bearings, in order to increase stiffness, at the cost of a slightly more complicated supply system.
487
APPLlCATlONS
t
-b-
-a-
t
1
2 3 2
3
4
c
.c
t
Fig. 14.6 Variable restrictors. a- Laminar-flow disc restrictor: 1-locking ring, 2-spacer disc, 3-restrictor disc, 4-bottom disc (ref. 14.2). b- Laminar-flow screw restrictor.
A typical application of the aforegoing standardized units is shown in Fig. 14.7. Another example is shown in Fig. 14.8: a spindle for a vertical grinding machine supported by two journal bearings and an opposed-pad thrust bearing (ref. 14.3). In the latter example a high degree of axial stiffness was required: for this reason it was decided to feed the thrust pads a t a constant flow rate, by means of a flow divider; the radial stiffness of the spindle, measured a t the nose, was found to be 180 N/pm under a 300 N load (the spindle diameter was 80 mm, the supply pressure 5 MPa), whereas axial stiffness was 500 N/pm under a 800 N load; maximum axial load was 14 KN, since the maximum supply pressure of the thrust bearings was limited to 8 MPa. Figure 14.9 shows a different type of spindle, used in a plane grinding machine, borne by a journal and an opposed-pad thrust bearing. Note that this type of combined bearing may be made to support large tilting moments using multirecess thrust pads (section 8.4) instead of the simpler annular-recess pads.
Fig. 14.7 Hydrostatic spindle with a journal bearing and a journal and thrust bearing (ref. 2.2).
488
HYDROSTATIC LUBRICATlON
E Fig. 14.8 Hydrostatic spindle for a grinding machine; the thrust bearing is fed at a constant flow rate (ref. 14.3).
Fig. 14.9 Hydrostatic spindle with a journal and a double-effect thrust bearing. (Reference 14.1).
For tapered-bearing spindles the most common configuration seems to be that to be found in Fig. 14.10, although different types of spindles have been built, for instance with cones arranged as in Fig. 8.19.a. Standardized spindle units are cur-
A PPLICA TIQNS
489
Fig. 14.10 Hydrostatic spindle with conical bearings. Ring R is used to adjust film thickness. (Reference 14.1).
rently produced, which are interchangeable with rolling-bearing o r hydrodynamic units produced by the same firm; of course, the main spindle dimensions comform with international standards for machine tools (ref 14.4). An example of a standard spindle unit is shown in Fig. 14.11and the main relevant data are to be found in Table 14.2(ref. 14.4).
L
a
-
Fig. 14.11 Standard spindle unit with cylindrical housing for boring, turning or milling (ref. 14.4).
Selection of the main hydrostatic parameters (number of recesses and their dimensions, film thickness, lubricant viscosity and so on) is generally made caseby-case by the manufacturer, on the basis of the operating conditions for which the spindle is designed (mainly load and velocity) and also on the basis of particular requirements, concerning stiffness and damping. Comparing the data in Table 14.2 with data concerning the equivalent spindle units equipped with ball bearings (ref. 14.41,it should be noted that the hydrostatic units show greater radial stiffness (although units equipped with special roller bearings are much stiffer). It should be borne in mind, however, that the stiffness of
490
HYDROSTAX LUBRlCATION
Size
D
a
d
(mm)
(mm)
(mm)
120
350
40
130
450
50
200
550
70
330
650
90
550
8
150 180 230 300
850
110
11
380
1050
150
3 4 5 6
CR
tl
t2
nmax
(W
@m)
(rpm)
0.5
0.5
8500
750
0.5 0.5 0.6 0.8
0.5 0.5 0.6 0.8
1000
1
1
7000 5500 4000 3000 2000
the hydrostatic units is proportional to supply pressure, and may be considerably affected by large axial loads (see section 8.5.2). A distinguishing feature of hydrostatic spindles is their very good running accuracy: values of tl and t 2 are always smaller than 1 pm, whereas the values of similar ball-bearing spindles range from 2 to 4 pm for tl and from 1.5 to 2 pm for t 2 (these values may even double int he case of roller-bearing spindles). Figure 14.12 shows an opposed-cone multirecess bearing that may be used to build hydrostatic spindles (ref. 2.11, as in Fig. 14.13. Note that, in this case, only the right-hand bearing sustains axial loads, whereas the other is used as a pure radial bearing.
Fig. 14.12 Hydrostatic opposed-cone bearing.
APPLICATIONS
491
Fig. 14.13 Hydrostatic spindle with a pair of opposed-conebearings. 14.2.2
Steady rests
Mounting of long and heavy rotors (e.g. turbine rotors, steel mill rolls, calenders, etc.) on lathes or other machine tools often requires the use of steady rests in order t o relieve the headstock and tailstock spindles from excessive loads and to reduce the bending of the axis of the workpiece. On the other hand, conventional steadies are characterized by high friction, with the relevant wearing and heating of the rubbing surfaces: these problems can be completely eliminated by means of hydrostatic lubrication.
