pva course material 03.08.10

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Practical Vibration Analysis

1 Reliability Maintenance Institute

1.0 INTRODUCTION TO VIBRATION ANALYSIS

1.1 INTRODUCTION Machinery Vibration or Vibrations are manifestation of the health condition of operating machines. The accurate measurement and correct interpretation of vibrations can help in precisely diagnosing machinery problems during operation. The use of Vibration Analysis on rotating machines is an accepted and effective method to assess :

1. Presence of minor inaccuracies, if any 2. To prepare corrective action plans 3. To hand over the machine in a good health condition for a continuous operation 4. To prepare an effective Dynamic Predictive Maintenance Programme

Vibration is the language of machines which if one can listen and understand is enough to diagnose their complaints and ailments. Vibration is not related with the direct performance of a machine. Machinery vibrations are produced due to defects and inaccuracies in the machine components or in the machine assembly. An ideal machine (a machine with Zero Defect) will not vibrate at all, and all the energy supplied to the machine would be channeled into useful work. Vibrations are harmful to the machinery health and therefore all efforts are to be taken up to enter the vibration levels on all operating machines. But, even a very healthy machine will produce vibrations due to the minor imperfections in the system. Such vibrations can be usefully measured and analysed to understand the health condition, to pinpoint and to identify the machinery defects. Machinery vibration is an indication of operating Machine’s Health. If the vibration amplitudes are high, the health of the machine is not good. If the vibration of a machine is smooth, it can be understood that the defects and inaccuracies in that machine are very minimum and it’s health can be considered as “Good”. Machinery vibration can be considered as the ‘Heart Beat’ of a machine. The vibration pattern depends on it’s health and the monitoring of machinery vibration will give a clear indication of the defects in each machine.

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MACHINE VIBRATION

Machinery vibration is, the movement of a machine part, “back and forth” from its neutral (normal) position. This back and forth movement of the machine part is caused by forces being generated in the machine. These vibratory forces are created by defects in machines. A simplest method to understand machinery vibration is to consider the motion of a “Spring-Mass system” as shown in Fig.1 .2

Fig no: 1.2 SPRING MASS SYSTEM

IMPORTANCE OF MACHINERY VIBRATION Machinery vibration is very important to understand the health condition of a machine because :

Ø All machines (which are in operation) will vibrate.

If we say that any operating machine is not vibrating, it means that the health condition of the machine is excellent and thereby the vibration amplitudes are very low.

Ø Machines vibrate because of imperfections. We, as human beings have not learned to make a perfect machine. Minor imperfections are always permissible and that is why , we have tolerances, permissible limits etc., for various parameters. When the imperfection is more, we call it a “Defect”. That means machines are vibrating because of “Defects”.

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Ø There are various defects (causes) that can produce vibrations. The vibration characteristics of different defects will be different in nature. Therefore, a careful understanding of the vibration characteristics will help in identifying the defects.

The method of vibration measurement and analysis is based on the three basic principles i.e., Ø The measure of vibration is an indication of machinery health Ø Machinery vibrations are caused due to defects in the machine Ø Different defects produce vibrations of different characteristics

The basic purpose of vibration measurement and analysis is to identify machinery health condition and detect various defects. Therefore, the first stage is to understand the various defects that can produce vibrations in plant machines. VIBRATION ANALYSIS Vibration Analysis is the measurement and interpretation of machinery vibrations. The purpose of vibration analysis are Ø To identify Health Condition Ø To detect the Defects and Inaccuracies Ø To Pinpoint the Defective locations of any rotating machine.

Therefore, in practical sense, the analysis involves answering the following three basic questions. Ø How much vibration is present on the machine ? Ø What are the defects in the machine ? Ø Where are the defects in the machine ?

These questions can be answered by measuring and interpreting the vibration characteristics.

IMPORTANT VIBRATION CHARACTERISTICS The vibration characteristics can be learned by considering a vibrating “Spring-mass” system. A Spring-mass is selected for understanding vibratory motion of a body, because all vibrating parts of a machine have Ø Mass Ø Springness in the system Ø Damping forces

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Example :- We can consider one of the Fans in a plant supported by two Bearings. Each Bearing supports certain weight of the Fan. Depending on the construction, lubrication type, bearings etc., there is always an effect of springness and damping in the system. During the operation of the machine, forces are generated in the system. Therefore, a vibrating machine is compared to a spring mass system as illustrated in the figure 1. It could be seen that , vibration of the Mass starts from the Neutral position by the application of a Force. It can be noted that until the force is applied on the weight to cause it to move, the mass will not vibrate. If we apply a force in the upward direction, the mass will move upwards and the spring will be compressed. If we release the mass, then it would drop downwards. Now the movement of the mass shall be in the downward direction and it will pass through the neutral position. Due to the effect of the spring, the mass will stop at a bottom position , known as its bottom limit of travel. The mass will then travel upwards through the neutral position to the top limit of motion and then back again through the neutral position. This motion will continue to occur in the same manner , as long as the force is reapplied on the mass. This one cycle of motion from its neutral position to the top limit of travel and to its bottom limit of travel and then back to the neutral position is known as “One Cycle of Vibration”. Now, we can consider this One Cycle of Vibration for learning various vibration characteristics. Let us consider the plot of One Cycle of Movement of the vibrating part against time. This plot is shown in Fig.1.5. The time taken to complete one cycle of vibration is called “The period of Vibration”. If a Period of One Second is required to complete One Cycle of Vibrations during One Minute, the vibration cycle will be repeated 60 times. Therefore, we can see that the vibration is occurring at 60 cycles per Minute. This is a measure of the number of cycles during a given interval of time and it is known as “Frequency of vibration”. For the practical purpose, of vibration measurement and analysis, it is always preferred to specify the vibration frequency in Cycles Per Minute (CPM) because we specify machine speeds in RPM. The important vibration characteristics are

• Vibration Displacement • Vibration Velocity • Vibration Acceleration • Vibration Frequency • Vibration Phase.

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Fig no:1.5 CHARACTERISTICS OF VIBRATION

Each of the characteristics are explained below:

Vibration Displacement The total distance moved by the vibrating part from one extreme limit to the other extreme limit of travel is defined as “Peak to Peak Displacement”. The peak to peak displacement is normally measured in Microns, where One Micron is called 1/1000 of a Millimeter. 1 Micron = 1/1000 mm or 1 Micron = One Millionth of a Meter

10 = 0.001 mm 1.5.1.1 Use of Vibration Displacement The use of vibration displacement, velocity and acceleration can be defined by considering the failure mode of various machinery part. A slow speed equipment normally fails due to excessive stress. The stress is proportionate to the deflection from its normal path. In terms of vibration, the displacement is a measure of deflection. Therefore, wherever stress concentrated failures are likely; the characteristic to be used for assessing severity is the “vibration displacement”.

Fig no: 1.5.1 VIBRATION DISPLACEMENT IN SPRING MASS SYSTEM

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1.5.2 Vibration Velocity The vibrating part is moving up and down as explained above. This movement is with some speed, which is continuously changing. At the top limit of travel, the speed is Zero, because the mass has to stop before it moves in the downward direction. The Speed or Velocity will be maximum as the part passes through the neutral position. The velocity of motion is a definite characteristic of vibration, but the velocity is continuously changing through out the cycle. The vibrating part will have the maximum or peak velocity while passing through the neutral position. This peak velocity is selected as one of the major vibration characteristics. The vibration velocity is measured in Millimeter per Second. The vibration velocity can also be expressed as Root Mean Square Velocity. 1.5.2.1 Root Mean Square Value The Root Mean Square value gives a direct relation to the energy content in vibration and therefore, the destructive abilities of vibration can be understood from the RMS value. The permissible limits of vibration is specified in terms of RMS velocity as per ISO 10816. 1.5.2.2Use of Vibration Velocity Most of the medium speed equipments fail due to fatigue. Fatigue is proportionate to the vibration displacement and the number of vibratory cycles. If ‘X’ is the displacement and ‘ω’ is the Angular velocity; the vibration velocity is equal to ‘X’ multiplied by ‘ω’ (Omega).

ω= 2 Πn / 60, where N is the rpm. In terms of vibrations N can also be considered as frequency. Therefore, the vibration velocity is proportionate to the vibration displacement and the vibration frequency.

