Wednesday, September 10, 2008

Alignment and Balancing


For most rotating machines used in the process industries, the trend is toward higher speeds, higher horsepowers per machine, and less sparing. The first of these factors increases the need for precise balancing and alignment. This is necessary to minimize vibration and premature wear of bearings, couplings, and shaft seals. The latter two factors increase the economic importance of high machine reliability, which is directly dependent on minimizing premature wear and breakdown of key components.

Balancing, deservedly, has long received attention from machinery manufacturers and users as a way to minimize vibration and wear. Many shop and field balancing machines, instruments, and methods have become available over the years. Alignment, which is equally important, has received proportionately less notice than its importance justifies. Any kind of alignment, even straightedge alignment, is better than no alignment at all. Precise, two-indicator alignment is better than rough alignment, particularly for machines 3600 RPM and higher. It can give greatly improved bearing and seal life, lower vibration, and better overall reliability. It does take longer, however, especially the first time it is done to a particular machine, or when done by inexperienced personnel. The process operators and mechanical supervisors must be made aware of this time requirement. If they insist on having the job done in a hurry. they should do so with full knowledge of the likelihood of poor alignment and reduced machine reliability. Figure 5-1 shows a serious machinery failure which started with piping-induced misalignment, progressed to coupling distress, bearing failure, and finally, total wreck.





Prealignment Requirements

The most important requirement is to have someone who knows what he is doing, and cares enough to do it right. Continuity is another important factor. Even with good people, frequent movement from location to location can cause neglect of things such as tooling completeness and prealignment requirements.

The saying that “you can’t make a silk purse out of a sow’s ear” also applies to machinery alignment. Before undertaking an alignment job, it is prudent to check for other deficiencies which would largely nullify the benefits or prevent the attainment and retention of good alignment. Here is a list of such items and questions to ask oneself: Foundation Adequate size and good condition? A rule of thumb calls for concrete weight equal to three times machine weight for rotating machines, and five times for reciprocating machines. Grout Suitable material, good condition, with no voids remaining beneath baseplate? Tapping with a small hammer can detect hollow spots, which can then be filled by epoxy injection or other means. This is a lot of trouble, though, and often is not necessary if the lack of grout is not causing vibration or alignment drift.

Baseplate Designed for adequate rigidity? Machine mounting pads level, flat, parallel, coplanar, clean? Check with straightedge and feeler gauge. Do this upon receipt of new pumps, to make shop correction possible-and maybe collect the cost from the pump manufacturer. Shims clean, of adequate thickness, and of corrosion- and crush-resistant material? If commercial pre-cut shims are used, check for actual versus marked thicknesses to avoid a soft foot condition. Machine hold-down bolts of adequate size, with clearance to permit alignment corrective movement‘? Pad height leaving at least 2 in. jacking clearance beneath center at each end of machine element to be adjusted for alignment? If jackscrews are required, are they mounted with legs sufficiently rigid to avoid deflection? Are they made of type 3 16 stainless steel, or other suitable material, to resist field corrosion? Water or oil cooled or heated pedestals are usually unnecessary, but can in some cases be used for onstream alignment thermal compensation.




Piping Is connecting piping well fitted and supported, and sufficiently flexible, so that no more than 0.003 in. vertical and horizontal (measured separately -not total) movement occurs at the flexible coupling when the last pipe flanges are tightened? Selective flange bolt tightening may be required, while watching indicators at the coupling. If pipe flange angular misalignment exists, a “dutchman” or tapered filler piece may be necessary. To determine filler piece dimensions,



measure flange gap around circumference, then calculate as follows:

Gasket 0. D . I Flange O.D. I = l/8 in. + (Max. Gap - Min. Gap) Maximum Thickness of Tapered Filler Piece1/8 in. = Dutchman Minimum Thickness (180’ from Maximum Thickness). Dutchman OD and ID same as gasket OD and ID.

Spiral wound gaskets may be helpful, in addition to or instead of a tapered filler piece. Excessive parallel offset at the machine flange connection cannot be cured with a filler piece. It may be possible to absorb it by offsetting several successive joints slightly, taking advantage of clearance between flange bolts and their holes. If excessive offset remains, the piping should be bent to achieve better fit. For the “stationary” machine

element, the piping may be connected either before or after the alignment is done-provided the foregoing precautions are taken, and final alignment remains within acceptable tolerances. In some cases, pipe expansion or movement may cause machine movement leading to misalignment and increased vibration. Better pipe supports or stabilizers may be needed in such situations. At times it may be necessary to adjust these components with the machine running, thus aligning the machine to get minimum vibration. Sometimes, changing to a more tolerant type of coupling, such as elastomeric, may help.

Coupling Some authorities recommend installation on typical pumps Installation and drivers with an interference fit, up to .OW5 in. per in. of shaft diameter. In our experience, this can give problems in subsequent removal or axial adjustment. If an interference fit is to be used, we prefer a light one-say .OOO3 in. to .OOO5 in. overall, regardless of diameter. For the majority of machines operating at 3600 RPM and below, you can install couplings with .0005 in. overall diametral clearance, using a setscrew over the keyway. For hydraulic dilation couplings and other nonpump or special categories, see manufacturers’ recommendations or appropriate section of this text. Many times, high-performance couplings require interference fits as high as .0025 in. per in. of shaft diameter.

Coupling cleanliness, and for some types, lubrication, are important and should be considered. Sending a repaired machine to the field with its lubricated coupling-half unprotected, invites lubricant contamination, rusting, dirt accumulation, and premature failure. Lubricant should be chosen from among those recommended by the coupling manufacturer or a reputable oil company. Continuous running beyond two years is inadvisable without inspecting a grease lubricated coupling, since the centrifuging effects are likely to cause caking and loss of lubricity. Certain lubricants, e.g., Amoco and Koppers coupling greases, are reported to eliminate this problem, but visual external inspection is still advisable to detect leakage. Continuous lube couplings are subject to similar problems, although such remedies as anti-sludge holes can be used to allow longer runs at higher speeds. By far the best remedy is clean oil, because even small amounts of water will promote sludge formation. Spacer length can be important, since parallel misalignment accommodation is directly proportional to such length.


Alignment Tolerances

Before doing an alignment job, we must have tolerances to work toward. Otherwise, we will not know when to stop. One type of “tolerance” makes time the determining factor, especially on a machine that is critical to plant operation, perhaps the only one of its kind. The operations superintendent may only be interested in getting the machine back on the line,fast. If his influence is sufficient, the job may be hurried and done to rather loose alignment tolerances. This can be unfortunate, since it may cause excessive vibration, premature wear, and early failure. This gets us back to the need for having the tools and knowledge for doing a good alignment job efficiently. So much for the propaganda-now for the tolerances.

Tolerances must be established before alignment, in order to know when to stop. Various tolerance bases exist. One authority recommends ‘h-mil maximum centerline offset per in. of coupling length, for hot running misalignment. A number of manufacturers have graphs which recommend tolerances based on coupling span and speed. A common tolerance in terms of face-and-rim measurements is .003-in. allowable face gap difference and centerline offset. This ignores the resulting accuracy variation due to face diameter and spacer length differences, but works adequately for many machines.

Be cautious in using alignment tolerances given by coupling manufacturers. These are sometimes rather liberal and, while perhaps true for the coupling itself, may be excessive for the coupled machinery. A better guideline is illustrated in Figure 5-2, which shows an upper, absolute misalignment limit, and a lower, “don’t exceed for good longterm operation limit.” The real criterion is the running vibration. If excessive, particularly at twice running frequency and axially, further alignment improvement is probably required. Analysis of failed components such as bearings, couplings, and seals can also indicate the need for improved alignment.




Figure 5-2 can be applied to determine allowable misalignment for machinery equipped with nonlubricated metal disc and diaphragm couplings, up to perhaps 10,000 rpm. If the machinery is furnished with gear-type couplings, Figure 5-2 should be used up to 3,600 rpm only. At speeds higher than 3,600 rpm, gear couplings will tolerate with impunity only those shaft misalignments which limit the sliding velocity of engaging gear teeth to less than perhaps 120 in. per minute. For gear couplings, this velocity can be approximated by V = (TDN) tancr, where






resimmmmmm Say, for example, we were dealing with a 3560 rpm pump coupled to a motor driven via a 6-in. pitch diameter gear coupling. We observe a total indicator reading of 26 mils in the vertical plane and a total indicator reading of 12 mils in the horizontal plane. The distance between the flexing member of the coupling, i.e., flexing member on driver and flexing member on driven machine, is 10 in. The total net indicator reading is [(26)2 + (12)?]''* = 28.6 mils. Tan CY = (1/2)(28.6)/10) = 1.43 milslin., or 0.00143 in./in. The sliding velocity is therefore [(~)(6)(3560) (0.00143)1 = 96 in. per minute. Since this is below the maximum allowable sliding velocity of 120 in. per minute, the installation would be within allowable misalignment.

Choosing an Alignment Measurement Setup


Having taken care of the preliminaries, we are now ready to choose an alignment setup, or arrangement of measuring instruments. Many such setups are possible, generally falling into three broad categories: faceand- rim, reverse-indicator, and face-face-distance. The following sketches show several of the more common setups, numbered arbitrarily for ease of future reference. Note that if measurements are taken with calipers or ID micrometers, it may be necessary to reverse the sign from that which would apply if dial indicators are used. Figures 5-3 through 5-8 show several common arrangements of indicators, jigs, etc. Other arrangements are also possible. For example, Figures 5-3 and 5-4 can be done with jigs, either with or without breaking the coupling. They can also sometimes be done when no spacer is present, by using right-angle indicator extension tips. Figures 5-6 and 5-7 can be set up with both extension arms and indicators on the same side,
          

                                             





rather than 180" opposite as shown. In such cases, however, a sign reversal will occur in the calculations. Also, we can indicate on back of face, as for connected metal disc couplings. Again, a sign reversal will occur. In choosing the setup to use, personal preference and custom will naturally influence the decision, but here are some basic guidelines to follow.

