The above is a simple but comprehensive schematic to understand and resolve vibration problems of industries.
- Resolving vibration problems
- Design Review
- Machine Testing
The above is a simple but comprehensive schematic to understand and resolve vibration problems of industries.
Recently I received an email which asked me give my option on a phenomenon the analyst observed.
Observing high vibs on pressing and lifting pinion Drive End (DE) and Non Driven End (NDE) bearing on a ball mill. Motor and main Gear Box drive are OK. Clear predominant gear mesh frequency is appearing in the spectrum along with harmonics and side bands. But 1st GMF (Gear Mesh Frequency) is predominant. side bands with pinion speed is also seen. no Girth Gear speed side band was observed
Some of the vibration data, spectrums and photos shown in the attachment. Phase measurements indicate inconsistency in the readings near pressing pinion bearing. Impacts were also seen in time waveform data along with modulation.pinion speed 122 rpm. On pressing side bearing 2.03 Hz side bands are seen, On lifting side i can see side bands spaced at 6.09 Hz (That is 3 times of pinion speed). Both Pinion lifting and pressing bearings are behaving differently. the vibs are high on DE as compared to NDE on both pinions. Can we suspect eccentric moment of the pinions with looseness. Why am i seeing 2.03 side bands on pressing and 6.09 side band on lifting side bearings. What is the significance of this. One sample of GG tooth photo shows uneven shining surface on either side (refer photo). In this case I am seeing (30 T = 1 X 2 X 3 X 5) and (210 = 1 X 2 X 3 X 5 X 7) 2 X 3 X 5 as the common factor. Pinion 30 teeth and GG has 210 teeth. Will this create gear ratio issues uneven locking and releasing of 2 mating teeths. But no 1/2 or 1/3 or 1/5 GMF seen in the data.
My reply was:
1. We are seeing 1/2, 1/3 and 1/5 of GMF — these appear due to common factors 2, 3, 5 as you wrote.
This means that the pinion is badly worn out and as the common factor teeth mesh they generate these fractional frequencies.
It also means that the GMF and the natural frequency are not separated by 2.5 times. [The natural frequency in the horizontal direction = 28.5 Hz; natural frequency = 30.9 Hz; Gear Mesh Frequency = 60.6 Hz]
Looking at the signatures it is clear that the GMF falls within 2.5 times the natural frequencies.
Also note how the GMF (60.6 Hz) falls right between two natural frequencies in both the vertical and horizontal directions. (31.1 Hz and 83 Hz). This makes the situation worse.
It is important to question as to what else we can do other than detect a problem or detect an incipient fault?
With the above analysis and information we can easily see the relationship between fractional GMF and lubrication and wear. It means we can build an algorithm that would warn us about an imperfect lubrication system that would in fact accelerate wear and put the system out of service.
Further, we can refine the specification of a purchase a gear box. The specification should state — a) number of pinion teeth should be a prime number to prevent accelerated wear b) if a prime number can’t be achieved then the natural frequency in the three directions must be away from the GMF by at least 2.5 times the GMF.
Similarly, we can specify the gear box top cover natural frequency should be at least 4 times the GMF.
Scheduled running checks may include — a) rate of lubricant flow b) motor current c) placement of lubricant nozzles etc.
Details of the case: (relevant data)
VIBRATION STUDIES OF CEMENT MILL
Steps for vibration measurements
Impact test was carried out at selected locations on the torsion bar to know its natural frequency
Normal vibration signatures were recorded with motor speed being 994 rpm and pinion speed 122 rpm
Vibration data was recorded on selected bearing locations of motor, gearbox and pinion bearings
Data was recorded along horizontal, vertical and axial direction with 90% load on the mill
Phase measurements were recorded to know the behavior of pinion DE with respect to pinion NDE of pressing and lifting side
Vibration signatures recorded on Pinion DE and NDE of both pressing and lifting side bearing shows predominant gear mesh frequency and its harmonics
Side bands were observed along with gear mesh frequency and its harmonics
Gear mesh frequency 60.9 Hz is appearing predominantly in all HVA direction
Time waveform recorded on pinion DE and NDE bearings clearly shows modulation which occurs due to above phenomena
Impacting of the gear teeth was also observed. Refer time plots provided in this report in subsequent pages
Only Side bands of pinion speed (2.03 Hz or 122 rpm) are seen, no side bands of Girth gear was seen in the data
The phase measurements recorded on pressing pinion DE and NDE along axial and horizontal direction shows the phase is not consistent with time suspecting looseness due to uneven movement of pinion
Vertical vibrations recorded on pinion DE bearing lifting side shows the vibrations are low (5 mm/sec) on one end while its high (11 mm/sec) on the other end even though it’s a common top cover of that bearing
For any normal 2 mating gears the selection of no. of teeth on each gear should be such that when factorizing is done no common factor should be found apart from 1
In this case pinion has 30 teeth and Girth gear has 210 teeth o Then as per calculations
Pinion 30=1x2x3x5,GG 210=1x2x3x5x7
So common factors are 2x3x5
This case is about a sudden failure of cooling tower fan motor of a copper mine.