Fig no: 1.5.2 VIBRATION VELOCITY IN SPRING MASS SYSTEM

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1.5.3 Vibration Acceleration

As we have seen in the vibratory movement, the vibration velocity of the part is falling between Zero at the extreme limit of travel to maximum value at the neutral position. Since, the part has to stop at the limit of travel and then to pick up speed to obtain a maximum value at the neutral position, it must accelerate to pick up speed, as the part travels towards the other limit of travel. This acceleration is another important characteristic of vibration. Acceleration can be defined as “The rate of Change of Velocity”. Referring to the motion plot indicated in Fig no: 1.5.3 acceleration of the vibratory part is maximum at the extreme limit of travel, where the velocity is Zero. As the velocity of the part increases, the acceleration decreases. At the neutral position, the velocity attains a maximum value and acceleration will decrease to Zero value. As the part passes through the neutral point, the velocity will decrease and hence the motion is decelerated. When the part approaches the other extreme limit of travel, the acceleration will be at its peak once again . Vibration acceleration is normally expressed in “g” peak, where 1g is acceleration produced by the force of gravity at the surface of the earth. The value of acceleration due to gravity is equal to 980.665 Centimeter per second .

Fig no: 1.5.3 Vibration Acceleration In Spring Mass System

1.5.3.1 Use of Acceleration In very high speed equipments, the failure is normally due to excessive force. As we know, F = m x a where m = Mass of the vibrating part ; a = The vibration acceleration; F = Force applied Normally, the mass of the vibrating part remains constant and varying parameter is the vibration acceleration. Therefore the force is directly proportional to vibration acceleration.

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Normally, acceleration measurements are used to determine the health of very high speed equipments. The displacement, Velocity and Acceleration will determine the quantity or the severity of the vibration . Therefore , these three characteristics are commonly known as “Vibration Amplitudes”. 1.5.4 Frequency The Spring-mass system illustrated in the Fig no: 1 is indicating a cyclic motion. Each cycle is completed in a definite time. If the machine takes one second to complete one cycle , it will complete 60 cycles per minute. This is called the Vibration Frequency. Vibration frequency is the number of completed cycles of vibration in unit time. This can be expressed either in cycles per second or in cycles per minute. In relation to vibration analysis, it is advantages to use the unit of frequency as cycles per minute, because we always measure the speed of the machine in rpm. (Revolutions Per Minute). The period of vibration is a simple and very meaningful characteristic which is very important in machinery problem diagnostics. The vibration frequency, f = 1/T period. Where T = time period of completion in one second, the frequency is equal to 1=1/T cycle per second. The frequency in cycles per minute, is 60 cycles per minute. Fig no: 1.5.4, visualize the vibration frequency with respect to the Spring mass system.

Fig no: 1.5.4

As explained earlier, it is always easily understandable to refer vibration frequency in cycles per minute, especially to compare with the speed of the machine in RPM. For example, if the speed of the Pre-heater Fan is 990 RPM, it is advantages to refer to the vibration frequency as 990 cycles per minute instead of 990/60 cycles per second. The vibration frequency resulting from various machinery defects will have a definite relation with the rotational speed. For example, if a Fan is having unbalance, it will produce one cycle of vibration during one rotation. If the machine is having looseness, it will produce the vibration frequency equal to twice the rotational speed. Like that the effects of common defects in a machine can be correlated with the vibration frequency which shall be equal to certain multiples of RPM.

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1.5.5 Phase Vibration Phase is another very important characteristic. Vibration Phase can be defined as the instantaneous position of the vibrating part with reference to a fixed reference or with reference to another vibrating part. The Phase measurement is useful to compare One vibratory motion with another. It also helps in identifying the relative position of one vibratory part with reference to another vibratory part. The measurement of vibration phase with reference to a stationary part is illustrated in Fig1.5.5

1.5.5 : Vibration Phase With Reference To A Stationery Part

In this Fig, the rotating shaft is taken as the reference mark and vertical axis is taken as the reference position. At the particular instant of vibration, it could be seen that the reference mark (arrow mark) is 90° in the anti-clockwise direction with reference to the vertical axis. Hence, we say that the phase of the vibratory part at this instant is 90° . 1.5.6 Significance of Vibration Characteristics

Displacement, velocity Acceleration

indicates Health of the Machine

Frequency indicates Causes of Vibration

Phase indicates Exact location of Defects. (Phase is also used for Dynamic Balancing of Rotating Parts)

1.6 Acceleration Enveloping Acceleration Enveloping is a method for intensifying the repetitive components of a dynamic signal to provide an early warning of deteriorating mechanical condition. Common applications are concerned with rolling element bearings and gear mesh fault analysis.

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The vibration frequency caused by a race or anti-friction roller bearing defect depends upon how often the defect strikes another part of the bearing. The rolling element impacts the defect and the repetitive impulse depends on number of balls, bearing geometry, and defect location. For example, if there is a chip on an outer race, each roller will strike it as it goes by and cause a vibration signal. This signal can often be identified as some multiple of shaft rotational frequency. The multiple is estimated by knowing a bearing’s geometry and number of rollers. A vibration signal from a defective bearing is made up of low frequency signals from rotational components, defect impulse signals, and machine noise. Often bearing fault signals are of very short duration which translate in the frequency domain as small harmonic amplitudes spread over a wide frequency range and buried in machine noise. Machine noise masks the early stages of bearing faults making spectrum analysis alone a difficult diagnostic tool. Envelope analysis first filters out the low frequency rotational components from the complex signal. The high frequency repetitive components are enhanced and converted down to the bearing spectrum range while machine noise is reduced by a significant signal-to-noise factor. If vibration amplitudes appear in the envelope spectrum that is related to bearing defect frequencies it can be deduced that an incipient bearing defect is in progress. Envelope analysis techniques permit an earlier prognosis of an eventual bearing failure by reducing masking noise and by enhancing the significant spectral components relating to bearing performance. 1.7 SEE (Spectral Emitted Energy) SEE Technology is a new type of bearing fault detection method developed by SKF in order to better monitor bearings. This method of monitoring bearings breaks away from the traditional approaches to the problem by using high frequency, acoustic emission detection in the frequency range of 250,000 Hz to 350,000 Hz. This technique has characteristics that set it apart from normal vibration analysis at 0 to 20 kHz and other enveloping techniques at 5 kHz to 60 kHz. SEE provides an excellent way of monitoring problems with bearings that other current technologies cannot provide. Some of these advantages are detection of: early bearing defects, lubrication problems, which stem from contamination, and fretting. An acoustic emissions transducer is sensitive to metal-to-metal contact that occurs when bearing elements roll over a bearing race without an intervening lubricating layer. The transducer emits a high-frequency, pulsed voltage that gives evidence of the defective event. Thus SEE Technology, like enveloping, provides an early warning of deteriorating mechanical condition.

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A higher-than-normal SEE reading, based on a statistical mean and standard deviation, would suggest either insufficient lubrication or the beginnings of a bearing defect. If the readings were to return to normal after the application of additional lubrication, you can assume that the proper corrective action was taken. SEE readings can appear to return to the normal range after a defect spreads. At this point, acoustic emissions may be reduced as the impact stress factors between ball and race become marginal. 1.7.1 Taking SEE Measurements Low frequency vibration is measured by firmly placing a sensor close to the part of the machine being measured. The more securely the sensor is placed, the better the reading is. This is not true with SEE readings. Secure placement doesn’t necessarily mean better readings. A vibration probe actually moves with the machine it is resting on. The machine’s vibration is transferred to the probe which has a crystal inside which shakes with the probe. The crystal converts the mechanical motion into electrical signals. The SEE sensor is, for all practical purposes, a microphone. It does not have to be held firmly against the machine to get a good reading. The SEE sensor is listening for acoustic signals to be transmitted from the surface under it. Since these acoustic signals attenuate very easily, and air is an excellent attenuator for acoustic signals, they certainly will not reach the SEE sensor if the transducer is just held against the machine. The secret is in the coupling between machine and sensor. The coupling most widely used is grease. Grease fills the air gap between the sensor and the surface of the machine resulting in the acoustic signals traveling from the machine, through the grease, to the SEE sensor. Because of the nature of the SEE signal generating mechanism, it is important to note that zero readings or trends do not necessarily indicate that there is no bearing defect. Experience with SEE measurements and case histories developed by successful use of SEE have provided the following guidelines: 0 to 3 SEE No identifiable problem. 3 to 20 SEE Lubrication problem, contamination, bearing defect with light load, or a small bearing defect with normal load. 20 to 100 SEE Bearing defect or contamination. 100 SEE Severe bearing problem. and above

