Condition diagnosis

Units which fail to meet acceptable criteria should be given a complete frequency analysis. This applies to all weights and speeds of machine. Again readings should betaken at the vari­ous positions. A typical frequency analysis is shown in Figure

15.13.

So far we have discussed vibration in a general sense and indi­cated permissible overall limits. For important and/or arduous applications however, we need to be able to identify the causes of vibration and their likely effects on the machine which could be catastrophic in the event of a total breakdown.

FAN VIBRATION SIGNATURE To ol — 1 V < . -‘c _ 1 -t M — I U . * ‘ " °l T — § jjj _ 5 •? a — Z. 8 < • < ° ■ 03 cc ^r -X ~r| : S y — < — O ■ Fsj — Њ — O — _ — — _ ? I-‘ W FF Ec DUG NC ;y * CPK A — "it

500 6 8 IK 2K 3K 4K 5K 6 0 !0K 20K 30K 40K 50K 6 8100K 200K 300K 400K 500K MACHINE +50„„ Fun Rvf>tt"c*UP ЦVEL □ ACCEL C|N/C MEASURE POSITION FAN BNP ФE. AR1N6 MACHINE LOCATION C i_ / , , Oupportea on oyton AMPLITUDE UNITS Q MILLS f*K PK Q W/StC PK O QPK | 1 MICHONS PH PK (7] MWSPC PK JDЈT/J, ft DATA SHEET no OPERATING CONDITIONS ^ , (BPM. LOAO. TeMP. nc) FREE A/& — JS50 / / f ev

The keys to the identification of the cause of a vibration are in its magnitude and frequency over most of the spectrum. Below about 10Hz displacement will be of primary importance, whilst above about 1kHz, acceleration is paramount. Over the re­mainder of the range, velocity measurements will be sufficient.

Different causes of vibration occur at different frequencies. For example, a faulty ball bearing would cause high frequency vi­bration at many times the rotational frequency whilst unbalance or misalignment produce vibration at the rotating speed fre­quency. To rely on an overall or linear response reading of vi­bration velocity could lead one to ignore a developing problem.

To obtain the installation’s “vibration signature”, a pick-up is used, which can either be handheld or more rigidly attached, feeding a meter giving a visual display. For analysis some type of tunable frequency analyser is necessary together with a stro­boscopic light. The strobe permits rapid tuning to rotational speed when it “views” the rotating element and apparently sees a stationary image. From the analyser an X-Y recorder can be fed to show the magnitude of the vibration in narrow frequency bands right across the spectrum which the analyser can iden­tify. By fixing the probe in either the vertical, horizontal or axial modes, different traces can be obtained which will inform the operator as to possible sources of vibration. Any “peaks” in the reading at specified frequencies will indicate the onset of cer­tain troubles.

It is to these discrete frequencies we now look. Remember vi­bration severity is a quality judgement whilst frequency will indi­cate the cause. An increase in the severity of a particular fre­quency during inspection, commissioning or operation may indicate the onset of a particular problem. By referring to the ini­tial signature and specifying that the particular frequency level should not increase by more than a predetermined amount, it is possible to construct planned maintenance procedures.

In the comments which follow fl7 is defined as rotational fre­quency and equals (rev/min + 60) Hz.

The machine in general

Unbalance As the heavy spot would give a “pulse” to the pick-up once every revolution, unbalance will be identified by
high readings in the horizontal and vertical directions at the ro­tational frequency, i. e. f|Hz. This is the most common cause of vibration.

High readings at commissioning can indicate residual unbal­ance in manufacture or “sag” if the rotating element has not been turned over regularly during storage on site. The build up of dust on a rotor, erosion or corrosion will also lead to increas­ing figures. Scaling at high temperature (above 400°C) or soak­ing in heat whilst stationary are other possible causes. 1

Misalignment This is almost as common as unbalance despite the use of self-aligning bearings, which should still be lined up as well as possible. Flexible couplings can be out of line both by height and angle. A bent shaft produces angular misalignment. Radial and axial forces are always produced, the size of such forces, and therefore the vibration which results being propor­tional to this misalignment.

As previously stated, the axial readings are usually 50% or more of the radial readings and again the frequency is normally f-) Hz. When the misalignment is severe however vibration at 2^Hz and even 3fiHz may be experienced. Misalignment can also occur where a machine has been distorted by tightening down onto foundations which themselves are not level. With sleeve bearings this will produce vibration according to the amount of residual unbalance, but with ball or roller bearings an axial vibration would be produced even if the unit were “per­fectly” balanced, which is physically impossible.

Another very common fault is when pulleys and ropes of vee belt drives are not correctly aligned. This results not only in de­structive vibration, but also leads to rapid wear of belts and pro­duction of frictional heat through to shafts and bearings.

