mywbut.com CONTENTS SL. TOPIC NO. 1. INTRODUCTION 2. FUNDAMANTALS OF VIBRATION 3. SOURCES OF VIBRATION 4. VIBRATION MEASUREMENT 5. VIBRATION ANALYSIS

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1 CONTENTS SL. TOPIC NO. 1. INTRODUCTION 2. FUNDAMANTALS OF VIBRATION 3. SOURCES OF VIBRATION 4. VIBRATION MEASUREMENT 5. VIBRATION ANALYSIS 1

2 1. INTRODUCTION Everyone in the course of our daily life encounters the phenomenon of vl 'bration. The effect of vibration is not only physically unpleasant but may also weaken the structure. It must therefore be regarded as a most undesirable condition, which must be eliminated for both comfort and safety. On the contrary, the vibration is often useful and may be essential in some application. Occasionally, for example vibration can be used to unmix things; as in sieves and other sorting devices, for conveying grain from one place to another, concrete will flow far -more readily into the furthermost recesses when it is poured into shuttering if it is suitably vibrated. Also vibration has got application in medical practice. For instance, it is used to massage away patients unwanted bulges and for removal of kidney stones. Large sums of money are spent nowadays on the study of various forms of vibration. The subject of vibration has acquired considerable importance, with the increasing pace of industrial and technological developments in the world over there has been a phenomenal increase in the speed and power of industrial machines. All devices which have mass and elasticity are capable of vibrating, however, rigid, they might seem. Whether it is desired to use vibration as a tool for failure and maintenance prediction or for using vibration control measure to avoid discomfort and failure, it is necessary to have a proper understanding of the subject. -This course material is concerned with fundamentals of vibration, sources of vibration, measurement of vibration and vibration analysis of rotating machines. 2. FUNDAMENTALS OF VIBRATION 2.0 INTRODUCTION The study of vibration is concerned w' ith oscillatory motions of bodies and the forces associated with them. All bodies possessing mass and elasticity are capable of vibration. Thus most engineering machines and structures experience vibration to some degree. The effects of vibration depend on the magnitude, frequency and duration of the vibration. Also, some times the vibration of a system emits lot of noise, which is harmful from human point of view. 2

3 2.1 WHAT IS VIBRATION Vibration is defined as the resp onse of an elastic system to a dynamic disturbance. There are two general classes of vibrations - free and forced. Free vibration takes place when a system oscillates under the action of forces inherent in the system itself, and when external impressed forces are absent. The system under free vibration will vibrate at one or more of its natural frequency, which is a property of dynamic system determined by its mass and stiffness distribution. Vibration that takes place under the excitation of external forces is called forced vibration. The simplest way to show vibration is to follow the motion of a weight suspended at the end of a spring as shown in figure 2. I. This is typical of all machines since, they too have weight and spring-like quality namely elasticity. Until a force is applied to the weight to cause it to move, we have no vibration. By applying an upward force, the weight would move upward, compressing the spring. If we release the weight, it would drop below its neutral position to some bottom limit of travel, where the spring would stop the weight. The weight would travel upward through the neutra position to tie top limit of motion, and then back again through the neutral position. This is vibration! This motion will dampen with time unless force is applied again. 3

4 2.2 CHARACTERSTICS OF VIBRATION A lot can be learned about a machine's condition and mechanical problems by simply not'mg its vibration characteristics. Refem'ng to the weight suspended on a spring, we can study the detailed,characteristics of vibration by plotting the movement of the weight against time. This plot is shown in figure 2.2. The simplest form of vibration motion is simple harmonic motion. The motion of the weight from its neutral position, to the top limit of travel back through the neutral position to the bottom limit of travel, and its return to the neutral position, represents one cycle of motion. This one cycle of motion has all the characteristics needed to measure the vibration. Continued motion of the weight will simply be repeating these characteristics. When the instantaneous displacement of the mass is plotted against time, the motion takes sinusoidal form as shown in figure. Fig: 2.2 CHARACTERSICS OF VIBRATION As vibrations are movements of the machines around a rest point, they may be quantified in terms of' displacement, velocity or acceleration. These characteristics of vibration are measured to determine.the amount of severity of the vibration. The displacement, velocity or acceleration of a vibration is often 17eferred to as the 'amplitude' of the vibration. 4

5 In terms of the operation of the machine, the vibration amplitude is the indicator used to determine how bad or good the operation of the machine may be. The greater the amplitude, the more severe the vibration DISPLACEMENT (PEAK TO PEAK) The total distance traveled by the vibrating part, from one extreme limit of travel to the other extreme limit of travel is referred to as the 'peak-to-peak displacement'. In Metric units, the peak-to-peak vibration displacement is usually expressed in microns, where one micron equals one-thousandth of a millimeter (0.001-mm). Peak-to-peak vibration displacement is sometimes expressed in mils, where 1 mil equals one thousandth of an inch (0.001 inch) VELOCITY (PEAK) Since the vibrating weight shown in the figure.2.2 is moving, it miist be moving at some speed- However, the speed of the weight is constantly changing. At the top limit of the motion the speed is zero since the weight must come to a stop before it can go in the opposite direction. The speed or velocity is greatest as the weight passes through the neutral position. The velocity of the motion is definitely a characteristic of the vibration but since it is constantly changing throughout the cycle, the highest or 'peak' velocity is selected for measurement. In Metric units, vibration velocity is expressed in millimeters per second peak. Vibration velocity is expressed in terms of inches per second peak for English or imperial units VELOCITY (RMS) The ISO in its work to establish internationally acceptable- units for measurement of machinery vibration decided to adopt VELOCITY (RMS) (root mean square) as the standard unit of measurement. This was decided in an attempt to derive criteria, which would determine an effective value for the varying function of velocity. It should be noted that IRD Mechanalysis instruments may be calibrated to read in -terms of VELOCITY (PEAK) or VELOCITY (RMS) ACCELERATION In discussing vibration velocity, we pointed out tfiat the velocity of the part approaches zero at the extreme limits of travel. Of course, each time that the part comes 5