Reverse-Indicator Method

This is the setup we prefer for most alignment work. As illustrated in Figure 5-9, it has several advantages:

1. Accuracy is not affected by axial movement of shafts in sleeve bearings.

2. Both shafts turn together, either coupled or with match marks, so coupling eccentricity and surface irregularities do not reduce accuracy of alignment readings.

3. Face alignment, if desired, can be derived quite easily without direct measurement.




4. Rim measurements are easy to calibrate for bracket sag. Face sag, by contrast, is considerably more complex to measure.

5. Geometric accuracy is usually better with reverse-indicator method in process plants, where most couplings have spacers.

6. With suitable clamp-on jigs, the reverse-indicator method can be used quite easily for measuring without disconnecting the coupling or removing its spacer. This saves time, and for gear couplings, reduces the chance for lubricant contamination.

7. For the more complex alignment situations, where thermal growth and/or multi-element trains are involved, reverse-indicator can be used quite readily to draw graphical plots showing alignment conditions and moves. It is also useful for calculating optimum moves of two or more machine elements, when physical limits do not allow full correction to be made by moving a single element.

8. When used with jigs and posts, single-axis leveling is sufficient for ball-bearing machines, and two-axis leveling will suffice for sleeve-bearing machines.

9. For long spans, adjustable clamp-on jigs are available for reverseindicator application, without requiring coupling spacer removal. Face-and-rim jigs for long spans, by contrast, are usually nonadjustable custom brackets requiring spacer removal to permit face mounting.

10. With the reverse-indicator setup, we mount only one indicator per bracket, thus reducing sag as compared to face-and-rim, which mounts two indicators per bracket. (Face-and-rim can do it with one per bracket if we use two brackets, or if we remount indicators and rotate a second time, but this is more trouble.)

There are some limitations of the reverse-indicator method. It should not be used on close-coupled installations, unless jigs can be attached behind the couplings to extend the span to 3 in. or more. Failure to observe this limitation will usually result in calculated moves which overcorrect for the misalignment.

Both coupled shafts must be rotatable, preferably by hand, and preferably while coupled together. If only one shaft can be rotated, or if neither can be rotated, the reverse-indicator method cannot be used. If the coupling diameter exceeds available axial measurement span, geometric accuracy will be poorer with reverse-indicator than with faceand- rim.

If required span exceeds jig span capability, either get a bigger jig or change to a different measurement setup such as face-face-distance. Cooling tower drives would be an example of this.

Face-and-Rim Method

This is the “traditional” setup which is probably the most popular, al- Advantages of face-and-rim:

1. It can be used on large, heavy machines whose shafts cannot be

2. It has better geometric accuracy than reverse-indicator, for large di-

3. It is easier to apply on short-span and small machines than is rethough it is losing favor as more people learn about reverse-indicator. turned. ameter couplings with short spans.

verse-indicator, and will often give better accuracy. Limitations of face-and-rim: 1. If used on a machine in which one or both shafts cannot be turned, some runout error may occur, due to shaft or coupling eccentricity.

If used on a sleeve bearing machine, axial float error may occur. One method of avoiding this is to bump the turned shaft against the axial stop each time before reading. Another way is to use a second face indicator 180" around from the first, and take half the algebraic difference of the two face readings after 180" rotation from zero start. Figure 5-10 illustrates this alignment method. Two 2411. tubular graphite jigs are used for light weight and high rigidity. If used with jigs and posts, two or three axis leveling is required, for ball and sleeve bearing machines respectively. Reverse-indicator requires leveling in one less axis for each.

Face-and-rim has lower geometric accuracy than reverse-indicator, for spans exceeding coupling or jig diameter.

Face sag is often insignificant, but it can occur on some setups, and result in errors if not accounted for. Calibration for face sag is considerably more complex than for rim sag.

For long spans, face-and-rim jigs are usually custom-built brackets requiring spacer removal to permit face mounting. Long-span reverse- indicator jigs, by contrast, are available in adjustable clampon models not requiring spacer removal.

Graphing the results of face-and-rim measurements is more complex than with reverse-indicator measurements.



Face-Face-Distance Method

Advantages of face-face-distance:

1. It is usable on long spans, such as cooling tower drives, without elaborate long-span brackets or consideration of bracket sag.

2. It is the basis for thermal growth measurement in the Indikon proximity probe system, and again is unaffected by long axial spans.

3. It is sometimes a convenient method for use with diaphragm couplings such as Bendix and Koppers, allowing mounting of indicator holders on spacer tube, with indicator contact points on diaphragm covers.

Limitations of face-face-distance:

1. It has no advantage over the other methods for anything except long spans.

2. It cannot be used for installations where no coupling spacer is present.

3. Its geometric accuracy will normally be lower than either of the other two methods.

4. It may or may not be affected by axial shaft movement in sleeve bearings, but this can be avoided by the same techniques as for face-and-rim.

Laser-Optic Allgnment

In the early 1980’s, by means of earth-bound laser beams and a reflector mounted on the moon, man has determined the distance between earth and the moon to within about 6 inches. Such accuracy is a feature of optical measurement systems, as light travels through space in straight lines, and a bundled laser ray with particular precision.

Thus, critical machinery alignment, where accuracy of measurement is of paramount importance, is an ideal application for a laser-optic alignment system.

The inherent problems of mechanical procedure and sequence of measuring have been solved by Prueftechnik Dieter Busch, of 8045 Ismaning (West Germany), whose OPTALIGN@ system comprises a semiconductor laser emitting a beam in the infrared range (wavelength 820 mm), along with a beamfinder incorporating an infrared detector. The laser beam is refracted through a prism and is caught by a receiver/detector.

These light-weight, nonbulky devices are mounted on the equipmcnt shafts, and only a cord-connected microcomputer module is external to the beam emission and receiverldetector devices. The prism redirects the beam and allows measurement of parallel offset in one plane and angularity in another, thus simultaneously controlling both. In one 360" rotation of the shafts all four directional alignment corrections are determined.

The receiver is a biaxial analog photoelectric semiconductor position detector, yielding mathematical results to within one micron. Data for computation are entered automatically through a cable direct from the receiver/ detector. The only information still to be entered manually is the relative position, 4 times, at 0, 90, 180 and 270".

With the data automatically obtained from the receiver/detector, the microcomputer instantaneously yields the horizontal and vertical adjustment results for the alignment of the machine to be moved.

Physical contact between measuring points on both shafts is no longer required, as this is now bridged by the laser beam, eliminating the possibilities for error arising from gravitational hardware sag as well as from sticky dial indicators, etc. The system's basic attachment is still carried out with a standard quick-fit bracketing system, or with any other suitable attachment hardware.

If the reader owns an OPTALIGN@ system, he does not have to be concerned with sag. Other readers must continue the checkout process.

Checking for Bracket Sag

Long spans between coupling halves may cause the dial indicator fixture to sag measurably because of the weight of the fixture and the dial indicators. Although sag may be minimized by proper bracing, sag effects should still be considered in vertical alignment, To determine sag, install the dial indicators on the alignment fixture in the same orientation and relative position as in the actual alignment procedure with the fixture resting on a level surface as shown in Figure 5-1 I. With a small sling and scale, lift the indicator end of the fixture so that the fixture is in the horizontal position. Note the reading on the scale. Assume for example that the scale reading was 7.5 Ibs. Next, mount the alignment fixture on the coupling hub with the dial indicator plunger touching the top vertical rim of the opposite coupling hub. Set the dial indicator to zero. Next, locate the sling in the same relative position as before and, while observing the scale, apply an upward force so as to repeat the previous scale reading (assumed 7.5 Ibs in our example). Note the dial indicator reading while holding the upward force.





Let us assume for example that we observe a dial indicator reading of - 0.004 in. Using this specific methodology, sag error applies equally to the top and bottom readings. Therefore, the sag correction to the total indicator reading is double the indicated sag and must be algebraically subtracted from the bottom vertical parallel reading, i.e., - (2) ( - .004) = + .008 correction to bottom reading. This method is a clever one for face-mounted brackets. For clamp-on brackets, however, it would be easier and more common to attach them to a horizontal pipe on sawhorses, and roll top to bottom. Figure 5-12 shows this conventional method which, except for the sag compensator device, is almost universally employed. The sag compensator feature incorporates a weight-beam scale which applies an upward force when the indicator bracket is located at the top of the machine shaft, and an equal, but opposite, force when the indicator bracket and shaft combination is rotated to the down position, 180" removed.

In any event, let us assume that we obtain readings of 0 and + 0.160 in. at the top and bottom vertical parallels respectively. We correct for sag in the following manner:


Bracket Sag Effect on Face Measurements

Bracket sag is generally thought to primarily affect rim readings, with little effect on face readings. Often this is true, but some risk may be incurred by assuming this without a test. Unlike rim sag, face sag effect depends not only on jig or bracket stiffness, but on its geometry. Determining face sag effect is fairly easy. First get rim sag for span to be used (we are referring here to the full indicator deflection due to sag when the setup is rotated from top to bottom). This may be obtained by trial, with rim indicator only, or from a graph of sags compiled for the bracket to be used. Then install a setup with rim indicator only, on Calibration pipe or on actual field machine, and “lay on” the face indicator and accessories, noting additional rim indicator deflection when this is done. Double this additional deflection, and add it to the rim sag found previously, if both the face and rim indicators are to be used simultaneously. If the face and the rim indicators are to be used separately, to reduce sag, use the original rim sag in the normal manner, and use this same original rim sag as shortly to be described in determining face sag effect-in this latter case utilizing a rim indicator installed temporarily with the face indicator for this purpose. If the face indicator is a different type (Le., different weight) from the rim indicator, obtain rim sag using this face indicator on the rim, and use this figure to determine face sag effect.