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1.8 HOW MUCH VIBRATION IS TOO MUCH ? As explained above, the amplitudes of vibration the displacement, velocity and acceleration is a measure of the severity of defects in a machine. Therefore, it is important to know how much of vibration is too much for a machine. The answer to these questions will help in identifying the health of the equipment that needs corrective actions. Our main objective is to use the vibration checks, to detect and pinpoint the defects and problems in a machine. Such identification should be done, preferably in early stages; so that required corrective actions can be implemented, before the machine fails. It should be borne in mind that the goal of vibration measurement is not to find out how much vibration a machine can withstand before it breaks down. The main objective is to get an advance warning of the impending problems and defects in a machine so that suitable corrective actions can be implemented at the right stage. The permissible vibration limits (vibration tolerances) cannot be specified in absolute terms, because it is impossible to select a vibration limit, which, if exceeded will result in immediate Break Down of the machine. The development of failures in a machine is a complex mechanism because of many possible ways of failures of machinery part. Since machinery vibrations are caused due to defects and inaccuracies in the machine, the vibration can be used as an effective indicator of machinery health condition. The guide lines given for the permissible limits of vibration as per ISO 2372 have been formulated after several decades of experience gained from machinery vibrations and their harmful effects on machine health condition. A careful implementation of these limits on machines shall be very effective Ø In Pinpointing machinery defects Ø To prepare action plan for implementation Ø To achieve maximum operating life of plant machines Ø To establish a Condition Based Dynamic Predictive Maintenance Ø To achieve the ultimate goal of obtaining maximum productivity from the

installed plant and machines, at minimum cost. 1.9 CONCLUSION The important vibration characteristics are explained in this chapter. The use of vibration characteristics to identify,

Ø Health Condition of Machine

Ø Defects in a Machine

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Ø Location of Defects

The use of vibration amplitudes (Displacement, Velocity and Acceleration) of machine in a plant to detect whether the machine is running smooth or the machines are having problems, is the first stage of assessing operating condition. These simple parameters are to be correctly understood and applied by operators and maintenance technicians.

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Notes:

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2.0 VIBRATION SIGNATURE ANALYSIS 2.1 INTRODUCTION All operating machines give rise to vibration. A deterioration in the machine’s running condition almost always produces a corresponding increase of the vibration level. By monitoring vibration level it is therefore possible to obtain information about a machine’s condition. As the main sources of faults in industrial machinery have a mechanical origin, it is logical to choose a mechanical phenomenon as the representative parameter of machine condition. Mechanical vibration has proved to be one of the most reliable parameters to use in machine health monitoring to check machine condition. A doubling of dynamic force will typically result in a doubling of vibration measured at the forcing frequency. 2.2 MACHINERY SIGNATURES When broken down into spectral components, the complex waveforms, referred to as machinery vibration, or acoustics may in general be defined as a sum of harmonic functions of discrete amplitudes and frequencies. This is often referred to as the `Machinery Signature’ typical of which is depicted in Fig. 2.2 For example, in the case of an unbalanced shaft rotating at an angular velocity `w’, in a circular orbit with radius `A’ the vibration displacement signal may be represented by a single harmonic function.

FIG NO.2.2

Fig No:1 Fauilts in differ ent machine e lement will genera te

peaks in specif ic frequency ranges. n b X (t) = A sin w t (1) where `t’ denotes time.

FAULTS IN DIFFERENT MACHINE ELEMENT WILL GENERATE PEAKS IN SPECIFIC FREQUENCY RANGES

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By differentiation, we have that the velocity and acceleration are given by : X (t) = A w cos w t (2) and X (t) = - A w2 sin w t (3) The preceding expressions (1), (2), & (3) illustrates how velocity and acceleration components of a complex signal (and thus, the signal itself) are dependent on the angular frequency `w’. The velocity increases in direct proportion to frequency while acceleration increases in proportion to frequency squared. For example, if displacement is held constant while the frequency is doubled, velocity doubles and acceleration increases by a factor of 4. It can be shown with a simplified model of a rotor element of mass `m’ operating in bearings with spring coefficient `k’ and viscous damping factor `c’ that the total dynamic force during vibration may be represented by an expression of the form: .. . F(t) = m x + c x + k x (4) We here recognize the terms as those of inertia, damping and elastic forces respectively. By substitution of equation (1), (2) and (3) into (4), we find that the relative size of the elements of (4) are dependent on factor `w’. An important observation to be made at this point is that at high frequencies, the force component due to acceleration (or velocity) may dominate. Thus, it is entirely possible to have large forces acting even with unmeasurably low displacement amplitudes. It is indeed these high frequency components, which remain undetected by displacement criteria, that often cause sudden catastrophic fatigue failures. One may conclude that displacement alone, as often applied, can be a poor measure of vibration severity since this should always be considered in conjunction with frequency. To overcome this particular difficulty, maximum velocity criteria have often been suggested in order to include the frequency parameter and total dynamic forces concept into the measurement of machinery condition. The adequacy of even such criteria is, however, suspect since frequencies of interest in process machinery analysis form a wide, continuous spectrum from the sub-audio range of 10 to 15 Hz through the ultrasonic range approaching 100 kHz. Within this spectrum are the subharmonic `whirl’ frequencies from approximately 40 to 90 Hz, once per revolution frequencies from 60 to 700 Hz, centrifugal compressor vane passing frequencies from 2 to 5 kHz, turbine and gear passing frequencies from 5 to 30 kHz, and finally, antifriction bearing characteristics from 20 to 100 kHz. With this vast range and amount of data to be considered, the problem facing the industrial user is one of how to acquire, reduce and present the information in a form conducive to effective and meaningful evaluation.

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2.2.1 MACHINERY VIBRATION SIGNATURE A Plot of the vibration amplitude verses various frequencies is known as a Vibration signature. It is also known as vibration profile. The Vibration signature or Profile breaks the total vibration signal into its frequency components and will exhibit the existing vibration amplitudes at each of these frequencies. In a vibration signature the X axis always indicates the vibration frequency either in the Linear Scale or in the Logarithmic scale. The Y axis indicated the vibration amplitudes like displacement, velocity or acceleration. Depending on the type of instrumentation and or the analysis requirements the vibration signatures can be arranged in different fashions in a single format or in a multiple formats. One very useful method of arrangement is to accommodate the horizontal, vertical and axial vibration signatures of a particular bearing in a single data format ; so that data interpretation becomes more effective. The arrangement of vibration signatures in a continuous fashion (in a 3D format is known as Water-fall for curve. This depicts the behavior of the vibration signature over a period of time which will be helpful in understanding the variations / changes taking place at different frequencies over the period under consideration. Since the method of enter r g this data computation and comparisons in different formats mainly depends on the type of instruments and softwares being employed for such purposes. Basically; the analysis in the frequency domain is the technique by which the vibration signals are resolved displayed as narrow band spectra components. Hence the frequency enter r converts the complex vibration signal, from the time domain to the frequency domain, spectrum displayed of various frequency verses amplitudes. Analysis in the frequency domain is a technique where the vibration signals are resolved and displayed as narrow band spectral components. Here the frequency enter converts the complex vibration signal from the time to the frequency domain producing a display of frequency versus amplitude. In its simplest form, a frequency enter may be a manually tuned filter slaved to a plotter with frequency and amplitude positioning the X and Y axes respectively. This method is far too slow and inaccurate for machinery vibration analysis purposes, however, and while a wave enter with automatic programmed frequency scan is more accurate, the analysis speed is till too low. The real time frequency spectrum enter (RTA) combines accuracy with rapid analysis to produce an output which may be displayed on an oscilloscope. This instrument receives the complex analog vibration signal directly or from magnetic tape and converts it to digital form. A time compression technique translates the digital representation to a high frequency where the signal is analysed with a broad band filter.