6 to a stop at the limit of travel, it must 'accelerate' to pick-up speed as it travels towards the other extreme limit of travel. Vibration acceleration is another important characteristic of vibration. Technically, acceleration is the rate of change of velocity. Referring to the motion plot, figure 2.2, the acceleration of the part is maximum at the extreme limit of travel where the velocity is zero point 'A'. As the velocity of the part increases, the acceleration decreases. At point 'B', (the neutral position) the velocity is maximum and the acceleration is zero. As the part passes through the neutral point, it must now 'decelerate' as it approaches the other extreme limit of travel. At point 'C', acceleration is at peak. Vibration acceleration is normally expressed in "g's" peak, where one is the acceleration produced by the force of gravity at the surface of the C2 earth. By international agreement, the value of cm/se equals C2 C2 inches/se also equals feet/se has been chosen as the standard acceleration due to gravity. 2.3 CONVERSION OF AMPLITUDES The displacement, velocity and acceleration of a vibration are directly related. If the peak-to-peak displacement and frequency of a vibration are known, the velocity of vibration can be found as follows: - V Peak = 52.3D ( F / 1000 ) X Where: - V Peak = vibration velocity (mm/sec) peak D = vibration displacement (microns) peak to peak F = vibration frequency (CPM Further to the above when it is required to calculate vibration acceleration, the following formula can be used. - G (Peak ) = 5.6 D ( F / 1000 ) 2 X.0001 Where: - G (Peak ) D F = Vibration acceleration = Vibration displacement (microns) (peak-to -peak) = Vibration frequency (CPM) 6

7 It is sometimes necessary to convert Metric measurement to Imperial, or the converse. To convert velocity or displacement measurement from Metric to Imperial: - Velocity (mm/sec) Velocity (inches/sec) = 25.4 Displacement (microns) Displacement (mils) = 25.4 From Imperial to Metric: - Velocity (mm/sec) = Velocity (inches/sec) X 25.4 Displacement (microns) = Displacement (mils) X DISPLACEMENT, VELOCITY OR ACCELERATION WHICH SHOULD WE USE? Since the amplitude of vibration can be measured in terms of displacement, velocity or acceleration, the obvious question is 'Which parameter should we use? Vibration amplitude readings taken for checking overall machinery condition indicate the severity of the vibration. But which is the best indicator of vibration severity: displacement, velocity or acceleration? To answer this question, consider what happens when a wire or piece of sheet metal is bent repeatedly back and forth. Eventually, this repeated bending causes the metal to fai'i by fatigue in the area of the bend. This is similar in many respects to the way a machine or machine component fail from the repeated cycles of flexing caused by excessive vibration. Of course, the time required to fail the wire or sheet metal can be reduced by: - 1. Increasing the amount of the bend (displacement). The further the metal is bent each time, the more likely it is to fail. 2. By, increasing the rate of bending (frequency). Obviously, the more times per minute the metal is flexed, the quicker it will fail. Thus the severity of this bending action is a function of both how far the metal is bent (displacement) and how fast the metal is bent (frequency). Vibration severity then appears to be a function of displacement and frequency. 7

8 However, since vibration velocity is also a function of displacement and frequency it is reasonable to conclude that a measure of vibration velocity is a direct measure of vibration severity. Through experience we have found this to be basically true. Vibration velocity provides the best overall indicator of machinery condition. Displacement and acceleration readings are sometimes used to measure vibration severity. However, when displacement or acceleration is used, it is also necessary to know the frequency of the vibration. Charts like those shown in figure.2.3 and figure. 2.4 are often used to cross-reference the displacement or acceleration with frequency to determine the level of severity. Note from figure 2.3 that a displacement of 25 microns occurring at a frequency of 1200 CPM is in the 'GOOD' range, however, the same displacement of 25 microns at a frequency of 20,000 CPM is in the 'VERY ROUGH' range. Note also, that the diagonal lines dividing the zones of severity are constant velocity lines. in other words, a velocity of 12.7 mms per second peak is in the 'ROUGH' range regardless of the frequency of the vibration. Referring to the chart, figure 2.4, we can note that an acceleration of 1.0 g at a frequency of 100,000 CPM is in the 'GOOD' region of the chart; however, 1.0 g at a frequency of 18,000 CPM is in the 'SLIGHTLY ROUGH' region. So the real significance of the characteristics of vibration lies in the fact that they are used to detect and describe the unwanted motion of a machine. Each of the characteristics of vibration tells us something significant about the vibration. Therefore, the characteristics might be considered to be symptoms used to diagnose inefficient operation or impending trouble in a machine. 8