The motor failed almost immediately after Planned maintenance, which was just about lubricating the motor bearings.
Electrical department conducted a scheduled PM task on this piece of equipment on 25.05.17. After 3 hrs of running; motor Non Driven End (NDE) bearing was damaged.
When the motor was opened it was observed:
1. One of the poles was severely damaged.
2. Bearing cage was also found damaged and all roller elements were crushed.
1. The vibration signature did not indicate lubrication starvation of the bearing.
Hence the question is — why stop a machine for re-lubrication when the activity isn’t needed at all?
2. In the future, if the system is stopped, then during start up it has to be ensured that the load is zero or near zero or it has to start at no-load condition.
If that isn’t possible, the system has to be started at low rpm and then the rpm can be gradually increased, all the while maintaining a steady state. It might take up to 6 hours for the system to stabilise after a bearing is lubricated.
Report on Thaisen Fan (Scrubber)
After cleaning of the fan blades, vibration of the fan gradually increases during operation and in a span of 10 to 14 days vibration level reaches an unacceptable level, which necessitates the next cleaning cycle. However, for so long, this matched the scheduled production window provided by operation. However, after the recent changing of the rotor and the bearings, the fan now reaches unacceptable level of vibration within a short span of time that does not coincide with the scheduled “production window” of the operation, which causes “unplanned downtime.”
Goal of the investigation: To correct the imperfection in the system so that the fan cleaning cycle coincides with.the scheduled production window.
Result of the investigation:
1. The problem of rising vibration within a short period of time is an inherent problem (a birth defect) of the fan. The main reason is the Coriolis effect on the fan. Coriolis force is a force exerted by a moving fluid on the disc or impeller rotating in the fluid. If the rotation is CCW (counter clockwise) then the fluid moves to the right of the impeller and away from the centre. Similarly, when the impeller moves in the CW (clockwise direction) the fluid moves towards the left of the impeller and away from the centre.
In this case, with the fan moving in the CCW direction the Coriolis force moves toward the right of the impeller in the same direction as the damping force. This effect (the fan moves in the CCW direction) produces negative damping (since the two forces are in the same line of action).
Negative damping is a phenomenon, when damping force, which usually opposes the driving force, acts in the same direction as the driving force. In such a case the vibration of the system is amplified.
Combination of negative damping and Coriolis effect produces this phenomenon of gradually rising vibration of the fan in a short period of time, which goes away upon regular cleaning. In the present context nothing can be done to eliminate the phenomena of Coriolis Effect and Negative damping. However, if a similar system is to be installed in the future, we would be pleased to provide necessary suggestions and recommendations so that such phenomena are eliminated right from the start.
2. Present signatures indicate misalignment and dynamic imbalance
3. Weak foundations
Actions to be taken to increase the cleaning cycle to match scheduled “production window.”
1. Take care to align the rotor properly. Care to be taken while putting shims.
2. Dynamically balance the fan in two planes to eliminate the imbalance
3. Cleaning cycle can be initiated when vibration of the fan on the bearings reaches 7 mm/sec (rms). It is safe to run the fan upto this point.
4. Monitor the condition of the foundation by taking vibration measurements in displacement and acceleration modes. Displacement should be taken in the horizontal direction on the topmost accessible point of the columns and at the base. Acceleration should be taken in both vertical and horizontal directions. Displacement should not cross 50 microns in the horizontal direction or at the base of the columns. Similarly acceleration both in the vertical and horizontal directions must not cross 1.5 g. This would ensure safety of the equipment. In case it crosses corrective actions are to be taken to rectify the foundation.
Speed dependent vibration is associated with forced mechanical vibration.
Application: rolls where the strips processed through the rolls exhibit roll chatter that leaves permanent imprint on the strip in the form of chatter (equally spaced markings of about 20 to 45 mm width) pattern. It is generally considered to be a defective product and often can’t be sold in the market.
The way to check the cause of such chatter patterns or marks is to take the vibration in displacement mode. When displacement increases by approximately 0.6 microns at the highest rolling speed it significantly points to surface roughness of the strip and so creates the pattern of chatter marking on the product (e.g. aluminium sheets). It indicates a loss of stiffness or the presence of variable stiffness, which may be coming from coupling, defective gears or from loose or defective anti-friction bearings.
Usually we may observe sidebands on either side of the forced vibration peak. The spacing of the sidebands is an exact multiple of the rotational frequency of the work roll. This is commonly seen for inner race defects where the inner race is rotating freely on the roll. The defect rotates through a variable load zone and produces a modulated time waveform. This is seen as a peak with sidebands in the vibration spectrum. Also pronounced on chocking and de-chocking.
This eliminates the strip chatter or markings.
Reasons of bending:
Phase measurement is an effective test to confirm presence of bent shaft. Phase at 1N measured in the axial direction at opposite ends of the components will be 180 degrees out of phase.
However, if the phase measurements are taken around the shaft we would find that they are all in phase since the shaft will appear to be moving back and forth in the axial direction.
In addition to the prominent presence of 1N and 2N in the axial direction we would also find higher than normal 1N and 2N peaks in the radial directions.