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The high frequency broad band filter permits a rapid scan rate while producing a very good frequency resolution. Thus, dynamic changes in the frequency domain such as beats and harmonics can be conveniently observed as they occur. The spectrum can be preserved either by directing the output to a plotter or by photographing the oscilloscope. In order to gain statistical accuracy and insure that the graphic spectral representation reflects an average rather than an instantaneous condition, a spectrum average is usually utilized with, the enter in this mode. 2.3 ROTATING MACHINERY SIGNALS Fig 2.3 shows a schematic drawing of an electric-motor-driven gearbox, driving a ball Mill

1 2 3 4

5 687

FIG 2.3 LOW FREQUENCY RANGE (Refer Fig 2.3) Frequency spectra obtained from measurements made on the motor bearings or gearbox bearings will reveal low frequency components at shaft revolution speed originating from unbalance, misalignment, bent shaft, etc. At the second harmonic, i.e. twice the rotation speed, components originating from bent shaft and misalignments will also be found. In other words, by comparing the change in these components with time the appearance of such faults can be detected and diagnosed. The major problem with journal bearings is hydrodynamic instability in the system consisting of shaft/oil film/bearing housing. Oil whirl is a self-excited vibration, when the center of gravity of the shaft moves around within the clearance of the bearing at sub-synchronous speed. It arises typically at lightly loaded high speed shafts, and at frequencies between 40% and 49% of rotational speed, although oil whirl has been seen at higher frequencies. It seems to depend on the surface of the shaft and bearing. Hysteresis whirl is another self-excited instability. It arises when the rotor passes through its critical speed, and then it will maintain that frequency, independent of the rotor speed. It originates from mechanical hysteresis in the rotating system. Another instability of

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flexible rotors is flow-induced, and it appears at the same frequencies as hysteresis whirl, but has another mechanical background. A last type of fault appearing in this low frequency area is mechanical looseness. In many cases mechanical looseness will create inter harmonic and sub-harmonic components, i.e., a `half’ harmonic, a `one-and-a half’ harmonic, a `two-and-a half’ harmonic etc. MEDIUM FREQUENCY RANGE Higher up in the frequency range, components originating from the tooth mesh in the gear box will be found and are in this context referred to as medium frequency components. They will be at a frequency corresponding to rotational speed multiplied by the number of teeth on the gear, and referred to as the tooth meshing frequency. A new and healthy gearbox will clearly exhibit this frequency, but not just by itself. Due to mechanical loading the tooth will deflect, but they will deflect differently depending on how many teeth there are in mesh. When the gearbox wears, the gear profile will gradually change due to sliding between two teeth in mesh at any point except at the pitch point. This indicates that changes due to wear in a gearbox will turn up at the second harmonic of tooth meshing frequency and, since the change is not sinusoidal, higher harmonics will be revealed as well. An initial local fault, on the other hand, will not increase the level at tooth mesh frequencies and its harmonics. Imagine a cracked tooth which is not yet broken, and will consequently not be noticed by the operation personnel. However, it will, due to its weakened mechanical condition, deflect more under load than the other (healthy) teeth when it goes into mesh. This can be viewed as an ordinary signal from a healthy (but may be worn) gearbox, superimposed by a series of pulses, originating from the excessive deflection. A series of pulses will create a line spectrum, with each line spread with the repetition frequency. The envelope of this spectrum will be identical to the spectrum of each individual pulse (although at different scales). Thus an incipient fault will turn up in the frequency spectrum as an increased level in the sidebands spaced with rotation speed below, as well as above the tooth meshing frequency. But below the tooth meshing frequency we find all the low frequency signals already mentioned, originating from unbalance, bent shaft, misalignment, etc. and as these signals have much more energy, the weak signals from the cracked tooth will not be seen here. The tooth meshing frequency and its harmonics also have much more energy than the signals from an incipient fault in the gearbox. It will however, often be possible to see these small increases between the harmonics of the tooth meshing frequency.

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As the fault spreads out in the gearbox to cover several defective teeth, the signal will change character from the pulse-like one just described to one which has more energy and which looks more like an amplitude modulated signal. Its spectrum shows similarities with this, and has high amplitude sidebands around tooth meshing frequencies and its harmonics, spaced at rotational speed. The effect of general wear is indicated by the slight increase in the level at the tooth meshing frequency, but much increases in the second and third harmonics of tooth meshing frequency. Significant increases in the level of components spaced at rotational speed are also revealed between the tooth meshing frequency and second harmonic. This indicates the presence of a local incipient fault in the gearbox. The Ghost component is due to geometrical inaccuracy of the gearwheel originating from the index-wheel during manufacture. It disappears with general wear. HIGH FREQUENCY RANGE Another source of signals to consider are those signals originating from incipient faults in rolling element bearings. An incipient fault in a rolling element bearing will typically be a crack or corrosion pit either on the inner race, on the outer race, or on the rolling element itself. This crack will create small impulses every time one of the rolling elements pass over it. These impulses will transmit energy to the bearing housing, which in turn will vibrate at its natural frequency and decay with the damping in the mechanical structure. 2.4 FREQUENCY ANALYSIS Vibration motions can range from a simple Harmonic Motion to very complex Non- Harmonic Transient signals. Hence both their wave form and their corresponding frequency components are important in Condition Monitoring. This conversion from a time domain to a frequency domain is mathematically carried out by Fourier transform. The same is carried out through the use of frequency analyzers. Fig.2.4.1 to Fig 2.4.4 shows some of these signals both in their time domain and frequency domain form. All of them are discrete frequency components. When random vibrations or shock or impulsive motions are involved, their frequency spectrum is rather continuous as can be seen in Figs2.4.5 to 2.4.8. As Already mentioned, random vibrations are caused by random forces, and at every instant the signals recorded look different but most of the practical random vibration problems are of stationary type, and they can be described by certain parameters through statistical analysis of a number of records. Important parameters involved in this type of motion are:

Ø Mean or average values Ø Mean square values (also known as Variance) Ø Root mean square values (also known as standard Deviation) Ø Power spectral density Ø Probability density and distribution functions

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Ø Auto and cross correlations Ø Coherence and transfer functions

Going in detail into all these parameters is beyond the scope of this lesson, and it would be enough, if we can recognize the random frequency spectrum, classified under:

Ø Wide Band Spectrum (Fig.2.4.5) Ø White Noise Spectrum (Fig. 2.4.6) Ø Narrow Band Spectrum (Sine random) (Fig. 2.4.7)

The non-periodic motions like shocks and impulses, which act only for a short duration have their frequency spectra almost similar to characteristics of random vibrations. This is shown in Fig. 8. Shocks are produced not only due to impulsive forces generated during the operation of machinery, but also due to pits, crevices and other irregularities on the surface of balls or races in anti-frication bearings. Usually such shock motions result in very high frequencies, which can be picked up by modern instruments like shock pulse Meter (SPM), Kurtosis Meter, Acoustic Emission (AE) very effectively and the defect identified long before major breakdowns occur. Frequency analysis is the next important step for signal identification, and for this purpose we have different types of equipment, which can operate with desired accuracy. These are: 1. Octave band filters

Ø Fractional octave filters Ø Constant percentage band width filters

2. Narrow band width analysers 3. Time compression analysers

Ø High resolution real time anlysers Ø Cepstrum analysers Ø Level recorders

The simplest analyzer is an octave filter, the word octave coming from music. In this the upper to lower frequency ratio is always two. It reads as Follows (all in Hertz or Cycles/Sec) : 31.5, 63, 125, 250, 500, 1000 (1K), 2K, 4K, 8K, 16K, etc.

f upper ------- = 2 f lower

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For representing the energy contained within that band, a central frequency f c is defined as :

If we observe the higher bands in this octave analysis the range is quite high and any important frequencies relating to a particular malfunction are not dominantly shown as it will be an average doubt over the whole band. To illustrate, if we are doing the frequency analysis between 2000 Hz – 4000 Hz and if we are looking for an important frequency say, 560 Hz relating to a particular defect, this cannot be identified. Hence, in practice, one has to use FRACTIONAL FREQUENCY ANALYSERS, Such as :

(a) Half Octave (The ratio of f u / f l = 2 ½) (b) Third Octave (The ratio is 2 1/3 ) etc. By using these fractional analysers, the band width is reduced. For example: In octave (1/1) analysis : 2000Hz , 4000 Hz In half octave (1/2) analysis : 2000Hz , 2800Hz , 4000 Hz In Third octave (1/3) analysis : 2000Hz,2500Hz,3150Hz ,4000Hz We will have a better resolution by going in for Narrow Band analysis. Fig. 2.4.9(a & b) shows the frequency spectra taken in the three types of analysers discussed above. The above method of frequency analysis is of sequential type, in the sense that the analysis is carried out one band after the other assuming that the signal is steady and not changing with time, particularly if it is near about resonance, for a parallel type filtering, in which the signal passes through the various bands simultaneously to give a frequency spectrum in REAL TIME. Such type of Fast Fourier Transform (FFT). Cepstrum Analysis is needed while identifying very complex signals like gearboxes, steam turbine casings etc. This is in principle, a kind of frequency analysis. To distinguish this form the conventional “Spectrum”, “Frequency”, used in normal frequency analysis, here the words “Cepstrum”, “Quefrency” etc., are used. A detailed discussion of this is beyond the scope of this lesson. The discussion is on frequency and how the machinery vibration can be related to frequency. The “Indication-and-cause” table given below will provide initial frame of mind for use as a guide, for FREQUENCY analysis.