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11 2.5 VIBRATION FREQUENCY Frequency is the number of complete cycle in unit time. From the figure 2.2, the amount of time required to complete one cycle of vibration is the period of If a period of one second is required to complete one cycle of vibration, then during one minute the cycle will be repeated 50 times or 50 cycles per minute (CPM). The measure of the number of cycle for a given interval of time is the frequency of vibration and usually expressed in cycles per second or Hertz (CPS or Hz) or cycles per minute (CPM). 2.6 VIBRATION PHASE Phase is defined as the position of a vibrating part at a given instant with reference to a fixed part or another vibrating part. By measuring the phase we can Compare one vibration with another Determine how one part is vibrating relative to another part Phase readings are normally expressed in degrees (00 to 3600) where one complete cycle of vibration equeals Phase angle of vibrations, like amplitude and frequency, is a useful parameter, for analysis of vibrations. Measurement of phase and its analysis can help in the diagnosis of a machinery problem. Figure 2.5 shows the phase diagram of a vibrating object relative to a fixed reference, which corresponds to the equilibrium position. The phase diagram gives the corresponding to any position 2,3... etc., as shown, as measured from a datum.. Figure.2.6 shows the displacement time diagrams, A and B, of two vibrating par't-s or objects. PDF created with pdffactory Pro trial version 11

12 Fig: 2.6 PHASE DIFFRERANCE BETWEEN TWO VIBRATING PARTS The two reach their peaks or zero values, at different instants. The time difference, being td, phase angle between the two vibrating objects is td X 3600, since the time period corresponds to a full cycle or a phase of 360'. In the case of a rotor, the phase angle gives the location of the rotor at any instant e.g. it defines the location of the heavy spot of the rotor at each measurement point relative to a fixed point and is useful for balancing. 12

13 The phase may be measured with a stroboscope, as shown in figure 2.7. This is shown for a rotor rotating at same speed. If the frequency of flash of the stroboscope equals the running speed, any mark on the rotor appears stationary and the reading against a fixed reference scale would give the phase difference. 2.7 VIBRATION SEVERITY Since vibration amplitude (displacement, velocity or acceleration) is a measure of the severity of the trouble in a machine, the next question may be; 'how much vibration is too much?' To answer this question, it is important to keep in mind that our objective should be to use vibration checks to detect trouble in its early stages for scheduled correction. The goal is not to find out how much vibration a machine will stand before failure, but to get a fair, warning of impending trouble so it can be eliminated before failure. Absolute vibration tolerance or limits for any given machine are not possi 'ble. That is, it is impossible to select a vibration limit which, if exceeded, wi ill result in immediate machinery failure. The development of mechanical failure is just far too,complex for such limits to exist. However, it would be impossible to effectivel utilise vibration as an indicator of machinery condition y unless some guidelines are available 13

14 and the years of experience of those familiar with machinery and machinery vibration have provided some realistic guidelines. The vibration velocity provides a direct measure of machinery condition for the, intermediate vibration frequencies (600 to 60,000 CPM). The velocity values in figure 2.3 and figure, 2.4 are offered as a guide for overall unaltered velocity readings. When vibration amplitude is measured in displacement or acceleration, the charts in figure 2.3 and figure 2.4 may be used as guides in selecting acceptable levels of machinery vibration. Displacement and acceleration measurements applied to these charts should be filtered readings only. The guidelines offered in the above figures apply io machinery such as motors, fans, blowers, pumps and general rotating machinery where vibration does not directly influence the quality of a finished product. Amplitude readings should be those taken on the bearings or structure of the machine. Of course, the vibration tolerances suggested in these references will not be applicable to all machines. For example, some machines such as hammer mills or rock and coal crushers will inherently have high levels of vibration. Therefore, the values selected using these guides should be used,' only so long as experience, maintenance records and history proves them to be valid. For machines such as gr' ders and other precision machine tools where vibration can affect the quality of a finished product, refer the 'Guide to Vibration Tolerance For Machine Tools' provided in Table 2.1. Applying vibration tolerances to machine tools is rather easy because they can be based on the machine's ability to produce a certain size or finish tolerance. The values shown in the table are the result of years of experience with vibration analysis of machine tools, and represent the vibration levels for which satisfactory parts have been produced. Of course, these values may vary depending on specific size and finish tolerances required. A comparison of the normal pattern of vibration on the machine and the quality of finish, and size control required would reveal what level of vibration is acceptable. The first time the quality of finish or size control deteriorates, an unacceptable vibration level would be indicated. The initial values selected from Table 2.1 can then be modified to the new, more realistic ones. 14

15 Another severity standard which is coming into increasing use is ISO 2372 (BS 4675) as given in Table 2.2. This standard differs somewhat to the general severity standards referred to as it seeks to establish classifications of various types of machinery. Annexure-A, which follows the standard, describes the machines covered in the classification. To use ISO 2372 it is first necessary to classify the machine. Next reading across the chart can correlate the severity of the machine condition. The severity of the machine condition is indicated by the letter A 5 B, C or D. Making the decision to correct a condition of vibration is often a very difficult one indeed, especially when it involves downtime of critical machinery. Therefore, when establishing acceptable levels of machinery vibration, e erience and factors such as safety, labour costs downtime costs and the importance of a machine's operation to. the company's profits must be considered. Table-2.1 TENTATIVE GUIDE TO VIBATION TOLERANCES FOR MACHINE TOOLS TYPE OF MACHINE Displacement of vibration as read with pickup on spindle bearing housing in the direction of cut. Grinders Tolerance Range Thread Grinder 0.25 to 1.5 microns Profile of Contour Grinder 0.76 to 2.0 microns Cylindrical Grinder 0.76 to 2.5 microns Surface Grinder (vertical reading) 0.76 to 5.0 microns Gardner or Besly Type 1.3 to 5.0 microns Centreless 1.0 to 2.5 microns Boring Machine 1.5 to 2.5 microns Lathes 5.0 to 25.Omicrons 15