A steady for a heavy machine tool may easily be built with a couple of self-aligning shoes (ref. 14.5)of the type shown in section 14.3.Each shoe must be mounted on a radially adjustable support to allow exact positioning of the workpiece. In the application described in ref. 14.6,steadies for sustaining rubber-coated cylinders (up to 600 KN in weight) on a grinding machine have been built. The hydrostatic shoe is provided not only with a spherical seat allowing tilt in all directions, but also with a screw and nut assembly for easily adjusting the radial position of the shoe (Fig. 14.14).I t should be noted that the intermediate piece of the bearing is fitted in a hydraulic cylinder which is widened in the base piece; pressure in the cylinder is the same as in the recess: in this way the fillets of the screw and nut are loaded with only a fraction of the force acting on the bearing. In this case the cylinder to be machined does not lean directly on the shoe bearings since intermediate rings are fitted on the necks of the cylinder: the same steadies can hence be used with different workpieces without needing to change the shoes, but using different rings, all of which have the same external diameter.
492
HYDROSTATIC 1UBRlCATION
Fig. 14.14 Adjustable hydrostatic shoe bearing (ref. 14.6).
14.2.3
Feed drives
Modern high precision machine tools require feed drives with high feeding accuracy, freedom from backlash and low friction. For these reasons recirculatingball lead screws and nuts are widely used. Hydrostatic lead-screw nuts meet the same requirements and also have other advantages as compared to recirculatingball nuts. In particular, they are inherently free from backlash (without the need for mechanical preload) and from wear (which ensures continuity of performance) and have better damping properties. This last feature has a certain importance in machines with roller-bearing or hydrostatic guideways, since the intrinsic lack of damping in the feed direction of frictionless guides can lead to poor stability against chatter in the same direction (ref. 14.7). Moreover, construction of the lead screw should be simpler in the case of hydrostatic nuts, since a very high degree of surface hardness is not required. Nevertheless, hydrostatic nuts are much less used than recirculating ball nuts (at least in small and medium-size machines). The main reasons are, probably, the following: recirculating-ball nuts are well proven and perform satisfactorily; hydrostatic lubrication requires a high-pressure lubricant source; construction of hydrostatic nuts is much more difficult and critical than other types of hydrostatic bear-
493
APPLlCATlONS
ings. This last is also obviously true for recirculating-ball units, but does not constitute a drawback in this case, since they are easily available in the stock of specialized manufacturers. Since some firms have recently begun to produce a wide range of standardized screw and nut assemblies, this type of feed system is expected to spread in the future. Data concerning a range of hydrostatic screws are to be found in Table 14.3 (ref. 14.8). The nut constitutes a compact unit, with built-in restrictors and seals, a n inlet port and an outlet port, requiring only a n adequate but fairly simple supply system. Feeding accuracy depends mainly on the pitch error of the male screw, but owing to the levelling effect of hydrostatic lubrication the manufacturer claims that actual feeding inaccuracy is less than one third of the pitch fluctuations of the male screw. Lo
T A B L E 14.3 Hydrostatic screw and nuts (ref. 14.8).
63
36.5
108
173 205 167 207
134 166 124 164
D
20
2.79 3.72 2.26 3.39
181 241 147 220
1.733 2.300 0.755 1.133
494
HYDROSTATIC L UBRICATlON
As already noted, the manufacture of hydrostatic nuts is somewhat difficult, either because of the relatively inaccessible position of the recesses, or because a small pitch difference in relation to the male screw can lead to a considerable loss of loading capacity (see section 7.3). Both problems can be easily overcome by means of a clever technique consisting in coating the inner surface of the nut with a thick layer of plastic, which is cast while the lead screw is held in position; recesses are obtained by means of patterns temporarily fixed to the flanks of the screw with an adhesive (note that hydrostatic nuts are in general of the multirecess type rather than of the continuous recesses type). The gap is obtained because of the shrinkage of cast plastic (ref. 14.9).
In large machine tools it may be preferable to substitute the screw-nut feed drive with rack and worm systems, which permit runs of practically any length, with a high degree of stiffness; furthermore, stiffness proves t o be independent from run length and the position of the slide. These systems can also obviously be assisted with hydrostatic lubrication. An example is that of the so-called hydrostatic "Johnson drive" (ref. 14.9)shown in Fig. 14.15.In this case, a short worm drives a long rack firmly fixed to the slide. The
Fig. 14.15 Hydrostatic Johnson drive (Ingersoll). 1-Slide, 2-rack, 3-pump pressure, 4-capillaries, 5-cells, 6-worm, 7-external gear teeth, 8-oil supply for forward flanks, 9-bed.