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FIG – 2.4.1

FIG – 2.4.2

FIG – 2.4.3

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FIG – 2.4.4

FIG 2.4.5 & 2.4.6

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FIG – 2.4.7

FIG – 2.4.8

FIG – 2.4.9

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The basic purpose of a vibration frequency analysis is to diagnose machinery problems. The following chart indicates the probable causes of vibrations based on the vibration frequency :

S.No INDICATION:

Frequency in speed (RPM)

Probable main cause

Possible other causes

1.

1 x RPM

Unbalance (a) Misalignment or bent shaft

(b) Bad belts of (belt RPM) (c) Eccentric Journals, gears or

pulleys (d) Electrical faults

(e) Resonance (f) Reciprocating forces

2.

2 x RPM

Component Mechanical looseness

(g) Bad Belts (If 2 belts are there)

(h) Misalignment (In case of high axial amplitude) (i) Resonance

(j) Reciprocating influences 3.

3 x RPM

Misalignment

(a) Higher axial clearance

4.

5 x RPM to15 x (RPM High Frequency ranges)

Damaged Rolling Element Brgs

(a) Flow turbulence in

sleeve Brgs-random high frequency vibration

(b) Friction in sleeve bearings

(c) Rubbing

5.

Many Times RPM

(a) Bad gears

(Gear teeth) (b) Fan (Blades

times RPM)

(a) Mechanical looseness

(2,3,4 and higher harmonics)

(b) Pump impeller no. of vanes x RPM (c) Reciprocating Forces

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S.No INDICATION:

Frequency in speed (RPM)

Probable main cause

Possible other causes

6. Less than 1 x RPM

Gil whirl 42% to 48% RPM

(a) Cross Talk or Back

Ground vibration (b) Beat Vibration

7.

Electrical supply Frequency

Electrical Problem (a) Broken Rotor Bars (b) Unequal air gap

(c) Unbalance Phases

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Notes:

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3.0 PHASE ANALYSIS

3.1 INTRODUCTION The Vibration Phase analysis is a very useful diagnostic method to distinguish one defect from the other. Once the Phase Analysis is properly understood the same can be used very effectively even to pinpoint various problems. DIFFERENT USES OF PHASE ANALYSIS

Ø By using stroboscope in conjunction with Vibration Analyser. Ø Phase measurement with a Phase Meter Ø Phase measurements by using reference pickups. Ø Measurement of phase by using Oscilloscope

Basically Phase measurements are used to distinguish:

Ø Misalignment from a bent shaft Ø Detecting reacting forces versus unbalance Ø Resonance conditions Ø Pinpointing mechanical looseness Ø Mode shape analysis etc.

Another very useful analysis technique which can aid in the detection and identification of machinery problems is “phase” measurement and analysis. Phase measurements are normally expressed in degrees, where one complete cycle of the vibration equals 360 0. Phase can be defined as ” that part of a cycle ( 0 0 – 360 0) through which one part has traveled relative to another part or a fixed reference”. Phase measurements simply provide a means of determining the relative motion of various parts of the machine.

Weights vibrating in phase

Refer figure given below Weights “C” and “D” are vibrating in-phase or 90 0 out-of-phase.

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Weights Vibrating 90 0 out-of phase

Weights vibrating 1800 out-of- phase 3.2 PHASE MEASUREMENT TECHNIQUES There are many techniques which can be used to obtain comparative phase measurements for analysis and balancing including the stroboscopic light, phase meter and oscilloscope. 3.2.1 STROBOSCOPE LIGHT PHASE MEASUREMENT Most portable Vibration Analyzers are furnished with high intensity stroboscopic (strobe) light which when triggered by he measured vibration, provides a quick and convenient means of obtaining phase readings for analysis and balancing. To obtain phase readings, a reference mark is placed on the rotor at the end of a shaft or some other location which can be readily observed. For accurate phase measurements an angular reference scaled in degrees can be super imposed around the shaft, making it possible to obtain phase readings with minimum error. With the analyzer strobe light being triggered by the machine vibration, the reference mark on the rotor will appear stationary

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at some angular position. The phase measurement is then obtained by noting and recording the indicated angle from the angular reference. Comparative phase readings are obtained by simply moving the vibration pickup to other measurement locations but using the same reference mark and angular reference used for the initial measurement. Although the strobe light is very easy and convenient to use for obtaining phase readings where the vibration frequency of interest is equal to 1 x RPM of the machine. If for example the vibration of interest is occurring at 2 x RPM then the strobe light will flash twice for each revolution of the shaft and a reference mark will appear in two positions when observed with the light. Obviously, it would be impossible to obtain comparative phase readings with two or more identical reference marks visible under the light. Where it is desired to obtain comparative phase measurements for multiple, sub-multiple or non harmonically related vibration frequencies, some other technique for phase measurement must be used such as a phase meter or oscilloscope. STROBOSCOPE

3.2.2 PHASE MEASUREMENT WITH A PHASE METER Where it is impractical or unsafe to obtain phase readings with a strobe light, an instrument is the IRD Model 360 Vibration Analyzer. The overlapping 0 0 - 18 0 0 and 18 0 0 - 36 0 0 scales used scales used on the Model 360 phase meter. The monitor phase module provides a direct readout of the measured phase on a digital meter with accuracy within +_1 0 . To obtain a remote readout of phase on an instrument like the Model 360 requires a voltage reference pulse at the desired vibration frequency. Where it is desired to observe the phase of the vibration occurring at 1 x RPM, a reference pickup such as a photocell, electro magnetic pickup or non-contact. Pickup is mounted close to the shaft which has been properly prepared to “trigger” the reference pickup. Since the photocell responds to changes in reflectivity of the target, a common method of preparing the shaft is to wrap black non-reflective tape around the shaft and then paint a white line across the tape, or attach a small piece of reflective material such as metal foil to the black tape. Some prefer to paint the target area with