16 Table 2.2 ROLERANCES BASED ON B.s : 1976 (ISO 2372) 'A BASIS FOR COMPARATIVE EVALUATION OF VIBRATION IN MACHINERY' Velocity (mm/ s) RMS Velocity (mm/s) Peak Velocity Ins/Sec Peak (1) (2) (3) A Class I M/C Small A Class II M/C Medium B A Class III M/C Large B A Class IV M/C Turboo C B C B D C D C D D A= Good B = Acceptable C = Still Acceptable D = Not Acceptable ANNEXURE-A CLASSIFICATION EXAMPLES FROM BS 4675 (ISO 2372) (For guidance purposes only) In order to show how the recommended method of classification may be applied, examples of specific classes of machines are given below. It should be emphasised, however, that they are simply examples and it is recognised that other classifications are possible and may be substituted in accordance with the circumstances concerned. As and when circumstances permit, recommendations for acceptable levels of vibration severity for particular types of machines will be prepared. At present experience suggests that the following classes are appropriate for most applications CLASS I Individual parts of engines and machines integrally connected with the complete machine in its normal operating condition. (Electrical motors of up to kw are typi@al examples of machines in this category). CLASS II 16

17 Medium sized machines. (Typically electrical motors with 15 to 75 kw output) without special foundations, rigidly mounted engines or machines (upto 300 kw) on special foundations. CLASS III Large prime movers and other large machines with rotating masses mounted on rigid and heavy foundations, which are, relatively stiff in the direction of vibration measurement. CLASS IV Large prime movers and other large machines with rotating masses mounte( on foundations which are relatively soft in the direction-of,vibration measurement (for example turbo-generator sets, especially those with lightweight sub-structures). 3. SOURCES OF VIBTRATION 3.0 INTRODUCTION 3.0 INTRODUCTION In a machine, vibration is a result of minor faults that are the natural consequences of manufacturing and material limitations. Common causes of vibration are as follows UNBALANCE: This is a major contributor to vibration in rotating machinery. It is caused by unequal distribution of mass in a rotating part. Points of unbalance produce additional forces in the radial direction and the machine bearings restrain these forces resulting in vibration. The unbalance can be static or dynamic as shown in figure In either case the frequency of vibration equals the rotational frequency (IxRPM). Amplitudes of vibrations are excessive in the radial directions. Phase measurement by a stroboscope shows a single steady reference mark. 17

18 In the case of overhung rotors, vibrations in axial direction are also encountered due to unbalance, in addition to those in radial directions. Vibrations at all bearings, in such a case, would all be in phase. Sometimes, reasons other than unbalance may also result in vibrations at rotational frequency. In such cases, the unbalance as a possible cause should be confirmed by the difference in phase between vibrations in the two radial directions - horizontal and vertical. If the phase is 900,the cause of vibrations is unbalance. Defects like eccentric pulleys may also cause vibrations at the frequency of rotational speed. In such a case, the phase difference between vibrations, in the two directions, may not be 90' due to the effect of reaction forces of the belt. 3.2 MISALIGNMENT Like unbalance, misalignment is another common problem. Inspite of selfaligning bearings and flexible couplings, it is difficult to align the shafts and their bearings so that no force exists which will cause vibration. There are three possible types of coupling misalignment as shown in figure 3.2 Angular the center line of the two shafts meet at an angle Offset the shaft center lines are parallel but di 'laced from one another A combination of angular and offset misalignment. Misalignment, even with flexible, couplings, results in two forces, axial and radial, which result in axial and radial vibrations. This is true even when the misalignment is within the limits of flexibility of the coupling. The Magnitude of the forces and therefore the amount of vibration generated will increase with increased misalignment. The significant characteristic of vibration due to misalignment is that it will be in both the radial and axial directions. 18

19 Normally, the vibration frequency is lx RPM; however, when the misaligmnent is severe, second order (2xRPM) and sometimes third order (3x RPM) vibration frequencies may appear. There can be a misalignment' not involving a coupling. The misalignment of a bearing with its shaft is one example. In the case of a misaligned sleeve type bearing, no vibration will result unless there is also unbalance. In such a case radial vibration as well as an axial vibration will be present due to the reaction of the misaligned bearing to the force caused by the unbalance. If the real cause of this vibration is unbalance, then both the axial and radial readings will be reduced when the part is balanced. When an anti-friction bearing is misaligned with a shaft, then axial vibration will exist even When the part is balanced. We have to install the bearing properly to eliminate the vibration. The misalignment of sheaves and sprockets used in V-belt drives and chain drives results in high axial vibration. The angular and offset misalignment conditions result in destructive vibration and leading to accelerated wear of sheaves, sprockets, chains and drive belts. A bent shaft acts very much like angular misalignment, so its vibration. characteristics are included with misalignment. We can suspect misalignment or a bent shaft whenever the amplitude of axial vibration is greater than one-half of the highest radial (horizontal or vertical). 2.3 MECHANICAL LOOSENESS The causes of mechanical looseness could be loose mounting bolts, excessive bearing clearance or a crack in' the structure of bearing pedestal. For the vibration characteristic of mechanical looseness to occur, it needs some other exciting force to cause it, say, unbalance or misalignment. But, when the looseness is excessive, just - a small amount of unbalance or misalignment will result in large vibrations. So, looseness simply allows more vibration to occur than would otherwise appear. This does mean that if we can eliminate the unbalance or misaligmnent forces, we can reduce the vibration, but it needs an extremely fine level of balance or alignment, which may be impractical. So, removing the looseness is more practical. 19