A PPLICA TIONS
495
worm is supported by means of hydrostatic thrust bearings; its circumference is toothed and is in mesh with a pinion driven by the feed gear. Recesses are hollowed in the flanks of the rack. A simple distributing device is needed to deliver lubricant only to the recesses covered by the worm, hence avoiding a considerable waste of power. In other applications (see, for instance, ref. 7.1 o r ref. 14.11) the rack is fixed to the bed, whereas the worm is supported by the slide, together with the relevant feed gear, which drives it by means of a toothed gear, fitted near the worm on the same shaft. Lubricant is supplied through ducts drilled in the worm; recesses may be hollowed in the flanks of the rack (as in Fig. 14.16)as well a s in the flanks of the male screw.
Fig. 14.16 Hydrostatic rack and worm;diarnete-270 mm,pitch=60 mm (INNSE).
In this case, too, a distributor is needed in order to cut off the high-pressure supply of lubricant to the ducts not ending on the flanks of the rack. When speed is high (speeds up to 750 rpm can be used) the centrifugal force may empty inactive ducts and that may cause aeration of the lubricant: hence the supply distributor should incorporate a pre-filling device whose task is to pump lubricant at low pres-
496
HYDROSTATIC LUBRICA TlON
sure into the inactive ducts, just before they become active again (a similar device is described in ref. 14.12). Hydrostatic worms are generally built with a pitch of between 36 and 60 mm and a n outside diameter of between 150 and 300 mm; load capacity may vary between 50 and 180 KN. The rack may be of virtually any length since it is built in sections (for instance, 1000 mm in length) that are bonded and firmly bolted to the slide bed after having been adjusted in relation to one another and measured to verify the pitch error (ref. 7.1). Accuracy may be about 70i-80 pm on a length of 25 m. Owing to this accuracy and to the very high degree of stiffness this feed system can also be used for monitoring the position of the slide during normal operation (by means of electronic compensation the relevant error can be further reduced to a very small value). 14.2.4
Guideways and rotating tables
Hydrostatic lubrication proves to be particularly suitable for guideways of modern high precision machine tools (especially those equipped with numerical control), because of their intrinsic characteristics: very low friction (and proportional to speed); freedom from stick-slip; freedom from wear (which means constancy of performance for an indefinite time); thickness of the oil film independent of the sliding speed (whereas for lubricated plain bearings it increases with speed); high damping capacity fin directions perpendicular to guide); levelling ability: the fairly high film thickness (commonly a few hundredths of a millimeter) allows the hydrostatic lubrication to compensate, a t least partially, for small geometric inaccuracies and deformations of the guides; possibility of building guides of virtually any length (which is difficult with roller guides). On the other hand, it should be noted that the virtual elimination of friction can enhance the effects of the flexibility of other parts of the machine and in particular of the feed drive. For instance, consider the experimental diagrams in Fig. 14.17 (see ref. 14.13 for further details): they refer to a milling machine and show that the displacement due to loading in the direction of the guides (mainly due to the flexibility of the ball screw and nut and of the relevant thrust bearing) is greatly reduced by the friction of the sliding ways. Diagrams in Fig. 11.18 (ref. 14.71, obtained with a similar experimental rig, show the reduction in damping connected with the use of frictionless guides (either
497
APPLlCATlONS
0
20000
kN
10000
load P
Fig. 14.17 Influence of hydrostatic ways on static stiffness, compared with sliding ways.
a- Hydrostatic system in action; b- without the hydrostatic system.
plain or ball screws were used as feed drives, without leading to any notably different behaviour). Problems of this kind are easily eliminated by means of simple clamping devices when feed rate is null, whereas in other cases they may be solved by stiffening the feed drive (for instance, in heavy machines, by selecting a worm and rack feed drive instead of screw and nut), by eliminating any backlash and increasing damp-
I
0
I
200
I
I
400
I
V
*
mmlmin
Fig. 14.18 Influence of feed rate V on maximum vibration amplitude A, (at resonance frequency) along feed direction for: a- sliding guideways; b- hydrostatic guideways; c- roller guideways.
498
HYDROSTATIC LUBRICATION
ing (for instance, introducing hydrostatic lubrication in the feed drive) or by means of external dampers (ref. 14.13). A number of different examples of layout for slideway guides are presented in Fig. 14.19; type ‘c’ and ‘d‘use an opposed-pad design: this is necessary when great stiffness and damping are required for a large range of loading conditions. The lower pads are in this case much smaller than the upper ones, in order to compensate for the weight of the slide.
-a-
-b-
-c-
-d-
Fig. 14.19 Sample layouts of hydrostatic guideways.