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flat black paint in lieu of using non-reflective tape. If the shaft is clean and shiny, the proper trigger for the photocell can be provided by simply attaching a small strip of non-reflective tape to the shaft or by painting a small strip with flat black paint. In any case, the objective is to produce an abrupt change in the reflectivity of the target area of the photocell for each revolution of the shaft. In the event that it is not possible to mount a reference pickup on the machine, it is still possible to obtain phase measurements with a remote phase instrument such as the Model 360. This can be accomplished by using the analyzer or oscilloscope output of another vibration measuring instrument as the reference source. Simply mount a “reference” vibration pickup at any convenient location on the machine or structure where a strong signal can be obtained. This “reference” vibration pickup is then connected to the vibration meter. The analyzer output of the vibration meter is then connected to the reference input of the phase measuring instrument in place of the reference pulse from a photocell or electromagnetic pickup. Adjust the amplitude range selector of the vibration meter until a suitable reference signal is obtained, as indicated by a steady phase reading. It is necessary to overdrive the vibration meter by one or two amplitude meter “chops” the output reference signal resulting in a simulated square-wave similar to the normal reference pulse. In any event, overdriving the vibration meter will not damage the instrument and does give the desired results. The “reference” vibration pickup must not be moved if comparative phase readings are to be made at several measurement locations. This technique for measuring phase is only useable for phase analysis and cannot be used for balancing. Applying a trial weight to the rotor would change the phase of both the phase reference signal and vibration signal. If the vibration is complex or if the vibration frequency of interest is not the predominate vibration, it may be necessary to substitute a vibration analyzer with tunable filter in place of the vibration meter. In this way, comparative phase measurements can be obtained for virtually any frequency of interest synchronous, multiples, sub-multiples and even non-harmonic frequencies. This offers a strong advantage over the strobe light for phase measurement since the strobe technique is limited to vibrations at 1 x RPM. The 1 x RPM voltage pulse from the reference pickup applied to the IRD Model 360 actually serves tow purposes. First, the reference pulse serves to automatically tune the filter of the analyzer to the rotating speed frequency. In this way, should machine RPM change, the filter will automatically adjust to the change to provide consistently accurate amplitude and phase data. Secondly, the reference pulse is electronically compared to the filtered vibration signal to provide a measure of relative phase. In simple terms, the instrument measures that portion of a cycle by which the reference pulse leads the neutral-to-positive ( 0 to + ) crossover of the vibration signal. This is then converted to a proportional D.C. voltage needed to drive the phase meter. (Refer figure 71 page 56 vec/sbv/7c)

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PHOTOCELL KIT

Figure illustrates various relationships between the reference pulse and vibration signal and the resulting phase indications. Examples of phase measurements using a reference pulse superimposed on the vibration waveform

= 00

= 900

= 1800

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Unlike the stroboscopic light, the phase meter can be used to obtain comparative phase data for vibration frequencies other than 1 x RPM. All that is needed is a reference signal at the desired frequency. For example if it were desired to analyze the phase of vibration for a frequency of 2 x RPM it would only be necessary to put two triggers on the rotating shaft 180 0 apart instead of the usual one trigger. In this way. Two reference pulses will be generated for each shaft revolution, the analyzer filter will automatically be tuned to 2 x RPM, and the phase readings noted on the meter will be those for the 2 x RPM vibration. Where it is necessary to obtain comparative phase data for sub-multiple or non-harmonically related vibration frequencies, a “reference” vibration pickup and a “reference” vibration analyzer with tunable filter can be used to provide a reference signal at any desired frequency of machine vibration. 3.2.3 OSCILLOSCOPE PHASE MEASUREMENT The oscilloscope is another useful device for obtaining comparative phase measurements and can be used where readings with a strobe light are not possible or an instrument such as the IRD Model 360 with remote phase readout is not available. Any standard single-trace or dual - trace oscilloscope can be used with standard IRD vibration analysis equipment. 3.2.4 MEASURING PHASE WITH A SINGLE – TRACE OSCILLOSCOPE To obtain comparative phase measurements with a single – trace oscilloscope requires a voltage reference pulse similar to that required for the IRD Model 360. The reference pulse can be applied to the oscilloscope in a couple of different ways to measure phase. One way is to switch the oscilloscope to the EXTERNAL SYNC mode of operation and connect the reference signal to the scope EXTERNAL SYNC (horizontal) input. The vibration signal at reference pulse frequency is connected to the oscilloscope VERTICAL input from the analyzer oscilloscope output. The resulting display on the CRT will be one complete cycle of the vibration, and the horizontal position and gain controls of the scope should be carefully adjusted to display the vibration waveform over the full width of the scope graticule. The vertical position of the trace should have been adjusted initially to the center line. The configuration of the vibration waveform on the scope display will depend on the phase relationship between the reference signal and the vibration signal. By scaling the horizontal axis of the scope from 0 0 to 360 0 0 . The neutral-to-positive crossover point is a common choice and is used here, although some other point such as the neutral –to-minus ( 0 to - ) crossover or peak amplitude could be used just as easily, as long as consistency is maintained. Another way to obtain phase measurements with a single-trace oscilloscope is to superimpose the reference pulse on the vibration waveform. With the reference pulse superimposed on the vibration waveform, phase measurements can now be obtained again by noting that portion of the vibration cycle which separates the reference pulse from a common point of the reference on the waveform such as the neutral-to-positive crossover. The addition, the convention of phase determination must

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also be established . Specifically, it must be decided whether the phase angle recorded is the angle by which the reference pulse leads or lags the neutral-to-positive crossover. For comparative phase readings, the same convention must be used for all measurements. A single-trace oscilloscope can be used to obtain comparative phase measurements for practically any vibration frequency of interest as long as a suitable reference pulse is available at the selected frequency. In addition, it is also possible to obtain phase measurements for vibrations occurring at frequencies which are multiples of rotating speed where only a 1 x RPM reference pulse is available. The phase of the 2 x RPM vibration can still be obtained by simply noting the angle by which the reference pulse leads the neutral-to-positive crossover of the vibration waveform. 3.2.5 MEASURING PHASE WITH A DUAL TRACE OSCILLOSCOPE In cases where it is not possible to obtain a reference pulse for the vibration frequency of interest for phase measurement with a phase meter or single-trace oscilloscope, it is possible to obtain comparative phase data for tow or more measurement locations by using a dual-trace oscilloscope. The relative phase between the two signals is determined by simply measuring that part of the vibration cycle which separates common oinks of reference such as neutral-to-positive cross-over points. Where comparative phase readings will be taken at more than two measurement locations, one of the first two signals should be retained for comparison with all other signals. In addition, the “lead” or “lag” convention for phase determination must also be selected to insure that al readings are directly comparable. 3.2.6 SOME PRACTICAL CONSIDERATION FOR PHASE MEASUREMENT There are several factors which can affect the accuracy of phase readings taken for analysis or balancing: They are 1. The direction of the vibration pickup axis, together with the selected phase reference system, established the “fixed reference” need to take comparative readings. The direction of the pickup axis should not be changed from one measurement point to the next. If pickup direction is changed this change must be noted so that the phase readings can be corrected accordingly for comparison. This is due for all phase measurement technique – strobe light, remote phase meter or oscilloscope. 2. When using a standard analyzer with manually tunable filter, it is essential that, when used the filter be precisely tuned for each reading taken. For example, if the machine speed change slightly between readings, then the analyzer filter is to be “detuned”,

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3. When taking comparative phase readings, it is important that all measurements

be taken using the same parameter of amplitude readout. Switching from displacement to velocity readout will result in a 90 0 change in the measured phase. Switching from displacement to acceleration will reveal a 180 0 phase difference.

3.3 PHASE ANALYSIS TECHNIQUES 3.3.1 PHASE MONITORING Many companies are currently including continuous phase monitoring systems on critical high-speed machines such as turbo-generators and centrifugal compressors for added protection against mechanical failure. The phase of the vibration being monitored at any point on the machine can be selected for readout on the digital phase meter. The concept of monitoring phase in addition to vibration amplitude is based on the fact that some machine problems may occur which could result initially in little or no significant change or increase in vibration amplitude. The vibration amplitude and phase readings were taken on a machine and indicated an amplitude of 2.0 mils and a phase reading of 90 0.. Then, a new reading was taken at the same location today indicating an amplitude of 2 mils but a phase reading of 270 0. The comparative amplitude readings show no increase and would suggest that there has been no change in machinery condition. However the phase readings show that there has been a substantial shift in phase which represents a 4.0 mil effective change in vibration amplitude. The phase change detected a significant change in machine condition possibly due to a thrown rotor blade, cracked rotor shaft or some other potentially serious problem. 3.4 APPLICATION OF PHASE MEASUREMENTS IN ROTOR DYNAMICS : Rotor and different modes of machinery vibrations are analysed and identified by introducing the influence of all static and dynamic forces. Whenever the vibratory mode changes; it is associated with a major phase change. Normally there will be a phase shift of 1800, when the vibration changes from one mode to the other.

Phase changes during different conditions :

Ø Force (or Static) unbalance is evidenced by nearly identical phase in the radial direction on each bearing of a machine rotor.

Ø Couple unbalance shows approximately a 1800 out-of-phase relationship when

comparing the outboard and inboard horizontal, or the outboard and inboard vertical direction phase on the same machine.

Ø Dynamic unbalance is indicated when the phase difference is well removed from

either 00 or 1800, but importantly is nearly the same in the horizontal and vertical

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directions. If the horizontal phase difference varies greatly from the vertical phase difference, this strongly suggests the dominant problem is not unbalance.