20 Mechanical looseness leads to a heavy beating action and these cause a vibration at a frequency of twice the rotating speed (2xRPM) and higher, orders of the loose' part. The highest amplitude of vibration occurs at 2xRPM of the equipment. The nature of mechanical looseness and the reason for the characteristic vibration at 2xRPM can be explained as follows in figure 3.3. Fig: 3.3 MECHANICAL LOOSENESS There are two applied forces for each revolution of the shaft. One is applied by the rotating unbalance, the other when the bearing drops against the pedestal. Therefore, the vibration frequency is 2xRPM. We can view this with an oscilloscope attached to the vibration analyzer. Also there will be some clearances inherent in every' machine, and it is normal that some vibration will occur at a frequency of 2xRPM whenever some unbalance or misalignment is present. Generally, we should suspect mechanical looseness to be the problem whenever the seventy of vibration at 2xRPM is more than one half the severity of vibration at rotating speed (lxrpm). Moreover, if we have great difficulty in eliminating the vibration by balancing or realignment, we should verify whether there is any looseness. 3.4 BAD BELT DRIVES V-belt drives are popular'for power transmission because of their capacity to absorb shock and vibration. They are also quiet in operation when compared to chain or gear drives. But, for machine tools where very low levels 20

21 of vibration must be maintained, they can be the source of vibration beyond limits. Such problems are of two types. Belt reaction to other disturbing forces in the equipment Vibration due to actual belt problems. When we see the whip and flutter of V-belts (the flexible strands between the pulleys), we conclude that they are the source of vibration. Because of the ease with which the belts could be changed and because the vibration of belt is readily visible than other parts, belt replacement often happens to be the immediate solution. But remember excessive unbalanced eccentric pulleys, is misalignment or mechanical looseness, all these may result in belt vibration. The belt may be just an indicator of other disturbances in the equipment. Hence, before replacing drive belts, make an analysis to pinpoint the root causes.. Looking at the frequency of the vibration we can do this. If the belt is reacting to other disturbing forces in the machine, the frequency of belt vibration will most probably be the same as the disturbing frequency. When we are using the strobe light of the analyser, that part of the machine, which is actually generating the disturbing forces, will appear to stand still. For multi belt drives all belts should have equal tension. If not, the slack belts may cause excessive vibration even for very minor condition. Additional problems are belt slippage and rapid belt and pulley wear. Vibration from actual belt defects will be at 1,2,3 & 4 times belt RPM. The articular p frequency found would depend on the nature of the belt problem as well as the number of pulleys and idlers over which the belt must pass. To summarise, the vibration due to belt drives can -be reduced by the following methodes. Make sure belts are in good physical condition Check whether the number and size of belts meet the load requirements Use matched set of belts in multi-belt installations to get equal tension Verify whether the pulleys and sheaves are round and accurately aligned with one another Check for wear of pulley grooves. Too much wear will allow the belt to ride in the bottom of the groove, causing slippage and poor efficiency. 21

22 Verify whether the belts are properly installed and adjusted to proper tension as recommended by the manufacturer. Keep other disturbing forces in the machine to a minimum. 3.5 ECCENTRICITY Eccentricity means that the shaft (rotating) centerline is not the same as the rotor (,geometric) centerline. It is not out-of-roundness or ovality. Eccentricity is a common source of unbalance and results in more weight on one side of the rotating centerline than on the other side. As an example, when the bore of the inner race is not concentric with the inner race geometric centerline in an anti-friction bearing, an apparent unbalance in the part mounted on the bearing will be introduced. If we balance the rotor, the forces causing the vibration will be compensated and the vibration will disappear. That is why balancing a rotor in its own bearing is recommended. Different sources of eccentricity are, shown in figure