A compromise, often used in rotary tables, may consist i n substituting the preloading effect of the hydrostatic recesses on the underside of the guide with a spring force applied by means of rolling bearings (in practice, this is a trick for increasing the weight of the slide without increasing its mass). At least two recesses must be used on each guide to absorb torque, but a larger number of smaller recesses (each fed independently) provide greater compensating ability for the geometric inaccuracies of guideways; moreover, since the load is more evenly distributed on the guides, better results should also be obtained from the point of view of elastic distortion. Recesses may be either of the conventional fully-hollowed type, or be reduced to narrow grooves, as in the guides i n Fig. 14.20. From the point of view of static load capacity both designs perform i n the same way, but the narrow-groove recesses have greater damping ability and a larger bearing area in the absence of lubrication (hence, they are less prone to damage in the event of failure of the supply system). On the other hand, friction is also much higher and this type of recess proves to be adequate only for low-sliding velocities. The ability of hydrostatic lubrication to even out inaccuracies due to manufacturing errors or deformations caused by external forces is limited by the thickness of the lubricant film. Especially in the case of very large and heavily loaded slides
APPLlCA TlONS
499
Fig. 14.20 Rototraversing table equipped with hydrostatic lubrication of the guides (INNSE). In a the thrust bearing of the rotary table is shown; the pinions of the feed drive are also visible, as well as four clamps that may be used to fix the angular position of the table and the laminar-flow restrictors. In b the same table is shown from another angle: the linear guideways are visible, as well as two clamps.
500
HYDROSTATIC LUBRICATION
and rotary tables (such as the rotary table of a large vertical lathe), elastic deformation might even force the designer to select an excessively thick film to avoid metalto-metal contact. A solution may be to build the guideway with self-aligning tilting pads, as will be shown in section 14.3. The geometric inaccuracies of the slideways (e.g. waviness) might be completely compensated by controlling recess pressure: the principle is outlined i n Fig. 14.21 (see also ref. 14.10). Pressure in each recess is controlled by a valve, piloted by a regulator which compares a reference signal with the signal produced by a transducer. This last is, for instance, a pneumatic sensor which monitors the position of the slide in relation to a reference straight edge, or a photoelectric sensor, which uses a laser beam a s a reference "guide".
1
2
Fig. 14.21 Scheme of compensating bearing control. 1-Guide, 2-reference guide, 3-distance transducer, 4-regulator, 5-set value, 6-controlled valve, 7-supply pressure.
An example of hydrostatic lubrication applied to guideways is presented in Fig. 14.20,in which details are shown of a hydrostatic rototraversing table: one of a wide range of such equipment, suitable for indexing and contour milling (ref. 14.14)with a load capacity varying from 400 to 5000 KN. A similar range of rotary and rototraversing tables is also suitable for turning operations, with a turning speed of up to 2565 rpm, depending on the diameter of the table (2.5+10m).
The rotary table in Fig. 14.20 has a circular thrust bearing (with a mean diameter of 1400 mm) made up of 12 pads, all fed independently through a set of laminarflow restrictors. These are made by cutting small-diameter (111.5mm) pipes to the appropriate length and are also visible in the photographs. With a supply pressure of 6 MPa, the table can bear loads of up to 600 KN. The radial forces are sustained by a tapered roller bearing, which also exerts a preloading force (150KN) on the hydrostatic thrust bearing, in order to increase its stiffness. The photographs also show
APPL ICATlONS
501
clamps that are able to hold the table firmly in any position, without affecting the film thickness of the hydrostatic bearings. Rotary motion is obtained by means of two controlled-preload pinions meshing with a helical crown gear, whereas a ball screw is used for linear axis transmission (the largest members of the same family of tables use hydrostatic worms and racks for axial feed drive). Hydrostatic lubrication is often also applied to the guides of ram-type milling arms (Fig. 14.22). The design of the guides is of course different from that of the guides of horizontal tables: in this case the ram is supported by two rows of eight recesses (two for each side). The recesses in the lower end of the guide are generally larger since they must support higher loads (in other applications there are three rows of recesses, two of which are set at the lower end of the guide). The supply system is made up of a set of multiple pumps (each pump directly feeds one recess),
Fig. 14.22.a- Hydrostatically lubricated milling arm (Pensotti).
502
HYDROSTATC LUBRlCAT/ON
Fig. 14.22.b- Hydrostatically lubricated milling arm:detail showing hydrostatic pads.
which are fed at constant pressure (-2.5MPa) by a larger pump. In this case, too, the recesses of the pads (which are made of bronze) are reduced to narrow grooves.
It is interesting that hydrostatic lubrication has also been used to compensate for the deflection of the ram due to the cutting force. The geometric adaptive control system described in ref. 14.15 measures the displacement of the milling head by means of a laser gun fixed to the milling arm, which emits a laser beam parallel to the undeformed axis of the ram, and a photoelectric scanner attached to the milling head. The signal produced by the measuring equipment is taken a s its input by a control unit which varies accordingly the speed of a servo-motor driving a further set of pumps. The flow produced by these compensating pumps is directed towards the appropriate recesses and added to the normal flow in order to produce a displacement of the milling head, realigning it with the laser beam.