Ø Angular Misalignment is indicated by approximately a 1800 phase difference

across the coupling, with measurements in the axial direction. Ø Parallel misalignment causes radial direction phase across the coupling to be

approximately 1800 out of phase with respect to one another. Ø Bent shaft causes axial phase on the same shaft of the machine to approach a 1800

difference when comparing measurements on the outboard and inboard bearings of the same machine rotor.

Ø Resonance is shown by a 900 phase change at the point when the forcing

frequency coincides with a natural frequency, and approaches a full 1800 phase change when the machine passes through the natural frequency (depending on the amount of damping present)

Ø Rotor rub causes significant, instantaneous changes in phase. Ø Mechanical looseness/weakness due to base/frame problems or loose hold – down

bolts is indicated by nearly a 1800 phase change when one moves his transducer from the machine foot down to its baseplate and then down to its support base.

Ø Mechanical looseness due to a cracked frame, loose bearing or loose rotor causes

phase to be unsteady with probable widely differing phase measurements from one measurement to the next. The phase measurement may noticeably differ every time you start up the machine, particularly if the rotor itself is loose and rotates on the shaft a few degrees with each startup.

Often, even though phase measurement capability is now offered by most data collectors, users do not use this powerful tool. If not used, this will severely limit the diagnostic capabilities of any program. However, currently it would be impractical to make phase measurements on all machinery during regular PMP surveys. Its greatest use comes into play when performing diagnostics on machines which have developed high vibration at 1X,2X or 3X RPM, requiring investigation to detect the predominate cause(s) prior to taking corrective actions. 3.4.1 APPLICATION OF PHASE MEASUREMENT FOR BALANCING

Phase measurement will help in distinguishing different types of unbalances. The effective time for balancing can be minimized once the exact type of unbalance existing in a rotor, is clearly established, before starting the balancing.

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Identification of unbalance location in rotors is one of the very major functions for phase measurement. Distinguishing one defect from the other; especially the other characteristics will not be helpful others in generalizing the type of inaccuracies. Once the frequency analysis and the amplitude analysis are clear, the phase observations and measurements will provide very powerful to precisely distinguishing and pinpointing machinery problems.

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Notes:

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4.0 MACHINERY PROBLEM DIAGNOSTICS

THROUGH VIBRATION ANALYSIS 4.1 INTRODUCTION

In today’s competitive Industrial environment, it is very essential to achieve maximum availability of plant and machines. The Industrial equipments are subjected to regress operating conditions and therefore the possibilities of break down from various inaccuracies and problems are quite significant. Therefore, a dynamic predictive maintenance based on machinery problem diagnostics is a must in ensuring trouble free operation. Though there are many methods of machinery fault diagnostics (Condition Monitoring), one of the most effective method is Vibration Analysis. The various defects that can occur in a machines can be categorised as follows:-

4.2 VARIOUS DEFECTS IN MACHINES

The most common defects which are associated with plants are listed below:

• Unbalance • Misalignment • Bent Shaft • Mechanical Looseness • Eccentricity • Rubbing • Distortion • Coupling Defects • Loose Foundation Bolts • Cracks in Foundation • Cracks in Structure • Piping Forces • Aero Dynamic Forces • Resonance Condition

Defects in Bearing

• Clearance induced vibration • Unequal Bearing Stiffness • Damage to Thrust Bearings • Defective Journal Bearings • Oil Whirl • Oil Whip

Defects in Gears and Gear Boxes

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• Tooth Meshing Faults • Worn out Tooth • Misaligned Gears • Cracked Teeth • Eccentric Gears

Defects in Belt Driven System

• Mismatched Belts • Misalignment of Pulleys • Damaged Pulleys • Belt Wear • Cracks & Lumps • Eccentric Pulleys

Electrical Defects

• Non Uniform Air Gap • Magnetic Centre Shift • Winding Defects • Unbalanced supply in 3 phase

It could be seen that all these defects in different machines can be grouped into 5 categories.

Ø Unbalance of One or more parts Ø Misalignment of system Ø Out of tolerances – Eccentricity and other dimensional deviations Ø Reaction to some external forces Ø Unbalanced Magnetic forces in the Drive Motors/Generators.

MACHINERY PROBLEM DIAGNOSTIC CHART

Cause Vibration Level Frequency Remarks

Unbalance Proportional to

unbalance. Largest in

radial direction

1 x rpm Most common cause of

vibrations.

Bent shaft Large in Axial

Direction

1 x rpm -

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Misalignment

(Coupling or

bearings)

Large in Axial

direction(50% or more

or radial vibrations)

1X rpm 2X rpm

most common. 3n

sometimes.

Base found by appearance of

large axial vibrations. Positive

diagnosis can be obtained using

dial indicators. If sleeve bearing

machine and no coupling

mialignment problem is probably

unbalanced of the rotor.

Eccentric Journals Usually not large 1 x rpm If on gears, largest vibrations in

line with gear centres.

Rubs No particular

characteristic if

continuous

Mainly 1 x rpm

plus 2 x rpm

Can excite high natural

frequencies of machine.

Amplitude at same speed may

vary between different runs.

Mechanical

Looseness

Variable 2n Usually accompanied by

misalignment and/or unbalance.

Not a loose bearing assembly.

Loose Bearing

Assembly

Variable 0.5 rpm Sometimes manifest at rotor

critical speed. Quite a common

problem.

Oil Whirl Severe, radial motion 0.43 – 0.47 rpm Only in high speed or vertical

rotor machines with pressure

lubricated

sleeve bearings.

Friction Induced

Whirl

Can be severe radial

vibrations

Usually less than

0.4 x rpm and

equal to first critical

speed of rotor

Rare. Can be caused by loose

rotor components.

Rolling Element

Bearing Distortion

Depends upon amount

of distortion

1 x rpm Large component in either

horizontal or vertical planes.

Taper roller bearings will also

have axial components of

vibration.

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Rolling Element

Bearing Damage

Unsteady High frequencies See later discussion.

Bad Gears or Gear

Noise

Low Very high T x rpm

(T number of teeth)

See later discussion.

Faults in Belt

Drives

Erratic or Pulsing 1,2,3 and 4 times

rotational frequency

of belt

Stroboscope can be used to

diagnose belt defects.

Electrical Low, Disappears when

power is turned off

1 x rpm or 1 or 2

times synchronous

frequency

See later notes on electrical

machines.

Aerodynamic or

Hydraulic

Variable 1 x rpm or b,n

(where b is number

of blades or lobes)

Rare as a cause of trouble,

except in cases of resonances.

See later notes on bladed

machinery.

Reciprocating

Forces

- 1 x rpm , 2 x rpm

and higher orders

Inherent in reciprocating

machinery.

Faulty

Combustion in

Diesels

High 0.5 x rpm Faults with injectors, fuel

pumps, calibration or timing

show unbalance. Resiliently

mounted unit rocks at 0.5n

which is close to natural

frequency of mounted unit, and

causes large amplitudes.

Foundation Faults Random and can be

high

Low and erratic Check foundation bolts.

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ANALYSIS OF NATURE & CAUSES OF VIBRATION

IN

INDUCTION MOTORS

FREQUENCY

OF

VIBRATION

TYPES OF

VIBRATION

CAUSE OF

VIBRATION

NATURE OF

VIBRATION

METHODS OF

ESTABLISHIN

G CAUSE

REMARKS

2X Line

frequency

Magnetic -Poor rotor

alignment

Beats would

show up

particular if

there is

associated

mechanical

problem

Vibrations

reduce

considerably

when power

supply is

disconnected

Some possibility of

vibration

frequencies of 2 X

RPM also there.

1 X RPM PLUS

2X slip X pole

pairs and

Mechanical -Shorted stator

turns

2 X slip X pole

pairs

Mechanical -Worn out

bearings

1 X RPM and 1 X

RPM +

Mechanical

-Local stator or

other

imperfections

resulting in

changing axis

of rotation

Strong beat

phenome-non

which can be

easily felt or

seen on the

vibration

meter.

Vibrations

reduce

considerably

when power

supply is

disconnected

from stator.

Some possibility of

vibration

frequency - 2

X RPM and 2 X

supply frequency

also there.