23 Fi2: 3.4 SOURCES OF ECCENTRICITY Though eccentricity can be corrected by routine balancing techniques, eccentricity can also result in reaction forces, which may not be corrected by simple balancing. In eccentric gear, the largest vibration will occur in the direction on a line through the centers of the two gears, at a frequency equal to I XRPM of the eccentric gear. It is not unbalance, though it will look like it. Eccentricity of the V-belt sheaves will result in reaction forces similar to the eccentric gear. Here, the largest vibration will occur in the direction of belt tension at a frequency equal to lx RPM of the eccentric sheaves. Again, the vibration looks like unbalance, but cannot be corrected by applying a balance correction. Eccentric fan, blower, pump and compressor rotors may also create forces, which result in vibration. Here, the forces are unequal aerodynamic and hydraulic forces against the rotor. These forces will be the greatest on the high side of the rotor, so will resemble unbalance. For this equipment, there is no positive test for eccentricity except that we can try to balance. If we get no results. verify if the impeller is concentric with the shaft journals. 3.6 FAULTY ANTI-FRICTION BEARINGS Rolling element bearings find many uses in today's machinery. They can be found in motors, slow-speed rollers, gas turbines, pumps, and many other machines. Some of the reasons for using the rolling element bearings are: low starting friction, low operating friction, ability to support loads at low speed (even zero), lower sensitivity to lubrication (compared to fluid film bearings so a simpler lubrication system can often be used) and the ability to support both radial and axial loads in the same bearing. Rolling element bearings have very little damping, so whenever a machine with. rolling element bearings traverses a balance resonance, large vibration can result. Also, compared to fluid film bearings, which generally have a long life, rolling element bearings have a limited fatigue life due to the repeated stresses involved in their normal use. Rolling element bearings, regardless of type (ball, cylindrical, spherical, tapered, or needle) consist of an inner and. outer race separated by the rolling elements, which are usually held in a cage. Mechanical flaws may develop on any of these components. 23

24 Using the basic geometry of a bearing, the fundamental frequencies generated by these flaws can be determined. The frequency of vibration caused by anti-friction is usually several times the rotating speed of the part, but it is unlikely to be an even multiple of shaft RPM. So, if we observe the rotating shaft with the strobe light, we may not see a stationary image (as. it would for vibration caused by unbalance, misalignment or gears', which occur at even multiples of shaft RPM) and also observe an unsteady frequency meter. Take the case of a bearing having a flat spot on only one ball. As the ball rolls, the flaw will intermittently come into contact with the bearing inner and outer races and will result in vibration at 1, and possibly, 2-times ball rolling frequency. Because the rolling frequency of the ball will be several times the RPM of the shaft, the resulting vibration will be high compared to rotating speed frequency. The amplitude of the vibration will depend on the extent of the bearing fault. In addition to the vibration occurring at or multiples of ball rolling frequency, these momentary impacts may excite vibration at natural frequency. Every object has its own unique natural frequency. A flaw on a rotating element of a bearing will produce the intermittent impacting type of force, which will cause the various parts (inner and outer races, shaft, and bearing housing) to vibrate at their respective natural frequencies. Normally, these will be high compared to the RPM of the machine. Hence, these vibration frequencies measured from a faulty bearing will also be high. Also, it is that these will be exact multiples of shaft RPM, Thus, the frequency of unlikely bearing vibration will probably not be a direct multiple of shaft RPM. Finally, there are many parts, hence many simultaneous vibration frequencies to varymg degrees, which cause the frequency meter to be unsteady or moving. 3.7 DEFECTIVE SLEEVE BEARINGS High levels of vibration or noise in sleeve bearings generally result from excessive bearing clearance (caused by wiping), looseness (babbitt loose in the housing), or lubrication problems. A sleeve bearing with excessive clearance may allow a relatively minor unbalance, misalignment or some other vibratory force to result in mechanical looseness or pounding. Here the bearing is not the actual cause; it simply allows more 24

25 vibration to occur than in the case where the bearing clearances were correct. A bearing, which has been wiped, can often be detected by comparing the horizontal and vertical amplitudes of vibration. Machines, which are securely mounted to a rigid foundation or structure, will normally reveal slightly higher amplitude of vibration in the horizontal direction. In several instances where the amplitude of vibration in the vertical direction appeared usually high compared to the horizontal, a wiped bearing was found to be the cause. Oil whirl is another problem associated with sleeve-type bearings. This' vibration occurs only on machines equipped with pressure-lubricated sleeve bearings and operating at relatively high speed - normally above the second critical speed of the rotor. Oil whirl vibration is often quite severe, but is easily recognized because the frequency is slightly less (5% to 8%) than one-half the RPM of the shaft. Under normal o eration, the shaft of the machine will rise up the side of p the bearing slightly, depending on the shaft RPM, rotor weight and oil pressure. The shaft operating at an eccentric position from the bearing center draws oil into a wedge to produce a pressurized load-carrying film. If the eccentricity of the shaft within the bearing is momentarily increased from its equilibrium position (say, as a result of a sudden surge, an external shock load or other' transient condition), additional oil will immediately be pumped into fill the space vacated by the shaft, thus increasing the oil film may drive the shaft into am pressure. This additional force developed by the oil film may drive the shaft into a whirling path around the bearing.. If the damping within the system is sufficient, the shaft will return to its normal position in the bearing; otherwise, the shaft will continue in a whirling path. Improper bearing design is normally attributed to the problem of oil whirl, however, excessive bearing wear, an increase in lubrication oil pressure or a change in oill viscosity are other possible causes. A temporary correction can some times be made by changing the temperature (viscosity) of the lubricant. Increasing the loading on the bearing by introducing a slight unbalance or misaligmnent is also some -crimes effective. Scrapping the sides. of the bearing or grooving the bearing surface to disrupt the lubricant wedge are also successful in some cases. 25