Now install face and rim setup on the actual machine, and zero the indicators. With indicators at the top, deflect bracket upward an amount equal to the appropriate rim sag, reading on the rim indicator, and note the face indicator reading. The face sag correction with indicators at bottom would be this amount with opposite sign. If zeroing the setup at the bottom, the face sag correction at the top would be this amount with same sign (if originally determined at top, as described).

Face Sag Effect-Examples


Example 1

Face and rim indicators are to be used together as shown in Figure 5-3. Assume you will obtain the following from your sag test: Rim sag with rim indicator only = .004 in. Rim sag with two indicators = .007 in. Mount the setup on the machine in the field, and with indicators at top, deflect the bracket upward .007 in. as measured on the rim indicator. When this is done, the face indicator reads plus .002 in. Face sag correction at the bottom position would therefore be minus .002 in. If you wish to zero at the bottom for alignment, but otherwise have data as noted, the face sag correction at the top would be plus .002 in.

Example 2

Face and rim indicators are to be used separately to reduce sag. Both indicators are the same type and weight. Other basic data are also the same.

Install face indicator and temporary rim indicator on the machine in the field, and place in top position. Zero indicators and deflect upward ,004 in. as measured on rim indicator. Face indicator reads plus .0013 in. Face sag correction at the bottom would therefore be minus .0013 in, If zeroing at the bottom for alignment, but otherwise the same as above, face sag correction at top would be plus .0013 in. Example 3

This will determine sag for “3-Indicator Face-and-Rim Setup” shown in Figure 5-4.

Set up the jig to the same geometry as for field installation but with rim indicator only and roll 180” top to bottom on pipe to get total single indicator rim sag - (Step 1).

Zero rim indicator on top and add or “lay on” face indicator, noting rim indicator deflection that occurs - (Step 2). Double this __ (Step 3).

Add it to original total single indicator rim sag (Step 1). - (Step 4). This figure, preceded by a plus sign, will be the sag correction for the rim indicator readings taken at bottom. With field measurement setup as shown, zero all indicators, and deflect the indicator end of the upper bracket upward an amount equal to the total rim sag (Step 4). Note the face sag effect by reading the face indicator. This amount, with opposite sign, is the face sag correction to apply to the readings taken at the lower position - (Step 5). Now deflect the upper bracket back down from its “total rim sag” deflection an amount equal to Step 3.

The amount of sag remaining on the face indicator, preceded by the same sign, is the sag correction for the single face indicator being read at the top position - (Step 6).

All of the foregoing refers of course to bracket sag. In long machines, we will also have shaft sag. This is mentioned only in passing, since there is no need to do anything about it at this time. Our “point-by-point” alignment will automatically take care of shaft sag. For initial leveling of large turbogenerators, etc., especially if using precision optical equipment, shaft sag must be considered. Manufacturers of such machines know this, and provide their erectors with suitable data for sag compensation. Further discussion of shaft sag is beyond the scope of this text.



Leveling Curved Surfaces

It is common practice to set up the “rim” dial indicators so their contact tips rest directly on the surface of coupling rims or shafts. If gross misalignment is not present, and if coupling and/or shaft diameters are large, which is usually the case, accuracy will often be adequate. If, however, major misalignment exists, and/or the rim or shaft diameters are small, a significant error is likely to be present. It occurs due to the measurement surface curvature, as illustrated in Figures 5-13 and 5-14. This error can usually be recognized by repeated failure of top-plusbottom (T + B) readings to equal side-plus-side (S + S) readings within one or two thousandths of an inch, and by calculated corrections resulting in an improvement which undershoots or overshoots and requires repeated corrections to achieve desired tolerance. A way to minimize this error is to use jigs, posts, and accessories which “square the circle.” Here we attach flat surfaces or posts to the curved surfaces, and level





them at top and bottom dead center. This corrects the error as shown in Figure 5-14.

For this method to be fully effective, rotation should be performed at accurate 90” quadrants, using inclinometer or bubble-vial device. In most cases, however, this error is not enough to bother eliminating- it is easier just to make a few more corrective moves, reducing the error each time.

Jig Posts

The preceding explanation showed a rudimentary auxiliary surface, or “jig post,” used for “squaring the circle.” A more common reason for using jig posts is to permit measurement without removing the spacer on a concealed hub gear coupling. If jig posts are used, it is important that they be used properly. In effect, we must ensure that the surfaces contacted by the indicators meet these criteria:

As already shown, they must be leveled in coordination at top and bottom dead centers, to avoid inclined plane error.

If any axial shaft movement can occur, as with sleeve bearings, the surfaces should also be made parallel to their shafts. This can be done by leveling axially at the top, rotating to the bottom, and rechecking. If bubble is not still level, tilt the surface back toward level for a half correction.

If face readings are to be taken on posts, the post face surfaces should be machined perpendicular to their rim surfaces. In addition to this, and to steps 1 and 2 just described, rotate shafts so posts are horizontal. Using a level, adjust face surfaces so they are vertical. Rotate 180” and recheck with level. If not still vertical, tilt back toward vertical to make a half correction on the bubble. This will accomplish our desired objective of getting the face surface perpendicular to the shaft in all measurement planes.

The foregoing assumes use of tri-axially adjustable jig posts. If such posts are not available, it may be possible to get good results using accurately machined nonadjustable posts. If readings and corrections do not turn out as desired, however, it could pay to make the level checks as described-they might pinpoint the problem and suggest a solution such as using a nonpost measurement setup.

interpretation and Data Recording


Due to sag as well as geometry of the machine installation, it is difficult and deceptive to try second-guessing the adequacy of alignment solely from the “raw” indicator readings. It is necessary to correct for sag, then note the “interpreted” readings, then plot or calculate these to see the overall picture-including equivalent face misalignment if primary readings were reverse-indicator on rims only. Sometimes thermal offsets must be included, which further complicates the overall picture. As a way to systematically consider these factors and arrive at a solution, it is helpful to use prepared data forms and stepwise calculation. Suppose we are using the two-indicator face-rim method shown in Figure 5-3; let’s call it “Setup #1.” To start, prepare a data sheet as shown in Figure 5-15. Next, measure and fill in the “basic dimensions’’ at the top. Then, fill in the orientation direction, which is north in our example. Next, take a series of readings, zeroing at the top, and returning for final readings which should also be zero or nearly so. Now do a further check: Add the top and bottom readings algebraically (T + B), and add the side readings (S + S). The two sums should be equal, or nearly so. If the checks are poor, takc a new set of readings. Do the checks before accounting for bracket sag. Now, fill in the known or assumed bracket sag. If the bracket does not sag (optimist!), fill in zero. Combine the sag algebraically with the vertical rim reading as shown, and get the net reading using (+) or (-) as appropriate to accomplish the sag correction. A wellprepared form will have this sign printed on it. If it does not, mentally figure out what must be done to “un-sag” the bracket in the final position, and what sign would apply when doing so.

Now we are ready to interpret our data in the space provided on the form. To do this, first take half of our net rim reading:




Formullll

This is because we are looking for centerline rather than rim offset. Since its sign is minus, we can see from the indicator arrangement sketch that the machine element to be adjusted is higher than the stationary element, at the plane of measurement. This assumes the use of a conventional American dial indicator, in which a positive reading indicates contact point movement into the indicator.

By the same reasoning, we can see that the bottom face distance is .007 in. wider than the top face distance. Going now to the horizontal readings, we make the north rim reading zero by adding - .007 in. to it. To preserve the equality of our algebra,

we also add - .007 in. to the south rim reading, giving us - .029 in. Taking half of this, we find that the machine element to be adjusted is .0145 in. north of the stationary element at the plane of measurement. Finally, we do a similar operation on our horizontal face readings, and determine that the north face distance is wider by .014 in. The remaining part of the form provides space to put the calculated corrective



Nelson’s method is easy to understand, and it works. It is basically a four-step procedure in this order:

1. Vertical angular Correction.

2. Vertical parallel correction.

3. Horizontal angular correction.

4. Horizontal parallel correction.

It has three disadvantages, however.

First, it requires four steps, whereas the more complex mathematical methods can combine angular and parallel data, resulting in a two-step correction. Secondly, it is quite likely that initial angular correction will subsequently have to be partially “un-done,” when making the corresponding parallel correction. Nobody likes to cut and install shims, then end up removing half of them. Finally, it is designed only for face-and-rim setups, and does not apply to the increasingly popular reverse-indicator technique.

We will now show two additional examples, wherein the angular and parallel correction are calculated at the same time, for an overall twostep correction. Frankly, we ourselves no longer use these methods, nor do we still use Nelson’s method, but are including them here for the sake of completeness. Graphical methods, as shown later, are easier and faster. In particular, the alignment plotting board should be judged extremely useful. Readers who are not interested in the mathematical method may wish to skip to page 192, where the much easier graphical methods are explained. But, in any event, here is the full mathematical treatment.

In our first example, we will reuse the data already given in our setup No. 1 data sheet.