2 x slip x pole

pairs and 2 x slip

x pole

Mecha

nical

-Broken rotor

bars

Mechanical -Rotor run-out

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FREQUENCY

OF

VIBRATION

TYPES OF

VIBRATION

CAUSE OF

VIBRATION

NATURE OF

VIBRATION

METHODS

OF

ESTABLISHI

NG CAUSE

REMARKS

Mechanical Bad electrical

connection

between rotor

bars & end

rings

Mechanical -Local rotor

heating caused

by shorted

laminations,

broken rotor

bars on loose

rotor

Several times

RPM or supply

Frequency

Mecha

nical

-Bad ball

bearings

Continuous type Frequency

closely linked

with: Number

of balls X

RPM

More often

misalignment of

the bearings.

Aerodyna-mic -Fan blades Continuous type Number of

blades X RPM

Magnetic Rotor & Stator

slot harmonics

Continuous type Number of

stator/ rotor

slots their

sums and

difference X

supply

frequency.

Sometimes beat

frequencies signal

can appear due to

interaction of

different

frequencies

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FREQUENCY

OF

VIBRATION

TYPES OF

VIBRATION

CAUSE OF

VIBRATION

NATURE OF

VIBRATION

METHODS

OF

ESTABLISHI

NG CAUSE

REMARKS

Others Mismatching

of radial

ventilation

ducts in rotor

and stator

Continuous type Number of

radial ducts X

RPM

Mismatching of

the ventilating

ducts in the stator

and the rotor

4.3 VIBRATION DIAGNOSTICS CHART

Causes Vibration Level Frequency Remarks

Unbalance Proportional to unbalance. Largest in radial direction

1 x RPM Most common cause of vibrations.

Misalignment Large in Axial direction (50% or

more of radial vibrations)

1xRPM, 2x RPM most common. 3x RPM sometimes.

Base found by appearance of large axial vibrations. Positive diagnosis can be obtained using dial indicators. If sleeve bearing machine and no coupling misalignment; problem is probably unbalance of the rotor.

Mechanical Looseness

Variable 2 x RPM Usually accompanied by misalignment and/or unbalance. Not a loose bearing assembly.

Bent shaft Large in Axial Direction

1 x RPM -

Eccentricity High in Radial and axial direction

1 x RPM and less than 1 x RPM

Some times symptoms will be like unbalance. If that is the case, can be solved by balancing

Coupling Inaccuracies

High in radial and axial direction

Non-synchronous high frequency and Number of

bolts in the coupling x RPM

Check for worn-out bolt holes

Rolling Element Bearing

Distortion

Depends upon amount of distortion

1 x RPM Large component in either horizontal or vertical planes. Taper roller bearings will also have axial components of vibration.

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Rolling Element Bearing Damage

Unsteady High frequencies and multiples of inner race, outer

race, ball spin and cage excited frequency

Harmonics of defective part of the bearing will be excited than other frequencies.

Loose Bearing Assembly

Variable 0.5 RPM Sometimes manifest at rotor critical speed. Quite a common problem.

Bad Gears or Gear Noise

High axial vibration

Very high T x RPM (T =number

of teeth)

Meshing problems can be significant due to increased clearances of gear shaft bearings.

Eccentric Journals

Usually not large 1 x RPM If on gears, largest vibrations in line with gear centres.

Faults in Belt Drives

Erratic or Pulsing 1,2,3 and 4 times rotational

frequency of belt.

Stroboscope can be used to diagnose belt defects.

Inadequate Rigidity

High in radial and axial direction

1x, 2x, & Unsteady Low

Frequency.

Vibration will be high in which direction the rigidity is less.

Rubs No particular characteristic if

continuous

Mainly 1 x RPM plus 2 x RPM

Can excite high natural frequencies of machine. Amplitude at same speed may vary between different runs.

Casing Distortion

Depend on the amount of distortion

1 x RPM & Number of vanes in the impeller x

RPM

High vibration will be felt in the fan or pump casing

Foundation Distortion

Depend on the amount of distortion

1 x RPM & Unsteady Low Frequencies.

Check the foundation.

Piping Force High axial vibration

1 x, 2 x RPM High vibration will be felt on the connecting pipe lines.

Friction Induced Whirl

Severe in radial vibrations

Usually less than 0.4 x RPM and equal to first

critical speed of rotor

Rare. Can be caused by loose rotor components.

Reciprocating Forces

- 1 x RPM, 2 x RPM and higher

orders

Inherent in reciprocating machinery.

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Faulty Combustion in Diesel Engines

High 0.5 x RPM Faults with injectors, fuel pumps, calibration or timing show unbalance. Resiliently mounted unit rocks at 0.5n which is close to natural frequency of mounted unit, and causes large amplitudes.

Aerodynamic or Hydraulic

Variable 1 x RPM or b x RPM (where b is number of blades

or lobes)

Rare as a cause of trouble, except in cases of resonance.

Resonance High radial & axial vibration.

1 x RPM Scientific method to be applied to solve the problem.

Electrical Low, Disappears when power is

turned off.

1 x RPM or 1 or 2 times power

supply frequency

Check for uniform gap between Stator & rotor and windings.

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Notes:

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5. ACCEPTABLE VIBRATION STANDARDS

5.1 INTRODUCTION International standard organization (USA) and various other institutes, world over have standardized vibration standards for various types of equipment/machine/systems. They are called acceptable levels of vibration for the specified system. Each piece of equipment has to undergo acceptance vibration test based upon those standards. The standard shown at ISO-10816-1 deals with over all acceptable level of vibration. In the case vibrations transmitted though foundations, structure or floor are also sensed on the machine. These standards are required to be adopted very intelligently as installation configuration at specific site induces and subtract vibration due its own impedance. To a beginner these standards provide the guide lines, but ultimately one has to generate his own standard for this specific equipment based upon data collected during operation. Experience shows that till vibration levels are kept within the limit of 2 times on high side nothing happens to equipment (Data based not standard based), care has to be taken when rising trend is seen and cut off point can be assumed when it touches 2 times level. 5.2 MACHINERY CLASSIFICATION IN ACCORDANCE WITH ISO 10816-1 Class 1 : Individual parts of engines and machines, integrally connected with

the complete machine in its normal operating condition. (Production electrical motors of up to 15 Kw are typical examples of machines in this category)

Class 2 : Medium-sized machines, (Typically electrical motors with 15 to 75 Kw

output) without special foundations, rigidly mounted engines or machines (up to 300 Kw) on special foundations.

Class 3 : 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 4 : Large prime movers and other large machines with rotating masses

mounted on foundations which are relatively soft in the direction of vibration measurement (for example turbo-generator sets, especially those with light -weight substructures).

Class 5 : Machines and mechanical drives system with unbalanceable inertia effects

(due to reciprocating parts), mounted on foundations which are relatively stiff in the direction of vibration measurement.

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Class 6 : Machines and mechanical drive systems with unbalanceable inertia effects

(due to reciprocating parts), mounted on foundations which are relatively soft in the direction of vibration measurements ; machines with rotating slack-coupled masses such as beater shafts in grinding mills; like centrifugal machines with varying unbalances capable of operating as self-contained units without connecting components; vibrating screens, dynamic atigue-testing machines and vibration on exciters used in processing plants

ISO STANDARD 10816 - 1

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Stands for Acceleration Enveloping

5.3 NATIONAL AND INTERNATIONAL CODE AND STANDARDS ISO - 2372 Mechanical Vibration of Machines ISO - 2373 Mechanical Vibration of Certain Rotating Electrical Machinery. NEMA-MGI-20 Motor and generator Balance Tolerances IEC - 222 Methods of Specifying Auxiliary Equipment for Vibration

measurements. API-RP-541 Recommended Practice for Form-Wound Squirrel-cage

Induction Motors, 200 hp and Larger API - 610 Centrifugal Pumps API - 611 General Purpose Steam Turbines API - 613 High Speed, Special Purpose Gear Units.

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API - 616 Combustion Gas turbines

API - 617 Centrifugal Compressors

API - 618 Reciprocating Compressors

API - 670 Non contacting vibration and Axial Position Monitoring

System.

API - 678 Accelerometer - Based Vibration Monitoring System.

pva course material 03.08.10

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