26 There are several special sleeve-beanng configurations to reduce the possibility of oil whirl. The axial-groove bearing is normally limited to smaller bearing applications such as those used in light gas turbines and turbo-chargers. The three-lobed bearing provides improved bearing stability against oil whirl. The three individual bearing surfaces generate pressurized oil films that act to center the shaft, Axial grooves are sometimes included at the intersection of the lobe segments to increase whirl resistance. The tilting pad bearing is a common choice on larger high-speed industrial, machinery. In a manner similar to the lobed-bearing, each segment or pad develops a pressurized oil wedge, which tends to center the shaft in the bearing. - The tilting feature allows each pad to follow the shaft, improving system damping and overall stability. Sometimes, a normal machine may exhibit oil whirl vibration. This may occur when an external source transmits vibration to the machine through the foundation or piping. If th is background vibration occurs at just the right frequency (ie. the probable oil whirl frequency of the machine) oil whirl will likely occur-. This condition is referred to as externally excited whirl. In a similar manner, a normally stable machine may be excited into oil whirl by -a foundation --or piping which is vibrating in resonance at a frequency equal to the probable oil whirl frequency. The resonant vibration of the piping or foundation may be the result of pulsation or flow turbulence. Oil whirl resulting from this condition is called resonant whirl. Whenever the vibration characteristic of oil whirl is found, we must carry out a complete vibration survey, of the installation including background sources, foundation and related piping to determine the true cause. Another problem encountered on machines equipped with sleeve bearings is called ffiction whirl or hysteresis whirl. It is similar to oil whirl except that'the vibration will occur on rotors operating above their first critical speed and the frequency of the vibration will always be the critical speed frequency of the rotor. For example, if a rotor operates at 3000 RPM and the first rotor critical speed is 2000 RPM. As is obvious, this vibration may not have the characteristic frequency of slightly less than 1/2 RPM associated with oil whirl. However, for machines operating above or near their second 26

27 critical speed, the frequency of hysteresis whirl may coincide with that of oil whirl resulting in an extremely severe vibration problem. In hysteresis or ffiction whirl, a rotor which, operates above critical speed will tend to deflect or bow in a direction opposite the unbalance heavy spot. As a result, the internal friction damping (hysteresis damping) of the rotor, which normally works to restrict deflection,. will be out of phase and this damping force will act to further deflect the rotor. This condition is normally kept in check by the damping provided by the bearings. However, if stationary damping is low,, compared to the internal dampin of the rotor, trouble is likely to occur. The usual solution. for hysteresis whirl is to increase stationary damping of the bearings and structure. We can change to a tilting pad bearing or other special bearing design. Sometimes, it can be solved by reduc'mg rotor damping for example, by replacing a gear-type coupling with a frictionless coupling such as a flexible disk coupling. Improper lubrication can also cause vibration in a sleeve bearing. If the bearing lacks lubrication or if the wrong lubricant is used, the result may be excessive ffiction between the rotating shaft and stationary bearing. This friction serves to excite vibration of the bearing and other related parts of the machine in a manner similar to the vibration we can generate by simply moistening our finger and rubbing it over a pane of glass. This vibration is called dry whip, is generally of high frequency and produces the distinctive squeal as for a dry bearing. The vibration frequencies generated are not likel y to occur at direct multiplies of-shaft RPM. Therefore, they will give no definite image under the strobe light and the vibration is similar to that caused by a faulty anti-ffiction bearing. Whenever vibration characteristic of dry whip is encountered, conduct an inspection of the lubricant, lubrication system and bearing clearances. This condition has been found on bearings with excessive and insufficient clearance. 3.8 GEAR PROBLEMS Common problems, which cause vibration in gear, are excessive gear wear, gear tooth inaccuracies, faulty lubrication and dirt or foreign material trapped in the gear teeth. Misaligmnent or a bent shaft can also be at fault. This is easy to identify because the 27

28 vibration normally occurs at a frequency equal to gear meshing frequency (the number of gear teeth x the RPM of the faulty gear). In complex gear arrangements where several meshing frequencies are possible, we have to examine the drawing of the gear box to determine the RPM and number of teeth on the various gears to identify which gear or gears are most likely at fault. However, if the axial vibration occurring at motor RPM frequency is relatively high on the gearbox and motor, misalignment may be the source of trouble. In this case, this misalignment condition should be corrected first; this may also eliminate the high frequency gear vibration. Sometimes, vibration at a frequency and equal to gear meshing frequency may be produced; for example, if a gear has only one broken or deformed tooth, a vibration at lx gear RPM may result. Viewing the vibration waveform on an oscilloscope connected to the analyzer will enable to differentiate this problem from unbalance because of the spike-like signal caused by a faulty gear tooth. If more than one tooth is deformed, a vibration frequency equal to the number of deformed teeth x gear RPM may result. An eccentrically mounted gear will also cause vibration at 1 x gear RPM, similar to unbalance. Where eccentricity is the problem, any attempt to balance in-situ will not be fruitful. Eccentricity, unbalance ind bent shafts have also caused gear vibration at submultiple frequencies of actual gear meshing frequency. The vibration amplitude and frequency from gears may also be erratic sometimes. This occurs with gears, which are operating under a very light load alcondition where the load may randomly shift back and forth from one gear to another. The impacts, which occur as the load is shifted, will excite the natural to frequencies of the gears, bearings and associated machine components. However, we can detect this gear vibration readily at two or more points on the machine and, thus we can distinguish from the bearing vibration which is predominate at the point of the faulty bearing. Because of the characteristic high frequency, gear vibration is also a common source of objectionable noise. For this reason, if we correct gear faults and other disturbances to reduce excessive gear vibration, noise level also will be reduced. 3.9 ELECTRICAL PROBLEMS 28