Eccentricity An example of this could be where an impeller centre with excessive bore is “pushed over” by a taper key. The centre of rotation does not then coincide with the geometric centre. As far as the impeller is concerned, this leads to more mass being on one side of the rotational centre than the other,

I. e. unbalance.

It can therefore be corrected by rebalancing provided that the rebalancing takes place in its own shaft and bearings and that with ball/roller bearings the position of the inner race on the shaft also does not change. The predominant frequency is of course f|Hz. Where a fan is gear driven, eccentricity can pro­duce reaction forces between pinions because of the cam-line action.

The largest vibration will occur along a line joining the centres of the two pinions at a frequency equal to (pinion rev/min h — 60) Hz of the one which is eccentric. It will look as if it is unbalanced but cannot be corrected by re-balancing. A similar situation can arise with vee belt pulleys which are eccentric. The largest vi­bration will be in the direction of belt tension at a frequency of (pulley rev/min ^ 60) Hz of the eccentric pulley. Again re-bal­ancing cannot cure.

Looseness Common forms are loose foundation bolts and ex­cessive bearing clearances. It will not be manifest unless there is some exciting force such as unbalance or misalignment to encourage it. Only small forces are necessary however to ex­cite the looseness and produce large vibrations. Although rebalancing or realignment may therefore assist, extreme ac­curacy would be necessary which may be impossible to achieve. It is essential to tackle the problem at its source.

1, 2,

Equ 15.3

Where

Fp =

To determine the characteristic frequency of looseness, let us consider an unbalanced rotor fitted to a shaft running in a bear­ing with loose holding down bolts. When the heavy spot is downward, the bearing will be forced against its pedestal. When the heavy spot is upward it will lift the bearing whilst at po­sitions 90° away, the force will neither lift nor hold down, and the bearing will drop against the pedestal due to weight alone. Thus there are two applied forces each revolution of the shaft and the vibration frequency is 2fi Hz. This is the characteristic frequency of looseness.

Resonance The section on fan response (Section 15.4), showed that every object has a natural frequency at which it “likes” to vibrate. Should the forcing frequency coincide with the natural frequency of a part, then resonance will occur. To over­come the problem with a vee belt-driven unit is simple, a small change in rev/min will normally suffice. With a direct drive unit stiffening or a change in the design may be indicated, although unlikely.

Many of these problems can only be identified at the commis­sioning stage or during service. Figure 15.14 shows a techni­cian taking readings on site.

I

Figure 15.14 Site testing with hand-held vibration analyser Courtesy of Schenck RoTec GmbH

Specific vee belt drive problems

Many of the problems found in impellers will also be present in vee belt drives. Often the balancing of pulleys has been over­looked and must be specified when ordering. Misalignment has already been mentioned.

In Chapter 14 a number of problems were identified, which could result in additional noise. However as these problems were essentially mechanical in origin, they are also manifest as vibration.

Such drives have good resistance to shock and vibration but may be blamed for causing trouble as they can be readily seen to whip and flutter especially when the belts are unmatched. Belts are often changed unnecessarily when the fault is really that of unbalance, misalignment etc. Nevertheless, the impor­tance of using matched sets of belts cannot be emphasised enough.

Vibration from faults in the belts themselves occur at multiples of belt speed. The relevant frequencies are:

„ . pulley diameter, ..

3or4x—————- xf Hz

Belt length

Pulley rev/min 60

Likely faults are pieces broken off, hard or soft spots etc.