First, we will solve for vertical corrections:

Using Nelson’s method, we found it necessary to make a 0.053 in. shim correction. Let us arbitrarily say this will be a shim addition beneath the inboard feet. At the coupling face, we then get a rise of:





As you can see, the values found this way are close to those found earlier. The main problem people have with applying these formulas is choosing between plus and minus for the terms. The easiest way, in our opinion, is to visualize the “as found” conditions, and this will point the way that movement must proceed to go to zero misalignment. For example, our bottom face distance is wide-therefore we need to lower the feet (pivoting at plane A) which we denote with a minus sign. The machine element to be adjusted is higher at plane A-so we need to lower it some more, which takes another minus sign. For the horizontal, our north face distance is wider, so we need to move the feet north (again pivoting at plane A). The machine element to be adjusted is north at plane A, so we need to move it south. Call north plus or minus, so long as you call south the opposite sign. Not really hard, but a lot of people have trouble with the concept, which is why we prefer to concentrate on graphical methods , where direction of movement becomes more obvious. We will get into this shortly, but first let’s do a reverse-indicator problem mathematically.

For our reverse-indicator example, we will use the setup shown earlier as Figure 5-6. Also, we must now refer to the appropriate data sheet, Figure 5-18. Finally, we resort to some triangles, Figures 5-19 and 5-20, to assist us in visualizing the situation.



Tekni┼čk Resimmm

Graphical Techniques

Now let’s see how we can do the same thing more easily. To do this, we will turn to graphical techniques. Reference 7 shows one version of this in an article on alignment of barrel-type centrifugal compressors. Its author stretches a strip-chart out on a drafting table, and rules in the machine element shafts, to scale, oriented as shown by reverse-indicator readings, much as we did in our previous example. Movements to achieve zero offset, or to compensate for thermal growth, can then be plotted on the same sheet. This technique has the advantage of giving a complete and permanent written visual record of what is happening, and is especially useful on multi-element machinery




After reworking our first example with the alignment plotting board, we will proceed to our second example, using the reverse-indicator setup (setup 4). This is illustrated for us in Figures 5-18 and 5-24. We will then go on and give worked-out examples using the other setups in Figures 5-25 through 5-33.

Three-Indicator Face-and-Rim

Figure 5-27 depicts a pump with ball bearings driven by a motor having sleeve bearings. External obstructions prevent the use of setups 4 or 5, and the coupling surfaces have considerable runout, making setup 1 inadvisable. It is therefore decided to use setup 2-the three-indicator face-and-rim. A sag check shows 0.002 in. sag for the rim indicator bracket. Sag for the face indicators will be ignored. A preliminary straightedge alignment is made in the horizontal plane. Both shafts are then turned together and a set of vertical readings taken as shown on the accompanying data sheet in Figure 5-28.


















trains such as the article discusses.

Another graphical technique, which we prefer for most situations, is the Alignment Plotting Board. It is fast, easy, and accurate, particularly on the simple two-element trains comprising the majority of all machines. It can be used on multi-element trains, but is less efficient than graph paper beyond two elements. To our knowledge, the plotting board is the only method which handles both the face-and-rim and reverse-indicator setups with equal ease and can convert from R-I to F&R. An additional advantage is its portability-it is only 8V2 X 11 in. and is made from an “indestructible” polycarbonate plastic. Figure 5-22 shows it in use.

Let’s return to our first example. For convenience, we will refer to the data sheet in Figure 5-15, then show how to find the corrections on the plotting board. This is graphically explained in Figure 5-23. By flipping back and forth between data sheet and plotting board sketches, you can see where the answers come from. It will help to also read the instructions that come with the plotting board, if you have one. With practice, the answers will appear in a minute or two.

movements. Although these have been filled in for our example, let’s leave them for the time being, since we are not yet ready to explain the calculation procedure. We will show you how to get these numbers later. If you think you already know how, go ahead and try-the results may be interesting.

You have now seen the general idea about data recording and interpretation. By doing it systematically, on a prepared form corresponding to the actual field setup, you can minimize errors. If you are interrupted, you will not have to wonder what those numbers meant that you wrote down on the back of an envelope an hour ago. We will defer consideration of the remaining setups, until we have explained how to calculate alignment corrective movements. We will then take numerical examples for all the setups illustrated, and go through them all the way. Calculating the Corrective Movements

Many machinists make alignment corrective movements by trial and error. A conscientious person can easily spend two days aligning a machine this way, but by knowing how to calculate the corrections, the time can be cut to two hours or less.

Several methods, both manual and electronic, exist for doing such calculations. All, of course, are based on geometry, and some are rather complicated and difficult to follow. For those interested in such things, see Refercnces 1 through 15. Also, the alignment specialist should be aware of programmable calculator solutions. These make use of popular calculators such as the TI 59 and HP 67. By recording the alignment measurements on a prepared form, and entering these figures in the prescribed manner into the calculator, the required moves come out as answers. A variation of this is the TRS 80 pocket computer which has been programmed to do alignment calculations via successive instructions to the user telling him what information to enter.

By far the simplest calculator is the one described earlier in conjunction with the laser-based OPTALIGN@ system. Next to it, we rank the IMS calculator. l4 This is designed specifically “from the ground up’’ for the more common face-and-rim and reverse-indicator calculations, rather than being a standard commercial calculator or computer adapted for alignment. For the more complex multi-element trains, larger central computers can be used directly with telephone linkups. Ray Dodd’s book (Total Alignment, Reference 3) describes one such system. Another that we saw demonstrated provides not only the numerical data, but a diagrammatic representation of the machines and their alignment relationship, on a cathode ray tube.

The foregoing electronic systems are popular, and have advantages in speed, accuracy, and ease of use. They have disadvantages in cost, usability under adverse field and hazardous area conditions, pilferage, sensitivity to damage from temperature extremes and rough handling, and availability to the field machinist at 2:OO A.M. on a holiday weekend. They also, for the most part, work mainly with numbers, and the answers may require acceptance on blind faith. By contrast, graphical methods inherently aid visualization by showing the relationship of adjacent shaft centerlines to scale.

Manual calculation methods have the advantage of low investment (pencil and paper will suffice, but a sliderule or simple calculator will be faster). They have the disadvantage, some say, of requiring more thinking than the programmed electronic solutions, particularly to choose the plus and minus signs correctly.

The graphical methods, which we prefer, have the advantage of aiding visualization and avoiding confusion. Their accuracy will sometimes be lcss than that of the “pure” mathematical methods, but usually not enough to matter. Investment is low-graph paper and plotting boards are inexpensive. Speed is high once proficiency is attained, which usually does not take long.

In this text, we will emphasize the graphical approach. Before doing so, let’s highlight some common manual mathematical calculations. Nelson” published an explanation of one rather simple method a number of years ago. A shortened explanation is given in Figure 5-16. For our given example, this would work out as follows:



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What is Fatigue

Fatigue is a technical term that elicits a degree of curiosity. When citizens read or hear in their media of another fatigue failure, they wonder whether this has something to do with getting tired or "fatigued" as they know it. Such is not the case.

One way to explain fatigue is to refer to the ASTM standard definitions on fatigue, contained in ASTM E 1150. It is difficult, if not impossible, to carry on intelligent conversations if discussions on fatigue do not use a set of standard definitions such as E 1150. Within E 1150, there are over 75 terms defined, including the term fatigue:
"fatigue (Note 1):
the process of progressive localized permanent structural change occurring in a material subjected to conditions that produce fluctuating stresses and strains at some point or points and that may culminate in cracks or complete fracture after a sufficient number of fluctuations (Note 2). Note 1--In glass technology static tests of considerable duration are called `static fatigue' tests, a type of test generally designated as stress-rupture. Note 2--Fluctuations may occur both in load and with time (frequency) as in the case of `random vibration'." (Ref 2). The words in italics (emphasis added) are viewed as key words in the definition. These words are important perspectives on the phenomenon of fatigue:

· Process
· Progressive
· Localized
· Permanent structural change
· Fluctuating stresses and strains
· Point or points
· Cracks or complete fracture
The idea that fatigue is a process is critical to dealing with it in design and to the characterization of materials as part of design. In fact, this idea is so critical that the entire conceptual view of fatigue is affected by it! Another critical idea is the idea of fluctuating stresses and strains. The need to have fluctuating (repeated or cyclic) stresses acting under either constant amplitude or variable amplitude is critical to fatigue. When a failure is analyzed and attributed to fatigue, the only thing known at that point is that the loads (the stresses/strains) were fluctuating. Nothing is necessarily known about the nucleation of damage that forms the origin of fatigue cracks. Design for Fatigue Prevention In design for fatigue and damage tolerance, one of two initial assumptions is often made about the state of the material. Both of these are related to the need to invoke continuum mechanics to make the stress/strain/fracture mechanics analysis tractable:· The material is an ideal homogeneous, continuous, isotropic continuum that is free of defects or flaws.· The material is an ideal homogeneous, isotropic continuum but contains an ideal cracklike discontinuity that may or may not be considered a defect or flaw, depending on the entire design approach. The former assumption leads to either the stress-life or strain-life fatigue design approach. These approaches are typically used to design for finite life or "infinite life." Under both assumptions, the material is considered to be free of defects, except insofar as the sampling procedure used to select material test specimens may "capture" the probable "defects" when the specimen locations are selected for fatigue tests. This often has proved to be an unreliable approach and has led, at least in part, to the damage-tolerant approach. Another possible difficulty with these assumptions is that inspectability and detectability are not inherent parts of the original design approach. Rather, past and current experience guide field maintenance and inspection procedures, if and when they are considered.The damage-tolerant approach is used to deal with the possibility that a crack-like discontinuity (or multiple ones) will escape detection in either the initial product release or field inspection practices. Therefore, it couples directly to nondestructive inspection (NDI) and evaluation (NDE). In addition, the potential for initiation of crack propagation must be considered an integral part of the design process, and the subcritical crack growth characteristics under monotonic, sustained, and cyclic loads must be incorporated in the design. The final instability parameter, such as plane strain fracture toughness (KIc), also must be incorporated in design. The damage-tolerant approach is based on the ability to track the damage throughout the entire life cycle of the component/system. It therefore requires extensive knowledge of the above issues, and it also requires that fracture (or damage) mechanics models be available to assist in the evaluation of potential behavior. As well, material characterization procedures are needed to ensure that valid evaluation of the required material "property" or response characteristic is made. NDI must be performed to ensure that probability-of-detection determinations are made for the NDI procedure(s) to be used. This approach has proved to be reliable, especially for safety-critical components.The above approaches often are used in a complementary sense in fatigue design. The details of all three approaches are discussed in this Volume.The fatigue process has proved to be very difficult to study. Nonetheless, extensive progress on understanding the phases of fatigue has been made in the last 100 years or so. It now is generally agreed that four distinct phases of fatigue may occur (Ref 3, 4):· Nucleation· Structurally dependent crack propagation (often called the "short crack" or "small crack" phase)· Crack propagation that is characterizable by either linear elastic fracture mechanics, elastic-plastic fracture mechanics, or fully plastic fracture mechanics· Final instabilityEach of these phases is an extremely complex process (or may involve several processes) in and of itself. For example, the nucleation of "fatigue" cracks is extremely difficult to study, and even "pure fatigue" mechanisms can be very dependent on the intrinsic makeup of the material. Obviously, when one decides to pursue the nucleation of cracks in a material, one has already either assumed that the material is crack-free or has proved it! The assumption is the easier path and the one most often taken. When extraneous influences are involved in nucleation, such as temperature effects (e.g., creep), corrosion of all types, or fretting, the problem of modeling the damage is formidable. In recent years, more research has been done on the latter issues, and models for this phase of life are beginning to emerge.
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Lubrication Oils Greases Lubricant Selection Lubricating Methods