A P P l ICA TlONS
503
A particular application of externally pressurized lubrication to the ram guide of a gear-shaping machine is described in ref. 14.10. A cross section of the guide is shown in Fig. 14.23: the ram is shaped like a spur gear with every third tooth removed. The accuracy of the internal bore of the sleeve is obtained by casting with a plastic material (this technique is briefly described in section 14.2.3).
Fig. 14.23 Hydrostatic ram guide of a gear-shaping machine (Liebherr). 14.3
LARGE TILTING PADS
Hydrodynamic bearings for very large rotating machine-members have been equipped for many years now with high-pressure hydrostatic pockets, used as jacking devices at starting (hydrostatic lifts). More recently, i t has been found to be expedient to retain the hydrostatic effect in normal running and then to substitute the hydrodynamic bearings completely with hydrostatic (or hybrid) bearings, in the case of slowly rotating machines in particular, or when irregularities in load or speed are expected. One problem connected with this type of bearing in certain machines (such as ore mills) is that the elastic deformation of the runner, due to the pressure of the lubricant, may greatly reduce the effectiveness of hydrostatic lubrication (Fig. 11.24.a). This problem may be overcome by foregoing the "optimum" design, obtained by assuming rigid surfaces and uniform film thickness, and displacing the recesses from the centre of the bearing (Fig. 11.24.b); separate pads may even be used instead of a multirecess bearing (ref. 14.16).
A further improvement in design, able to eliminate most of the problems connected with elastic deformation, machining tolerance, thermal expansion and so
504
HYDROSTATIC LUBRICA TlON
Fig. 14.24 Trunnion deformation due to bearing pressure: a- bearing as designed; b- improved concept; e- most effective concept (ref. 14.16).
on, consists in supporting the large journal by means of a set of self-aligning hydrostatic shoes, as shown in Fig. 14.25 (ref. 14.17). Each shoe is split up into two parts: the upper part rests on a spherical seat and hence can tilt in all directions. The underside of the upper part is shaped like a piston which fits into a cylinder in the base: since the piston area, on which the recess pressure acts, is slightly smaller than the effective area of the pad, the load on the spherical seat is quite low during normal operation.
Fig. 14.25 Arrangement of tilting-pad hydrostatic bearings (ref. 14.17).
505
APPLICATIONS
The spherical rest of each inner shoe (slave shoe) is pushed against the runner by a further piston on which, thanks to a hydraulic connection, the recess pressure of the relevant outer shoe (master shoe) acts. Clearly, if the sum of the two piston areas equals the effective pad area, the slave shoe must necessarily have the same film thickness, and thus the same recess pressure, a s the relevant master shoe (each pad is fed by the same flow rate). Thus when the load direction is vertical all four shoes have the same film thickness and recess pressure, regardless of the deviation of the runner from the ideal circular shape. When the load deviates from the vertical direction the two shoes on each side have an equal part of the load component falling along the line between the two shoes (ref. 14.17). The shape of the recess is also of particular interest. It is known that when a cylindrical pad with a simple recess (as in Fig. 5.30) is tilted from the concentric configuration the pressure field on the land surface is altered and produces a moment that tends to realign the pad; however, this self-aligning capacity is too small to ensure the stability of the shoe in all conditions and i t is hence necessary to use multirecess pads. In Fig. 14.26 the main recess is surrounded by four auxiliary recesses, situated in the corners of the pad, which are fed with the lubricant which passes from the central recess over the bearing lands and through small drilled ducts (this is a compromise aimed a t avoiding dependence upon the direction of rotation: for the greatest stability the auxiliary recesses on the trailing side should only be supplied over the lands). The hydrostatic system described in ref. 14.17 supported a large tube mill for crushing ores: each bearing runner had a diameter of 2700 mm, the maximum
I -.
I
I
L Fig. 14.26 Improved recess pattern (ref. 14.17).
-
W
506
HYDROSTAT C LUBRlCAT/ON
load was 3500 KN and the velocity was 0.24 reds. Each shoe was 640 mm long and 500 mm wide and was fed at 25 Ym with a lubricant whose viscosity was 0.1 Ns/m2 at 50°C. Film thickness in normal operation was 0.1410.15 111111.