L

Motor rev/min 60

Where

F3=v

1 + — cos A D

Hz

F2 — fi x ^

1 — cos A D

Hz

Hz

1—r-cos A D2

Hz

1 — cos A D

Line frequency , slip frequency = 2 x————————————— r-i-:——fm Hz

No. of poles

Faults in pulleys, such as chipped grooves etc., will be identified at the speed of the relevant pulley fp Hz. Electric motor problems Most electric motor vibrational problems are mechanical in ori­gin e. g. unbalance, misalignment, bolting down to foundations which are not level, loose foundation bolts, faulty bearings etc. Previously described frequencies are therefore applicable us­ing, of course, the motor rev/min where this differs from the ma­chine rev/min. Again the noise sources identified in Chapter 14 will also be identified as vibration. With induction motors, forces act in the air gap between rotor and stator tending to pull these together and produce vibration at 2 x line frequency Hz. Normally such vibration is small except in 2-pole motors, but if the air gap varies, or if the tightness of stator laminations or winding in the stator vary, then this vibra­tion will increase considerably. The second and third harmon­ics may also be important. Generally,
Equ 15.5
Flaw in inner raceway:
	Equ 15.6
Flaw in ball or roller:
	Equ 15.7
Irregularity in cage or rough spot on ball/roller:
	Equ 15.8
Where: N D D A
= number of balls or rollers = diameter of balls or rollers = pitch circle diameter of race = angle of contact of ball/roller
Equ 15.4
Such vibrations are not easily transmitted to the rest of the fan (except where there are large flat mounting surfaces) and will therefore be recognised by velocity readings on the bearing housing. Severe misalignment of a race will sometimes result in a fre­quency at n xf| Hz, even when the bearing itself is satisfactory. Selection and life of rolling element bearings Bearing parameters Modern ball and roller bearings are a precision made item. With correct selection, installation and lubrication, premature failure is unlikely. When this does happen it has usually been caused by machine out-of-balance, misalignment or use at speeds/ loads/temperatures in excess of those recommended by the manufacturers. The demand for high quality and low price, necessitates quan­tity production of all anti-friction bearings. Machine designers then have to select from a standard range the items which most closely meet their requirements as to: • Dimensional and speed properties • Frictional drag and heat generated • Noise output • Deflection under load • Rate of wear and lubrication • Life in relation to load Of these, the life is of most importance, especially at moderate speeds and loads. Correct selection for life will usually ensure that performance under the other headings is acceptable. Fatigue life Earlier in Fans & Ventilation, we considered a rolling element bearing to have point or line contact between the raceways and the ball or roller. In realitytheseconditionscannotexistwherea load is applied since the smallest force would induce an infinite stress. Deformation therefore takes place and the contact is over an area sufficiently large to result in a stress value which can be accepted by the bearing materials. To ensure that the stress is within the elastic limit, and to keep the contact area to a minimum, the steels used are through hardened. Accordingly high stresses still result, and the major cause of failure becomes metal fatigue.
This will not in itself be important as it will be of very low fre­quency. However, its interaction with higher frequencies can produce pulsations. If the rotor is severely unbalanced, the high spot will come closer to the stator than other points. As it passes the stator poles more pull is exerted. Thus vibration occurs at 2 x slip fre­quency on a 2-pole motor, 4 x on a 4-pole motor and so on. The magnitude of those readings in these frequencies can indicate whether the problem is simply due to the lack of balance, change in the air gap, worn journals, broken rotor bars etc. Vibrations may be produced at a frequency equal to no. of rotor bars x f-i Hz and at no. of stator slots x U Hz. Vibrations at inter­active frequencies may also be important. If a resonance condition exists within the motor at line fre­quency, then large vibrations can be produced. More often however this is the fault of an unbalanced magnetic pull and can be cured by changing stator connections. With all suspected electrical sources of vibration, the simple check is to switch off the motor when they should “die”. The specific problems of bearings Sleeve bearings Problems with these generally result from ex­cessive clearance, wiping, erosion of the journal surfaces (e. g. builders’ dust on site entering the bearings before start up), looseness of the white metal, inadequate lubrication (poor maintenance), lubrication with an incorrect grade of oil, or chemical corrosion. Characteristic frequencies are fi Hz, 2fi Hz or random, for the reasons already given. It will be appreciated that some of these problems are prevalent under modern conditions, becoming especially important on high speed fans, and have encouraged the trend to ball and roller bearings. Ball and roller bearings Races which have flaws on the balls, rollers or raceways will not only cause additional noise but also high frequency vibration identified in Chapter 14 but repeated here as follows: Flaw in outer raceway or variation in stiffness around the hous­ing:

B-

T

0.01

The time at which the first fatigue crater appears cannot be measured precisely. A batch of apparently identical bearings run under the same conditions of speed, load, lubrication and temperature will fail at different times. By using a statistical ap­proach and analysing data accumulated over the years, it is possible for the manufacturer to quote the probability that a bearing running under the specified conditions will last for a given period of time, but this cannot be predicted with certainty.

The calculation of bearing life has now been made the subject of a Standard, namely ISO 281. The L10 life in hours or running is defined as that at which 10% of a group of apparently identi­cal bearings can be expected to have failed by rolling fatigue. Being a statistical forecast, the result will be more accurate, the greater the numbers tried. Conversely 90% of all bearings can be expected to exceed their L10 lives, whilst the average life should be five times as great.

The need for early warning techniques

With such a wide spread of hours to failure, it has become desir­able to monitor bearings, so that we may predict their lives much more accurately. When a bearing does fail, the damage to associated machine parts, and the production losses, are of­ten far in excess of the actual cost of replacement.

The characteristic of fatigue failure is to increase the very high frequency vibrations at between 2.5 kHz and 80 kHz. Whilst these will always be present at a low level in nominally perfect bearings, they may be expected to increase by a factor of many hundreds before the onset of complete failure. During this time, the vibrations at the lower frequencies related to rotation and its multiples may not increase very much or may be attributed to other causes. Many vibration analyzers have a maximum cut-off frequency of about 1 kHz and by the time they are able to detect a significant increase in this vibration level, failure due to fatigue may be imminent.

Other techniques have therefore been developed, and these all monitor the high frequency vibrations in some form or another. It is to these methods which we now turn, describing the fea­tures and advantages of each. It is left to you to detect by inference or omission their respective disadvantages!

Posted in Fans Ventilation A Practical Guide