Adequate lubrication of rolling bearings is required for achieving the life calculated for any bearing. In a correctly operating rolling element bearing, a thin film of lubricant separates the rolling elements from the raceways. This film should be of sufficient thickness to actually prevent the rolling element surfaces from touching the inner and outer ring raceways. Contact of the raceway surfaces will result in wear, scoring, and possible seizure. Providing this film is the primary function of the lubricant to four types of internal bearing contact:true rolling contact of the rolling elementhaceways sliding contact between the cage and other bearing components partial slidhgh-olling contact in some bearing types sliding contact between the rollers and guide ribs in roller bearings In addition, the lubricant has several important secondary functions: protection from corrosion exclusion of contaminants flushing away of wear products and debris dissipation of heatThe requirements of a lubricant for rolling element bearings are often more severe than realized. In a rolling element bearing, there are conditions of both rolling and sliding with extremely high contact pressures. The lubricant must withstand high rates of shear and mechanical working not generally prevalent in other mechanical components. For these reasons, proper attention to lubrication is vital for successful bearing operation. Oil is a liquid lubricant which can be pumped, circulated, atomized, filtered, cleaned, heated, and cooled, making it more versatile than grease. It is suitable for many severe applications involving extreme speeds and high temperatures. On the other hand, it is more difficult to seal or retain in bearings and housings and, in general, involves a more complicated system than grease.
Viscosity, the measure of an oil’s thickness, is the most important property of lubricating oil. The selection of proper viscosity is essential and is based primarily on expected operating temperatures of bearings. Excessive oil viscosity may cause skidding of rolling elements and high friction. Insufficient oil viscosity may result in metal-tometal contact of the rolling surfaces.There are two general categories for liquid lubricants: petroleum or mineral oils, and synthetic oils. Mineral oils are lower in cost and have excellent lubricating properties. Synthetic lubricants have been developed to satisfy the need for a wider operating temperature range than is possible with mineral oils. This development has been prompted prompted by the extreme environmental demands of military and aerospace applications. There is a wide range of synthetic types with varying temperature limits. The maximum temperature limits for the common types are given in Table 11. The major disadvantage of synthetics is that they do not have the same load-carrying capacity as do mineral oils at typical industrial equipment operating temperatures. Also, synthetics are rarely compatible with mineral oils, so care must be taken when both types are being used in proximity.Grease is a combination of mineral oil or a synthetic fluid and a suitable thickener (often called soup). The percentage of the oil in grease is usually about 80%, but can range from 70% to 97%. Grease consistency or stiffness is determined primarily by the thickener and base oil viscosity. Greases of a given consistency may be formulated from various combinations of thickener and base oil viscosity so that greases of equal stiffness are not necessarily equal in performance. Greases considered satisfactory lubricants for rolling element bearings are combinations of soap or nonsoap agents, mineral oil, and additives. Soaps such as sodium, calcium, barium, aluminum, lithium, complexes of these soaps, and nonsoaps such as silica and special clays are generally used. Rust and oxidation inhibitors and extreme pressure agents are often added. Lithium and lithium-complex thickeners seem to give the best all-purpose performance, but each type has its advantages.Performance of a gEase depends on several different factors. The lubricating capability of the grease is mainly dependent on the properties of the base oil used. The corrosion protection of the grease is determined by the thickener.the base oil also, it is usually restricted by the thickener. Table 12 shows the maximum temperature limits for the most common thickener types.Be careful to avoid mixing greases of Merent soap bases. The combination will usually be worse than either one by itself, and sometimes worthless. Care should be taken when mixing greases with the same soap bases from different manufacturers, although this is usually not a problem. However, it is not always obvious what the soap base is unless the manufm’s data sheet is consuld. In no case should mineral oil greases be mixed with greases using a synthetic oil. The major criteria for selection of a lubricant is the viscosity. The selection of the proper viscosity oil is especially important for bearings operating in the high load, speed, or temperature ranges. As mentioned in the section OR life adjustment factors, it is necessary to have A between 1.1 and 1.3. The A factor is a mure of the mtio of the oil film thickness to the surface roughness of the raceways in contact. In other words, the oil film needs to be thicker than the race way roughness so that there is never metal-to-metal contact. The viscosity of most oils changes dramatically with a change in temperature. When determining the operating temperature, it is the oil temperature that is important. Generally, the oil temperam is 5" to 20°F greater than that of the bearing housing.The oil film thickness is calculated through the theory of elastohydrodynamic lubrication. This involves the elastic deformations of the raceway contacts, and the pressure-viscosity effects and hydrodynamics of the lubricant. This theory is very complex and best left to computer programs. However, a simplified method for determining if the oil film thickness is sufficient involves using Figures 12 and 13. From Figure 12, find the oil viscosity needed based on the beating size and the operating speed. Then, using Figure 13, combine that viscosity with the operating temperature of the oil to determine what grade of oil is needed. This should give a ballpark estimate of which oil to use: It assumes the use of a mineral oil with a viscosity index of 95. If a much different oil is to be used w e a synthetic, for example), it is best to consult with a bearing manufacturer or oil supplier to make a more detailed calculation. If the A factor is less than 1.1, the bearing life will be reduced. If A is greater than 1.3, the bearing life will be increased. This implies that the highest viscosity possible should be used. However, as the viscosity goes up, so does the operating temperature. As temperature goes up, the oil viscosity goes down and maintenance activity goes up. This all means that there is a practical limit to the life improvement from higher oil viscosity. There are a variety of methods to apply the proper amount The simple oil bath method is satisfactory for low and moderate speeds. The oil level in the housing should not be less than the lip of the outer ring, nor higher than the center of the lowest rolling element. The oil level should only be checked when the bearing is not rotating. Circulating oil is an excellent way to lubricate a bearing, especially on large machines, and can reduce maintenance and prolong the life of the oil in severe operating conditions. of oil to a bearing. The most common are as follows:Oil bath Circulating systems Jet lubrication Mist lubrication Wick feed It can be used with either a wet or dry sump, with the oil usually introduced on one side of the bearing and drained on the other. A system shutoff with loss of pressure is a desired feature. This system is good for all speeds and loads. The main drawback of such a system is the cost. Jet lubrication is a special type of circulating oil system used on very high speed bearings such as in a gas turbine engine. Most of the oil in this method is used for cooling of the bearing. This method can be used at speed levels up to 1.5 million DN with proper design. The jet should be aimed at the largest space between the cage and the ring lands. Mist lubrication is of two types, which are distinguished by the method of generating the mist. In some applications, the mist is generated by a flinger that dips into the oil and throws it into the air in the vicinity of the bearing. Sometimes gears substitute for the flinger. Another way to generate the mist is to spray a jet of oil against the side on the inside of the machinery. This method can be very effective for bearings where cooling is not needed.The second type of oil mist lubrication is when the mist is produced by a special mist generator. The oil mist is formed in an atomizer and supplied to the bearing housing under suitable pressure. This method of lubrication has proven very effective in reducing the operating temperature, not so much by air cooling as by the flow of air, preventing excess oil from accumulating in the bearing. Since the air pressurizes the housing and escapes through the seals, the entrance of moisture and grit is retarded. No drain is needed, as the quantity of oil supplied is very small. The problems with mist oil generators is that the immediate area may be coated with oil and if the oil generator shuts off, the bearings cannot survive long because of the small amount of oil supplied. Also, if there are air pressures created by other parts of the mechanical system, they should be checked to make sure they are not restricting the flow of the air mist under all operating conditions.Wick feed is also suitable for high speeds because, again, a small amount of oil is delivered to the bearing. Careful maintenance is needed to make sure the cup never runs dry and that the wick is always in contact with the source. Grease systems are not as numerous as those for oil. Many bearings come from the manufacturer with a supply ings by hand, filling the internal volume of the bearing one-third to one-half full. A grease gun with a grease fitting on the housing can be used, but care should be exercised not to overfill the housing and cause overheating of the bearing. One method of gauging the amount of grease to add is to add grease slowly while the beating is running until some grease is just visible coming out either seal. The grease fitting should then be removed briefly to allow any grease backpressure to relieve itself, and then be reinstalled.There are automatic grease systems on the market that can relieve a lot of maintenance activity when a number of bearings can be grouped into a system. The bearings and/or their housings need to be packed with grease before the system is operated. The disadvantages of these systems is their initial cost and that all bearings on any one system must be able to use the same grease. Bearings with special needs would need a separate system.
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Friction Welding