A range of hydrostatic shoes based on the foregoing working principles is currently produced by the same firm (ref. 14.5): a sketch of them is to be found in Fig. 14.27 and their main dimensions are given in table 14.4. The recess pattern is similar t o that shown in Fig. 14.26, but the main recess is now annular in shape, in order t o increase the bearing area a t rest (in the absence of hydrostatic lubrication) virtually without affecting bearing performance during normal operation. Hydrostatic shoes may be used to support horizontal as well as vertical rotating equipment. In the first case the rotating drum may lean on the shoes by means of trunnions (Fig. 14.28.a) or by means of girth rings (Fig. 14.28.b). The latter arrangement, which is often inapplicable with rolling bearings due to their size limits, permits large feed openings and a simplified (and less expensive) design. Each -b-
-a-
Fig. 14.27 Bearing shoes: a- master shoe; b- slave shoe. T A B L E 14.4 H
410 SO0 600
530 640 756
(mm)
Master
Slave
300
180+190 2601270 2951305 320+330 4201430
425
507
APPLlCA TlONS
ring (or trunnion) is supported by two master shoes, to each of which one or two slave shoes may be added to boost the load-carrying capacity (Fig. 14.29). The suggested ring diameter D varies between 500 and 5400 mm, with a load capacity F ranging from 480 to 12000 KN, depending on pad size and the total number of pads. -a-
-b-
Fig. 14.28 Horizontal rotating arrangements: a- trunnion arrangement; b- girth ring arrangement.
D
IMaster
shoe: O S l a v e shoe
Fig. 14.29 Shoe arrangements for horizontal rotating cylinders.
In the case of vertical equipment three master shoes are obviously required in order t o obtain a statically determined load distribution; to each master shoe a slave unit can be added, thereby doubling the load capacity. A typical arrangement is shown in Fig. 14.30, in which two alternatives are also proposed for the radial guidance of the runner: a rolling bearing mounted on the shaft, or a set of hydrostatic guiding pads (see below). For six pad arrangements the load capacity ranges from 1900 t o 14000 KN (depending on the pad size) and correspondingly the minimum pitch diameter D varies from 800 to 2400 mm. Besides the hydrostatic shoes described above, the same firm produces a range of smaller tilting pads of simplified design (see Fig. 14.31 and table 14.5). These still retain a self-aligning capacity, since they have a spherical seat and multiple recesses, but are not equipped with hydraulic cylinders. They are mainly proposed (ref. 14.5) as guiding pads for the axial location of a girth ring (Fig. 14.32) or for the
508
HYDROSTATIC LUBRICATION
D
r=
ALT I
I
ALT II
Fig. 14.30 Shoe bearing arrangement for vertical rotating equipment. radial guiding of platforms (Fig. 14.30). Compact assemblies are also available consisting in a master shoe bearing with two guiding pads (in an opposed-pad configuration) mounted on the fixed part of the shoe (Fig. 14.32.b). Another type of tilting pad (ref. 14.18) can be used to build spherical thrust bearings with a very large diameter. In practical terms, it consists of a circular recess pad laid on a spherical rest whose position can be adjusted by means of a wedge. In Fig. 14.33 a set of twenty pads is used to build a large bearing (with a mean diameter of 5000 mm) for a large parabolic antenna. The dimensions of the bearing and angle a depend on the value of the axial and radial components of the load: bearings with an external diameter of up to 8000 mm can be built.
APPLlCATlONS
509
Fig. 14.31 Guiding pad.
TABLE 14.5 Dimensions of guiding pads (ref. 14.5).
Fig. 14.32 Axial guiding pads: a- separate axial guidance; b- axial guidance integral with a master shoe.
A further type of tilting shoe is shown in Fig. 14.34(ref. 14.19):it can tilt around the cylindrical rib on the underside and align itself thanks to the multiple recesses (two or four) which are fed independently through capillary restrictors. These pads can also be used to sustain radial loads as well as the axial thrust of a large rotating platform. In the latter case, bearings with diameters exceeding 5000 mm may be built, which sustain thrusts greater than 5000 KN and rotating a t more than 20 rpm. These pads prove to be particularly suitable for building rotary tables for large machine tools (e.g. for vertical lathes): an example is given in Fig. 14.35.A different application is described in ref. 14.20,concerning the supporting ring of a 3.5 m telescope.
510
HYDROSTATIC LUBRICATION
P I
Fig. 14.33 Spherical pad arrangement,
Fig. 14.34 "Hydro-tilt"shoe bearing (ref. 14.19).
APPLlCA TlONS
51 1
Fig. 14.35 "Hydro-tilt" shoe arrangement.
Lastly, Fig. 14.36 shows a large-size spherical bearing (ref. 2.2); it has three recesses fed a t a constant flow rate. Bearings like this can sustain heavy loads (up to 10,000 KN) and in general their rotating speed is low. For instance, the bearing depicted in Fig. 14.36 was made to support the rotor of an air preheater weighing 800 KN and rotating at 2 rpm.