29 Vibration of rotating electrical machinery can be mechanical or electrical in origin. We have already seen mechanical problems. Electrical problems normally consist of unequal magnetic forces acting on the rotor or stator. These y unequal magnetic forces may be due to Rotor not round Eccentric armature journals Rotor and stator misaligned (rotor not centered in the sta,tor) Elliptical stator bore Broken bar Open or shorted windings The frequency of vibration will be lxrpm, and will appear similar to unbalance. An easy way to identify this source of vibration is to observe the change of vibration amplitude; the instant electrical power is disconnected from the unit. Make this check with the analyzer filter on the out position. If the vibration disappears the instant power is shut off; the vibration is likely due to electrical problems. If so, conventional, electrical testing procedures can be carried out to. pinpoint the true cause of the problem. If it decreases only gradually after power is removed, the problem is probably mechanical in nature. Electrical problems with induction motors often cause swinging or pulsating amplitude in nature. The blasting noise and vibration is caused by the slip frequency, characteristic of this type. Slip frequency = motor RPM synchronous frequency of the rotating magnetic field. The synchronous frequency is always equal to or an exact sub multiple of the AC. line frequency Therefore, if the motor has both electrical and mechanical problems such as unbalance, there will actually be two different vibration frequencies present. Since these two are close, their amplitudes will alternately add together and then subtract at a rate equal to the difference between their frequencies. The result will be a noticeable with steady beat as well as the corresponding swing of the amplitude meter. If the amplitude of this pulsating vibration is excessive, we must correct it. If we observe the amplitude meter, the instant the power is shut off, we can decide whether it is mechanical or electrical cause. The pulsation may not be detrimental to the performance 29

30 of the machine, but a pulsating noise is more noticeable than a steady one and affects the personnel psychologically. Electric motors have inherent vibration due to torque pulses. These torque pulses are generated as the motor's rotating magnetic field energizes the poles in the stator. The frequency of vibration resulting from torque pulses will be 2 times the AC line frequency powering the motor. Thus, if AC line frequency is 50 Hz (50 cycles per second) or 3000 CPM, torque pulse frequency will be 6000 CPM. This is rarely troublesome except where extremely low vibration levels are required, or if the torque pulses should behappen to excite a resonant condition in the machine or structure. If resonance is excited, this can also result in excessive noise. In eccentric motor armature though the armature itself may be balanced or in terms of rotor weight distribution, a lxrpm force is generated between the armature and stator because of varying magnetic attraction between the eccentric armature and motor poles. Increased load increases the magnetic field strength and results in increased vibration. To check this, measure the vibration, with the motor operating under power. Then, turn the power off and observe what happens to the amplitude of vibration. If the amplitude decreases gradually as the motor coasts down, the problem is likely unbalance. On the other hand, if the vibration amplitude disappears the instant power is turned off, the problem is electrical and possibly due to armature eccentricity. Other electrical problems causing vibration are shorted windings, broken rotor bars, or a rotor, which is not properly centered in the stator. A visual inspection using standard motor testing procedures will reveal the nature of the electrical problem RESONANCE We already know that every object and every part of a machine has natural frequency. If we strike a bell, it vibrates at its own natural frequency. This continued vibration, called free vibration, will eventually diminish because of inherent damping In addition to free vibration, there are forced vibrations where the frequency depends on the frequency of the driving force applied to the machine or structure. For example, the driving force of rotor unbalance 30

31 may cause the forced vibration of a motor. In such case, the frequency of this forced vibration is determined by the speed (RPM) of the motor. To confirm whether or not a part is vibrating in resonance, we can apply the bump test.. With the machine shut down, simply bump the machine or structure with a force sufficient to cause it to vibrate. Since an object 'will undergo free vibration at its natural frequency when bumped or struck, the frequency of free vibration generated in this way will be indicated on the analyzer's frequency meter. The analyzer's filter must be in the out position for this test. If the vibration diminishes very quickly, it may be necessary to bump the machine several times in succession m order to sustain free vibration long enough to register on the frequency meter. We can also record the amplitude and phase of vibration versus the rotating speed of the machine. We can use FFr analysis or make@ a similar plot by operating the machine at a number of selected speeds and plotting the amplitude and phase of vibration for each speed. If resonant conditions do exist, they will be clearly identified by a characteristic peak vibration and by a large phase shift (around 180'). There are several ways to correct a resonance problem. We can change the frequency of the exciting force so that it no longer coincides with the natural frequency of the machine or structure. Either increasing or decreasing the RPM of the machine can do this. If the exciting frequency cannot be changed, change the natural frequency - by increasing or decreasing the stiffness or mass of the, object AERODYNAMIC AND HYDRAULIC FORCES Machine which handle fluids will have vibration and noise due to the reaction of the vanes or blades on the irnpeller striking the fluid. This type of vibration is common on pumps, fans and blowers. The frequency will be equal to the number of vanes or blades on the impeller times the RPM of the machine. Aerodynamic and hydraulic vibrations are rarely troublesome unless they excite some part of the machine, piping or ductwork to vibrate at resonance. When we encounter this type of vibration, carry out the tests for resonance to determine which part of the machine is causing the problem. If no resonance condition can be found, the problem may be due to improper design of the machine or related piping or du'ctwork. 31