Unlike the other processes covered in this book friction welding is a solid phase pressure welding process where no actual melting of the parent metal takes place. The earliest version of the process utilised equipment similar to a lathe where one component was held stationary and the other held in a rotating chuck (Fig. 8.11). Rubbing the two faces together produces suf- ficient heat that local plastic zones are formed and an end load applied to the components causes this plasticised metal to be extruded from the joint,carrying with it any contaminants, oxides, etc. Thus two atomically clean metal surfaces are brought together under pressure and an intermetallic bond is formed. The heat generated is confined to the interface, heat input is low and the hot work applied to the weld area results in grain refinement. This rapid, easily controlled and easily mechanised process has been used extensively in the automotive industry for items such as differential casings, half shafts and bi-metallic valves. Since the introduction of this conventional rotating method of friction welding many developments have taken place such as stud welding, friction surfacing, linear and radial friction welding, taper plug welding and friction stir welding.
One very important characteristic of friction welding is its ability to weld alloys and combinations of alloys previously regarded as unweldable. It is possible to make dissimilar metal joints, joining steel, copper and aluminium to themselves and to each other and to successfully weld alloys such as the 2.5% copper–Al 2618 and the AlZnMgCu alloy 7075 without hot cracking. The primary reason for this is that no melting takes place and thus no brittle intermetallic phases are formed.
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Fly Wheels

The energy-storage capacity of a flywheel is determined from its polar moment ofinertia / and its maximum safe running speed. The necessary inertia depends on thecyclic torque variation and the allowable speed variation or, in the case of energystorageflywheels, the maximum energy requirements. The safe running speeddepends on the geometry and material properties of the flywheel.Flywheels store energy. Indeed, flywheels are used as energy reservoirs, and this usewill be discussed in Sec. 18.4. Their principal use in machine design, however, is tosmooth the variations in shaft speed that are caused by loads or power sources thatvaryirTa cyclic fashion. By using its stored kinetic energy 0.5/co2 to absorb the variationsin torque during a machine cycle, a flywheel smooths the fluctuating speed of amachine and reduces undesirable transient loads. The effect of a flywheel is thereforefundamentally different from that of a regulator: A flywheel limits the speedvariation over one cycle and has minimal effect on the average speed; a regulatoruses negative feedback to maintain a selected average speed with only secondaryeffects on the speed during a cycle.The flywheel has other features which have to be considered in design. Its size,speed, and windage effect can all be used to advantage in providing a secondaryfunction as part of a clutch, gear, belt pulley, cooling fan, pump, gyroscope, or torsionaldamper.
If the torque-angle curve for a machine cycle is available from experimental dataor a dynamic analysis, U is determined from the areas between the curve and theaverage-torque line. If the external torque input or load is not constant, it can becombined with the torque-angle curve for the machine. If the loading torque and thedriving torque are not synchronized or have an unknown phase difference, a worstcasecombination should be used. The areas under the curve can be determinedusing a planimeter or by graphic or numerical integration as shown in Chap. 4. (Seealso Example 4 or consult the user handbook for your programmable calculator orcomputer.) Unless C5 is accurately known and the curve is from a worst case or ishighly repeatable, precision in integrating is not warranted.
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Types of Turbines

Turbines can be classified according to the direction of the water flow through the blades, e.g. radial, axial or combinedflow turbines, or as reaction, impulse or mixed-flow turbines. In reaction turbines there is a change of pressure across the turbine rotor, while impulse turbines use a high velocity jet impinging on hemispherical buckets to cause rotation. There are three basic types of turbine broadly related to low, medium or high heads.
Propeller or axial flow turbines are used for low heads in the range from 3 to 30 metres. They can have relatively inexpensive fixed blades, which have a high conversion efficiency at the rated design conditions but a poorer par€-load efficiency, typically 50%, at one third of full rated output.
Alternatively, the more expensive Kaplan turbine has variable-pitch blades which can be altered to give much better part-load efficiency, perhaps 90% at one third of full rated output. The Francis turbine is a mixed-flow radial turbine and is used for medium heads in the range from 5 to 400 m. It has broadly similar performance characteristics to the fixed-blade propeller type and its speed is controlled by adjusting the guide vane angle. The best-known impulse turbine is the Pelton wheel.
Each bucket on the wheel has a centrally placed divider to deflect half the flow to each side of the wheel. It is normally used for heads greater than 50 m and has good performance characteristics over the whole range, very similar to the Kaplan turbine, reaching 60% efficiency at one-tenth of full rated output.

The speed is controlled by a variable inlet nozzle, so that with a constant head, the delivered torque to the generator is proportional to the flowrate and the turbine speed can be held at that required for synchronous generation at the particular grid frequency. This type of installation is known as a constantspeedkonstant- frequency system and optimization of the power output is relatively easy. I1O In smaller installations, optimum power cannot be obtained at constant speed where the hydraulic head is both relatively low and variable over a wide range.
A detailed description of methods which can be used for optimizing electric power from small-scale plant has been given by Levy.”’ He points out that small hydroelectric systems will become more financially attractive through developments of low-cost power converters (from 100 W upwards), special variable-speedkonstant-frequency generators and cheap computing units for on-line power measurement and optimizing control. This means that many run-of-the-river sites that were considered in the past to be unsuitable for electricity generation can now be used
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Bearing Maintenance and Replacement

The fundamental purpose of a bearing is to reduce friction and wear between rotating parts that are in contact with one another in any mechanism. The length of time a machine will retain its original operating efficiency and accuracy will depend upon the proper selection of bearings, the care used while installing them, proper lubrication, and propcr maintenance provided during actual operation. The manufacturer of the machine is responsible for selecting the correct type and size of bearings and properly applying the bearings in the equipment. However, maintenance of the machine is the responsibility of the user. A well-planned and systematic maintenance procedure will assure extended operation of the machine. Failure to take the necessary precautions will generally lead to machine downtime. It must also be remembered that factors outside of the machine shaft may cause problems. Engineering and Interchangeability Data Rings and Bulls-The standard material used in ball bearing rings and balls is a vacuum processed high chromium steel identified as SAE 52100 or AISI-52100. Material quality for balls and bearing rings is maintained by multiple inspections at the steel mill and upon receipt at thc bcaring manufacturing plants. The 52 100 bearing steel with standard heat treatment can be operated satisfactorily at temperatures as high as 250°F (121 "C). For higher operating temperatures. a special heat treatment is required in order to give dimensional stability to the bearing parts.


Seals-Standard materials used in bearing seals are generally nitrile rubber. The material is bonded to a pressed steel core or shield. Nitrile rubber is unaffected by any type of lubricant commonly used in anti-friction bearings. These closures have a useful temperature range of -70" to +225"F (-56" to 107°C). For higher operating temperatures, special seals of high temperature materials can be supplied. Ball Cages-Ball cages are pressed from low carbon steel of SAE 1010 steel. This same material is used for bearing shields. Molded nylon cages are now available for many bearing sizes. The machined cages ordinarily supplied in super-precision ball bearings are made from laminated cotton fabric impregnated with a phenolic resin. This type of cage material has an upper temperature limit of 225°F (107°C) with grease and 250°F (121 "C) with oil for extended service. For periods of short exposure, higher temperatures can be tolerated.

Removal of Shaft and Bearings from Housing

The first step in dismantling a spindle or shaft is to remove the shaft assembly from the housing. To do this, it is generally necessary to take off the housing covers from each end.

Most machine tool spindle and API pump housings are constructed with bearing seats as an integral part of the housing. This contributes to the rigidity of the spindle. However, it makes disassembly more difficult and extreme care must be taken to avoid bearing damage. Also, it is not generally possible to remove bearings from the shaft unless the shaft assembly is first removed from the housing.

On most spindle assemblies this can be done by first placing the entire spindle in an arbor press and in alignment with the press ram. Next, carefully apply pressure to the end of the shaft making sure that there is clearance for the expulsion of the shaft assembly on the press table. As pressure is applied, the shaft is forced from the housing along with the bearing mounted on the opposite end of the shaft.

The bearing on the end where pressure is applied remains in the housing. It is removed from the housing either with hand pressure or by carefully pushing it out of the housing from the opposite side with rod tubing having a diameter slightly smaller than the housing bore. The tubing should contact the bearing outer ring and should push it from the housing with little or no pressure on the balls and inner ring. Following this procedure will help avoid brinelling of the raceways due to excessive pressure on the rolling elements and races.

Electric motor shafts are generally constructed to permit removal of one end bell, leaving the shaft and bearings exposed. The rotor or shaft assembly is then free to be removed by drawing it through the stator.

Bearing Removal from Shaft

Removal of bearings from spindle shafts is a highly important part of the maintenance and service operation. In most cases, it is far more difficult to remove a bearing from the shaft than to put it on. For this reason, a bearing can be damaged unnecessarily in the process. Every precaution must be taken to avoid damage to any of the parts including the bearings. If the bearings are damaged during removal, the damage often is not noticed and may not become known until the spindle is completely reassembled. Bearing damage during removal from the shaft can occur in many ways, of which these are the most common:

The smooth, highly-polished surface of the ball raceways may be brinelled, i. e., indented, by the balls (Figure 7-4). Brinell marks on the surface of the races are usually caused when a bearing is forced off the shaft by applying excessive or uneven pressure through the rolling element complement. Any shock load, such as hammer blows on the inner or outer rings, is apt to cause brinelling. Major brinelling can sometimes be discovered on the job by applying a thrust load from each direction while rotating the inner or outer ring slowly. As the ring is turned through the brinelled area on either of the race shoulders, it can often be felt as a catch or rough spot. A brinelled bearing is unfit for further use. Never put it back into service. Ball raceways may be roughened due to dirt particles or metal chips working into the bearing. As soon as the shaft has been removed from the housing, it should be placed in a clean work area and suitably covered so that no contaminant can become lodged in the bearing prior to removal from the shaft. If contaminants enter the housing and the bearing is subsequently rotated, it is possible that they will roughen and damage the raceways.