14.4
OTHER APPLICATIONS
Apart from those quoted in the foregoing sections, hydrostatic lubrication has a number of different applications. For instance, let us consider the pump in Fig. 14.37 (ref. 14.21): the pistons (1) lean on the tilted plate (2) by means of the spherical pads (3) which are hydrostatically borne by the same circulating fluid. Another special application is quoted in ref. 14.22, that is the lower journal bearing of the main pump of the Super-PhBnix nuclear power plant. This bearing has a
51 2
HYDROSTATIC LUBRICATlON
L.L.1
r't'i
Fig. 14.36 Large spherical bearing.
Fig. 14.37 Piston pump.
diameter of 0.85 m and a width of 0.3 m; it has twelwe recesses carved in the shaft. In this case, too, the lubricant is the fluid circulating in the plant, i.e. liquid sodium. Hydrostatic lubrication has been successfully used in a number of testing rigs. An example is shown in Fig. 14.38:an experimental rig for testing rolling bearings (ref. 14.23).The bearing being tested (1) is made to rotate by a motor (2)by means of a
APPLICATIONS
513
Fig. 14.38 Experimental rig for rolling bearings.
belt drive (3) and is loaded by a jack (4) through the hydrostatic bearing (5);this last leans on a cell (61, which measures load Fa, by means of a spherical seat. Friction moment M R is measured by means of dynamometer (7)and the angular speed n by means of thacheometer (8). 14.5 14.5.1
HYDRAULIC CIRCUITS
Simple layout
A typical supply system for hydrostatic spindles, such as the one in Fig. 2.24, is shown in Fig. 14.39 (ref. 2.2). The bearings are fed at a constant pressure, which is usually in a range between 3 and 7 MPa. A gear pump supplies lubricant at a rate which is 30% greater than the calculated value: the surplus flows back to the reservoir through the pressure regulating valve. Lubricant is pushed through two filters, the first of which is coarser (15 pm), while the other is narrower (5-10pm). A pressure switch prevents the spindle from running until the pressure reaches the established value and stops it when pressure falls: in the last case, an oil accumulator
51 4
HYDROSTATIC LUBRICATION
Fig. 14.39 Supply system for a hydrostatic spindle: 1-oil tank; 2-pump; 3-motor; 4-pressure regulating valve; 5-pressure filter; 6-pressure switch; 7-check valve; 8-piston accumulator; 9-pressure gauge; 10-cooler; 11-thermostaticsystem; 12-heater.
1
2
Fig. 14.40 Supply system for the hydrostatic bearing of an air preheater: 1-pump; 2-motor; 3-pressure filter; 4-pressure switch; 5-check valve; 6-flow divider; 7-piston accumulator; 8-pressure-limiting valve; 9-cooler.
515
APPLICA TlONS
keeps on feeding the bearings for the time needed for the spindle to come to a complete stop. A thermostatic system keeps the temperature of the lubricant close to the design value. Sometimes a further pump may also be needed (generally inserted upstream from the cooler) to bring the lubricant back from the spindle to the reservoir. Flow dividers
14.5.2
Figure 14.40 shows the supply circuit for the three-recess preheater bearing in Fig. 14.36.The flow rate produced by the main pump is divided into three equal streams by means of a flow divider made up of three equal gear pumps mounted on
8
1
1
13
9
\
5
6 4
3
2
17
15
16
14
1
Fig. 14.41 Supply system for the hydrostatic bearing of an ore mill: 1-oil tank; 2-pump; 3-pressure switch; 4-pressure filter; 5-check valve; 6-pressure limiting valve; 7-pressure gauge; 8-piston accumulator; 9-nitrogen gas bottle; 10-flow divider; 11-shoe bearing; 12-circulation pump for cooling circuit 13-throttle valve; 14-oil cooler; 15-water flow control valve; 16-temperature-sensing device.
51 6
HYDROSTATIC LUBRICATlON
a common shaft. To ensure continuous operation a second pump is ready to be started up automatically when supply pressure drops below a safe value. A further spare pump is available for replacement, to permit maintenance operations. In the case of a n electric mains failure a diesel generator can provide power for the motors of the pumps. The last emergency device is a set of oil accumulators which can supply lubricant to the bearings for a short time. The hydraulic circuit for the bearing arrangement in Fig. 14.25 is shown in Fig. 14.41. The flow rate produced by the main pump is divided into four equal streams by means of a flow divider. To ensure continuous operation a second pump is ready to be started up automatically in case of failure of the other one and a set of piston accumulators (driven by pressurized nitrogen bottles) can feed oil to the bearings for a certain time in case of power failure, allowing the runner to stop without damaging the bearings. 14.5.3
Multiple pumps
The constant-flow supply circuit of the guideway presented in Fig. 14.22.b is shown in Fig. 14.42. The pre-feeding pump (1)delivers lubricant at a pressure of 25 bar to two multiple pumps (2). Each pump can feed ten recesses independently, each with a 0.33 m3/s flow rate, a t a pressure of 40 bar.
Fig. 14.42 Supply system, with multiple pumps, of a hydrostatic slide.