The ball cage may be damaged if the bearing puller is used incorrectly. Use of improper tools such as a hammer or chisel to pound or pry the bearing off the shaft may result in damage to the bearing in addition to the hazard of contaminating the bearing.

Removal From Shaft

Because of operating conditions or location of the shaft, bearings are often tight and resist easy removal. This holds true even though they were originally mounted with a “push” fit, usual in most machine tool spindle applications. A “push” fit means ability to press the bearing on the shaft with hand pressure.

If these conditions occur, mechanical means such as a bearing puller (Figure 7-5) or the use of an arbor press (Figure 7-6) should be employed to effect bearing removal. The hammer and drift tube method, sometimes used to pound the bearing from the shaft, generally is not recommended, especially on machine tool spindle bearings. There is always the chance that the hammer shocks conducted through the tube will cause brinelling. For some types of bearings, electrical means of removal are possible as well. These removal methods will be described later.




Bearings are mounted on shafts or spindles in several ways so that dismounting must be accomplished by different means. Here are the most common conditions:

The bearing is free of grease and/or other parts. Place the shaft in an arbor press in line with the ram and with the inner ring of the bearing supported by a split ring having a bore slightly larger than the shaft (Figure 7-7). Press the shaft from the bearing with an even pressure, making sure it does not drop free and become damaged. If the split ring is not available, two flat bars of equal height could support the bearing (Figure 7-8).


Another means of removing a bearing from the shaft is by use of a bearing puller, several of which are shown in Figures 7-13 to 7-15. The bearing mounted with gears and/or other parts abutting it (Figure 7-9). In most cases, a bearing in this location can only be removed by a bearing puller which applies pressure on the outer ring (Figure 7-10). Extreme care must be exercised when applying pressure to make sure that the pull is steady and equal all around the outer ring. If the gears or other parts are removable, it may be possible to apply prcssure through them to force the bearing off the shaft. An arbor press may be employed to do the job if the bearing or gear can be adequately supported while pressure is applied




Applying Pressure with Bearlng PullerWhenever possible, bearings always should be moved from the shaft by square and steady pressure against the tight ring. Thus with a tight fit on the shaft, pressure should be against the inner ring; with a tight fit in the housing, pressure should be against the outer ring. If it is impractical to exert pressure against the tight ring, and the loose ring must be used, it is imperative that the same square and steady pull method be used. Pressure may be applied in either direction on bearings with shoulders of equal height (Figure 7-1 1). On counterbored bearings with one deep and one low shoulder, pressure should be applied against the deep shoulder. If pressure is applied against the low shoulder, disassembly of the bearing or serious damage may result. When the pairs of bearings on each end of the shaft are mounted in a back-to-back (DB) relationship, the counterbored outer ring is always exposed. In such cases, it will be necessary to apply the pressure against the low shoulder (counterbored ring) to effect bearing removal from the shaft even though the risk of damage to the inboard bearing is great.

Most machine tool spindles employ Type R or angular-contact bearings (7000 Scrics) that do not havc scals or shields. However, it is possible that a Conrad type bearing equipped with seals or shields may be used in some applications. When using pullers for bearing removal, care must be exercised to avoid damage to the seal or shield (Figure 7- 12). If dented and then remounted, an early bearing failure during operation could result. Bearing removal damage can be caused by the selection of the wrong puller type as easily as it can with improper use of the correct puller. No matter which puller is used, remember: the bearing must be pulled off squarely under steady pressure.Identification and Handling of Removed Bearings.

As it is possible that bearings may be suitable for remounting after servicing, it is necessary to replace them in exactly the same position on the shaft. Therefore, each bearing must be specifically tagged to indicate its proper location. Duplex bearings should be tied together in their proper relationship, DB, DF, or DT and the tag should also indicate the relationship. If a spacer is used between duplex bearings, the tag should indicate its position and relationship to the bearings. On jobs where the bearing is being removed because performance has not been fully successful, it is often desirable to find out why. Be sure to preserve the bearing until it is practical to examine it. The bearing frequently contains direct evidence as to the cause of failure. It should not be permitted to rust badly and the parts should be abused as little as possible during disassemble.



If the bearing is being removed for reasons other than bearing failure, be certain that it is thoroughly cleaned and oiled immediately after removal. Otherwise there is a good chance that it will get dirty and rusty. which would prevent its reuse.Bearing PullersThere are numerous types of bearing pullers on the market, any of which would be satisfactory to use depending upon the dismounting situation encountered. A conventional claw type is used where there is sufficient space behind the bearing puller claws to apply pressure to the bearing. In the illustration (Figure 7-13), the claws are pressing against the bearing preloading spring pack which in turn will force the duplex pair of bearings and spacers from the spindle. Another type of puller (Figure 7-14) uses a split-collar puller plate (Figure 7-15), the flange of which presses against the inner ring of the bearing. The puller bolts must be carefully adjusted so that the pulling pressure is equal all around the ring. The collar must be made in two pieces so that it can be slipped behind the bearing. The collar hole should be large enough so that the two pieces may be bolted together without gripping the shaft


Most bearing companies do not manufacture bearing pullers, but many bearing distributors stock a variety of the various pullers described above.Bearing Removal Through Application of Heat

The application of heat via special devices provides a rather straightforward way of removing inner bearing rings without damaging shafts. The device shown in Figure 7-16 is initially heated by an induction heater (see Figures 7-59 through 7-61, later in this chapter). To remove the inner ring from a bearing assembly [Figure 7-17( l)], the outer race and rolling elements must first be removed [Figure 7-17(2)]. The device is then heated to approximately 450°C (813°F) and slipped over the exposed ring, Figure 7-17(3). By simultaneously twisting and pulling [Figure 7- 17(4)], the operator clamps the heated pull-off device onto the ring. Within approximately 10 seconds, thc ring will have expanded to the point of looseness [Figure 7-17(5)] and can be removed.



Cleaning and inspection of Spindle Parts

Insufficient attention is paid to small dust particles which constantly blow around in the open air. But should a particle get in one’s eye. it becomes highly irritating. In like manner, when dirt or grit works into a ball bearing, it can become detrimental and often is the cause of bearing failure.

It is so easy for foreign matter to get into the bearing that more than ordinary care must be exercised to keep the bearing clean. Dirt can be introduced into a bearing simply by exposing it to air in an unwrapped state. Within a short period of time, the bearing can collect enough contaminants to seriously affect its operation. Special care must be taken when the bearing is mounted on a shaft, a time when it is most susceptible to contamination. This cleanliness requirement also extends to the handling of spindle parts, as everything must be clean when replaced in the assembly.Cleaning the BearingDuring the process of removal from a shaft, the bearing is likely to have become contaminated. The following procedure should be used to clean the bearing for inspection purposes as well as to prepare it for possible remounting on the shaft:




1. Dip the bearing in a clean solvent and rotate it slowly under very light pressure as the solvent runs through the bearing (Figure 7- 18). Continue washing until all traces of grease and dirt have been removed. Do not force the beuring during rotation.

2. Blow the bearing dry with clean, dry air while holding both inner and outer rings to keep the air pressure from spinning them. This avoids possible scratching of balls and raceways if grit still remains in the bearing. A slow controlled hand rotation under light pressure is advisable.


3. After blowing dry, rotate the bearing again slowly and gently to see if dirt can still be detected. Rewash the bearing as many times as necessary to remove all the dirt.

4. When clean, coat the bearing with oil immediately. Special attention should be given to covering the racewayr and balls to ensure prevention of corrosion to the highly finished surfaces. Rotate the bearing gently to coat all rolling surfaces with oil.After cleaning, the bearing should be wrapped with lint-free material such as plastic film to protect it from exposure to all contaminants. Unless this is done, it may be necessary to repeat the cleaning procedure immediately prior to remounting. As other spindle parts are cleaned, they also should be covered to exclude contamination which could ultimately work into the bearing.Cleaning the ShaftThe shaft must be cleaned thoroughly with special attention being paid to the bearing seats and fillets. If contaminants or dirt remain, proper seating of the shaft and/or against the shaft shoulder could be impossible. Don’t overlook the cleaning of keyways, splines, and grooves.

Cleaning the Housing.
Care should be taken to remove all foreign matter from the housing (Figure 7-19). Suitable solvents should be used to remove hardened lubricants. All corrosion should be removed. After cleaning, inspect in a suitable light the bearing seats and corners for possible chips, dirt, and damage, preferably using low power magnification for better results. The most successful method to maintain absolute cleanliness inside a clean housing is to paint the nonfunctional surfaces with a heat-resisting, quick-drying engine enamel. Do not paint the bearing seats of the housing. This would reduce the housing bore limits, making it difficult, if not impossible, to mount the bearings properly.

Painting seals the housing and prevents loose particles such as core sand from contaminating the bearing lubricants and eventually the bearings. It also provides a smooth surface which helps to prevent dirt from clinging to the surfaces. The housing exterior also may be painted to cover areas where old paint is worn or chipped; but do not paint any of the locating mating surfaces.