A PPLlCATlONS
517
Figure 14.43 shows the supply circuit of the pad arrangement in Fig. 14.33. Each pad is directly fed a t a constant flow rate; that is, a set of five multiple pumps is used and each pump delivers four equal streams which are supplied to four pads situated a t 90 degrees from each other. Thanks to the layout mentioned, emergency operation of bearing system is possible even if a pump fails.
Ic
Fig. 14.43 Supply system for the bearing system of a large-beam antenna: M-motor; P-multiple pump; 1-20 pads.
REFERENCES
- Herzstuck Leistungsfahiger Werkzeugmaschinen; FAG publ. 02-113DA (1985);68 pp. Hallstedt G.;Standardized Hydrostatic Bearing Units; Instn. Mech. Engrs., C48 (1971),420-430. Lewinschal L.; Contributo dei Cuscinetti Zdrostatici allXumento d i Produttivitb delle Rettificatrici; La Rivista dei Cuscinetti/SKF, 196 (19781,24-27. FAG Spindeleinheiten fur das Bohren-Drehen-Frasen; FAG publ. 02-1OW3 DA (1985);12 pp. Hydrostatic Shoe Bearing Arrangements; SKF Publication 3873 E (19881,28 PP. Bi l dt sh C., Htillnor G.; Problema Risolto con 1'Adozione di Pattini Idrostatici;La Rivista dei CuscinettYSKF, 181 (1974),18-20. Polseck M.,Vavra Z.; The influence of different types of guideways on the static and dynamic behaviour of feed drives; Proc. 8th Int. MTDR Conf. (19671, pt. 2, p. 1127-1138.
14.1 Die Arbeitsspindel und Hire Lagerung 14.2 14.3 14.4 14.6 14.6 14.7
518
HYDRCSTATIC LUBRICATlON
14.8 Catalog B1025E; Nachi Corp., Japan, 1984; 4 p. 14.9 Weck M.; Handbook of Machine Tools, Volume 2 (Construction and Mathe-
matical Analysis); J. Wiley & Sons, 1980; 296 pp.
14.10 Rohs H. G.; Die Hydostatische Bewegungspaarung i m Werkzeugmaschinenbau; Konstrudion, 22 (1970); 321-329. 14.11 Andreolli C.; Eliminazione dell'Attrito e dei Giochi nelle Macchine Utensili;
Controlli NumericiIMacchine a CN/Robot Industriali, anno XI1 (19791, n. 7, p. 32-45. 14.12 Appoggetti P.; Perfezionamento negli Accoppiamenti Vite-Cremagliera a Sostentamento Zdrostatico; Patent IT 51829 N69; Bollettino Tecnico RTM n. 9, 1969; p. 47-51. 14.13 Umbach R., Haferkorn W.; Some Examples and Problems in Zmplementation of Mwlern Design Features on Large Size Machine Tools; 10th Int. MTDR cod., Manchester, 1969; paper 34; 30 pp. 14.14 Rototraversing Tables for Indexing, Milling and Turning; INNSE Publication DMU/27 (1985),4pp. 14.15 Weck M.; Handbook of Machine Tools, Volume 3 (Automation and Controls); J . Wiley & Sons, 1980; 451 pp. 14.16 Rippel T., Hunt J. B.; Design and Operational Experience of 102-Znch Diameter Hydrostatic Journal Bearings for Large Size Tumbling Mills; Instn. Mech. Engrs., C16 (1971), 76-100. 14.17 Arsenius H. C., Goran R.; The Design and Operational Experience of a SelfAdjusting Hydrostatic Shoe Bearing for Large Size Runners; Instn. Mech. EWS., C303 (19731,361-367. 14.18 Supporti idrostatici FAG; FAG Publication 44109 IB (19711, 8 pp. 14.19 Andreolli C.; Guida Circolare Idrostatica Assiale per Tavola Portapezzo Rotante; Patent IT 2353CA, 1975; 15 pp. 14.20 Andreolli C.; Sopporto Zdrostatico per 1'Asse Azimutale del 3.5 m New Technology Telescope (NTT) dell %SO; Convegno AIM-AMME (Tribologia-Attrito, Usura e Lubrificazione), Sorrento, 1987; p. 421-430. 14.21 Giordano M., Boudet M.; Thermohydrodynamic Flow of a Piezoviscous Fluid Between Two Parallel Discs; J. Mech. Eng., 1980. 14.22 F r h e J., Nicolas D., Deguerce B., Berthe D., Godet M.; Lubrification Hydrodynamique; Edition Eyrolles, Paris, 1990; 488 pp. 14.23 Martin F. J.; Prove Funzionali e di Qualificazione nello Sviluppo dei Cuscinetti Volventi; La Rivista dei CuscinettiBKF, 224 (1986),28-36.