This type of work should be done in a place outside of the spindle assembly area.Keep Spindle Parts Coated With OilAs most of the spindle parts are usually of ferrous material, they are subject to corrosion. When exposed to certain atmospheric conditions, even nonferrous parts may become corroded and unusable in the spindle. Therefore, it is important to make certain that parts are not so affected from cleaning time until they are again sealed and protected in the spindle assembly. The best protection is to keep parts coated with a light-weight oil, covered, and loosely sealed with a plastic film or foil. Such a covering will exclude contaminants such as dust and dirt. When it is necessary to handle parts for inspection, repair, transportation, or any other purpose, precautions must be taken to ensure they are recoated with oil as some may have rubbed off during handling of the part.

Inspect All Spindle Parts
After the spindle parts have been cleaned thoroughly, the various parts should be inspected visually for nicks, burrs, corrosion, and other signs of damage (Figure 7-20). This is especially important for locating surfaces such as bearing seats, shaft shoulders, faces, and corners of spacer rings if any are used in the spindle, etc.Sometimes damage may be spotted by scuff marks or bright spots on the bearing, shaft, or in the housing.

This scoring may be caused by heavy press fits or build-up of foreign matter drawn onto the mating surfaces. Bright spots may also indicate early stages of “fidgeting” or scrubbing of mating surfaces. The shaft should also be checked for outof- round and excessive waviness on both two-point and multiple point gauging or checking on centers.Bearing Seats on ShaftThe shaft seat for the inner ring of a ball bearing is quite narrow and subject to unit pressures as high as 4,000 lbs per square in.

Because of this pressure, particular attention must be paid to the shaft fit to avoid rapid deterioration of the bearing seats due to creepage under heavy load and/or “fretting.”The required fit of the inner ring on the shaft will vary with the application and service. It is dependent on various factors such as rotation of the shaft with respect to the direction of the radial load, use of lock nuts, light or heavy loads, fast or slow speeds, etc.

In general, the inner ring must be tight enough not to turn or creep significantly under load When the bearing has too tight a fit on the shaft, the inner race expands and reduces or eliminates the residual internal clearance between the balls and raceways. Usually bearings, as supplied for the average application, have sufficient radial clearance to compensate for this effect. However, when extremes of shaft fit are inadvertently combined with insufficient radial clearance, extreme overload is caused and may result in heating and premature bearing failure.



Tight fits in angular-contact type bearings used for machine tools may cause changes in preload and contact angle, both of which have an effect upon the operating efficiency of the machine. Finally, rings may be split by too heavy a fit.Excessive looseness under load is also very objectionable because it allows a fidgeting, creeping, or slipping of the inner ring on the rotating shaft (Figure 7-22).



This action causes the surface metal of the shaft and bearing to fret, scrub, or wear off which progressively increases the looseness. It has been noticed that, in service, this working tends to scrub off fine metal particles which oxidize quickly, producing blue-black and brown oxides on the shaft and/or the bore of the bearing.



The bearing should be tight enough on the shaft to prevent this action. If any of these conditions are noticed on a shaft that has been in service, it may be necessary to repair it to correct size and condition. If the shaft is machined for the bearing seat, it is important not to leave machining ridges, even minute ones. The load very soon flattens down the tops of these ridges and leaves a fit that is loose and will rapidly become looser.



For best results, bearing seats should be ground to limits recommended for the bearing size and application.Shaft ShouldersCorrect shoulders are important because abutment against the shoulder squares the bearing. The bearing is actually squared up when it is pushed home against the shaft shoulder and no further adjustment is necessary. If a heavy thrust load against the shaft shoulders has occurred during operation, it is possible that the load may have caused the shoulder to burr and push over. Therefore, check the shoulder to make sure that it is still in good condition and square with the bearing seat. If it is not, the condition must be corrected before the spindle assembly operations are begun.Poor machining practices may result in shaft shoulders that do not permit proper bearing seating. The shoulder in Figure 7-23A is tapered. This results in poor seating of the bearing against the corner of the inner ring.



Thc shaft shoulder in Figure 7-23B is so low that the shoulder actually contacts the bearing corner rather than the locating face of the bearing. With the condition shown in Figure 7-23C, contact between the shoulder and the bearing face is not sufficient. Under heavy thrust loads, the shoulder might break down.Figure 7-23D is exaggerated to illustrate distortion of the inner ring when forced against off-square shoulder. An off-square bearing shortens bearing life.Some of these conditions can be corrected when repairs are made on the inner ring seat of the shaft. Such work should be done away from the clean assembly area to avoid possible contamination of the bearing and spindle parts by metal chips or particles from the machining or grinding operations.The shaft shoulder should not be too high as this would obstruct easy removal of the bearing from the shaft. As described previously, a pulling tool must be placed behind the inner ring and a surface must be left for the tool. Preferably, the inner ring should project somewhat beyond the shaft shoulder to permit pulling the bearing off against this surface.

This may not be possible in the case of shielded or sealed bearings where the bearing face is srrlall.Shaft Fillets and UndercutsDuring shaft repair work, it is important to pay attention to the fillet. When it is ground, the fillet frequently becomes larger as the wheel wears? causing an oversize fillet. This in turn locates the bearing on the corner radius instead of the shaft shoulder.

In other cases, the corner fillet is not properly blended with the bearing seat or shaft shoulder. This too may produce incorrect axial location of the bearing. The bearing corner radius originally may be a true 90" segment in the turning. but when the bores, OD's, and faces are ground off, it becomes a portion of a circle less than 90" while the shaft fillet may be a true radius (Figure 7-24A).Shaft fillet radius specifications are shown in bearing dimension tables with the heading “Radius in Inches” or “Corner Radius.” This dimension is not the actual corner radius of the bearing but is the maximum shaft fillet radius which the bearing will clear when mounted.
The radius should not exceed this dimension. The actual bearing corner is controlled so that the above mentioned maximum shaft fillet will always yield a slight clearance. Figure 7-24B illustrates the conventional fillet construction at the shaft shoulder. Where the shaft has adequate strength, an undercut or relief may bc preferred to a fillet. Various types are shown in Figure 7-24 C, D, and E. Where both shaft shoulder and bearing seat are ground, the angled type of undercut is preferred.
Break Corners to Prevent BurrsWhen the shaft shoulder or bearing seat is repaired by regrinding, it is desirable to break the corner on the shaft. This will help prevent burrs and nicks which may interfere with the proper seating of the inner ring face against the shaft shoulder (Figure 7-25). If left sharp, shoulder corners are easily nicked, producing raised portions which, in turn, may create an off-square condition in bearing location. The usual procedure to break a corner is to use a file or an abrasive stone.
This should be done while the shaft is still in grind position on the machine after regrinding the bearing seat and shoulders. The corner at the end of the bearing seat also should be broken, thus providing a lead to facilitate starting the bearing on the shaft.If nicks or burrs are found during an inspection and no other work is necessary on the shaft, they can be removed by careful use of a file or stone (Figure 7-26).
This work should be done elsewhere than in the clean assembly area. Any abrasive material should be removed from the part before returning it to the assembly area.Check Spindle Housing SurfacesIn many cases, housings will require as much preparatory attention as the shaft and other parts of the spindle. Check the surfaces which mate with the machine mount.
Frequently burrs and nicks will be evident and they must be removed before remounting the bearings. Failure to do so may cause a distortion in the bearing, resulting in poor operation and reduced life. These precautions apply to both bearing seats and shoulders.Shaft and Housing Shoulder

DiametersRecommended shaft and housing shoulders (Figure 7-27) for various sizes of bearings are shown in Table 74.Checking Shaft and Housing MeasurementsAfter all repair work on the shaft has been completed, shafts should be given a final check to make sure the repairs are accurate and within the recommended tolerances.
This work may be done with suitable gauging equipment such as an air gauge, ten-thousandths dial indicator, electronic comparator, an accurate micrometer, and other instruments as necessary. Accuracies of readings depend on the quality of equipment used, its precision, amplification; and the ability and care exercised by the operator. It is usually advisable to use a good set of centers which will hold the shaft and permit accurate rotation. The center points should be examined to make sure they are not scored and should be kept lubricated at all times to prevent possible corrosion.
Center holes of the shaft must also be of sufficient size, clean and smooth, and free from nicks. Be sure to remove particles of foreign matter that could change the centering of the shaft on the points.V-blocks will also be helpful to hold the shaft while making various checks. It is important that the V-blocks are clean on the area where the shaft contacts the blocks. Foreign matter and nicks will change the position of the shaft in the blocks and affect any measurements taken.Check Bearing Seat for Out-of-RoundA simple check may be made with a hand gauge on the bearing seat (Figure 7-28).
This will provide a reading at two points on the shaft 180" apart. However, it does not indicate how those points are related to other points on the shaft.For a more accurate reading on out-of-round (radial runout) of a bearing seat, mount the shaft between centers and place a suitable indicator in a position perpendicular to the axis of the shaft and contacing the bearing seat. On rotating the shaft slowly by hand, a check is obtained on all points of the shaft which the indicator contacts (Figure 7-29). Another method of measuring out-of-round is the three-point method using a set of V-blocks and a dial type indicator (Figure 7-30). The shaft should lay in the V-blocks and be rotated slowly while the indicator is centrally located between the points of shaft contact with the V-blocks and perpendicular to these lines of contact. This method will reveal outof-round which would not have been found by the two-point method of gauging. Therefore, if the equipment is available, it is desirable to check bearing seats using centers or V-blocks as well as two-point gauging. In all of these checks, the gauge should be placed in different locations on the bearing seat. This will give assurance that the seat is within the recommended tolerances in all areas. While the spindle is mounted on centers, the high point of eccentricity of the bearing seat should be located. Using a dial type indicator, find the point and mark it with a crayon so that it can be easily located when the bearing is to be remounted. The high point of eccentricity is covered in more detail later.
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