In this case time waveform would not prove to be a good indicator for bent shaft. However, a sinusoidal waveform is expected in the axial direction if the vibration is predominately 1N. In the case of a predominate presence of 2N there would be a “wobble” depicting the classic “M” or “W” pattern depending on the phase angle, if the bend is closer to the coupling.
Typical Symptoms: High 1x in the axial direction and 2x in the radial directions; at time 3 x is also present in severe cases (e.g. when coupled to coupling imbalance).
Reasons for misalignment:
Types of misalignment:
Usually we would find high 1x peak in the axial direction with small 2x and 3x peaks depending on the “linearity” of the vibration. There may be both 1x and 2x (at times accompanied by 3x) in the radial directions.
Time waveform in the axial direction would be dominated by sinusoidal 1x vibration
Phase: Motor and say Pump would be out of phase axially due to angular misalignment (across the coupling in the same direction).
Typical Symptoms: 1x radial (in Vertical and Horizontal directions)
Like eccentric pulleys, Eccentric gears generate strong 1x radial components, especially in the direction parallel to the gear.
They would also generate sidebands of the running speed of the eccentric gear around the GMF (gear mesh frequency). However, harmonics of GMF may also be generated (depends on the severity of the problem). Natural frequency might also be excited.
Time waveform: The waveform will have combination of 1x running speed of input and output shafts plus strong gear mesh vibration modulated by the running speed of the shaft having the eccentric gear.
Phase: Not applicable.
Typical Symptom: High 1x in the direction parallel to belts. Though 1x component can be found on both Vertical and Horizontal directions.
Instead of the typical Vertical and Horizontal directions it is best to choose the directions parallel and perpendicular to the belts.
The high 1x can be found on both sub-assemblies (e.g. the motor and fan). Since the motor and the fan would run at different speeds we would also find two distinct peaks on the signature corresponding to the motor and fan running speeds. Confirmation about which pulley is eccentric can be obtained by removing the belts and checking for the presence of high 1x on motor in the direction parallel to the belts.
Time waveform would be sinusoidal when viewed in velocity.
Phase: Phase reading taken parallel and perpendicular to belts will either be in phase or 180 degrees out of phase.
The inherent reliability of a system is determined by the system’s design. It means that the design of the system would determine the upper limit of reliability the system exhibits during operation. Suppose, for example, a system, with the best possible maintenance is able to achieve availability of say 90% we can say that this is the upper limit of the system’s capability that is determined by its design. A good “preventive maintenance” plan can never improve a systems inherent reliability. In other words, preventive maintenance, contrary to what many believe, cannot make a system “better”. It may, at best, only help realise the inherent reliability as determined by the physical design.
Hence the suggested process to “improve” the inherent reliability of a system, may be framed as follows: –
Understand the dynamics through tools like vibration analysis
Monitor changes and rate of change
Eliminate unnecessary maintenance tasks
Change the design of the system interactions to eliminate inherent “imperfections” and revise the maintenance plan.
In most cases, this would be the general approach.
Until we can effectively undertake some design changes (Design Out Maintenance – DOM) or take measures to eliminate inappropriate maintenance actions (Review of Equipment Maintenance – REM) it would not be possible to go beyond inherent reliability of an equipment, specially if it is undesirable in the business context. For example, a vertical pump of a power plant kept failing very frequently or had had to be stopped quite often when vibration shot beyond the trip limits. This behaviour of the system is determined by the design of the system. Unless the design (specifically the interactions between components) is corrected for improvement; the system (vertical pump) would continue to behave in that manner for all times. Likewise if the MTBF of a machine is say 90 days, it would not be possible to considerably improve the MTBF way beyond 90 days unless some undesirable interactions (which I call system “imperfections”) are corrected for improvement and a proper review of existing maintenance system is carried out. Such “imperfections” can be both physical and non-physical. Design features, most importantly, the interactions between physical/non-physical components are arguably the most important characteristic of a system that determine a system’s inherent reliability.
In addition, there are many physical design features that influence reliability like redundancy, component selection and the overall integration of various pieces of the system.
In the context of RCM, design extends far beyond the physical makeup of the system. There are a number of non-physical design features that can affect, sometimes profoundly, the inherent reliability of a system. Among these are operating procedures, errors in manufacturing, training and technical documentation. When a proper RCM analysis is conducted on a system or sub-system, there’s a good chance that the resulting maintenance actions will enable the system to achieve its inherent reliability as determined by its physical design features. However, if the inherent reliability is below user’s expectation or need then the design features are to be improved to achieve the desired level of inherent reliability.
Moreover, if unwarranted maintenance tasks are eliminated as it will greatly reduce the risk of suffering the Waddington Effect. There is also a good chance that if operating procedures, training, technical documentation and so forth are found to negatively impact inherent reliability, these issues will be identified and corrected. As evidenced by the Waddington Effect. In virtually every case, less than optimal, non-physical design features almost always have a negative impact on inherent reliability. Therefore, in RCM analysis a through review of existing maintenance plan (REM) along with DOM is necessary to improve inherent reliability of a system.
In brief, right amount of Condition Based Maintenance (CBM) tasks, Scheduled Inspections (which is a part of CBM activity) REM and DOM would not only help us realise the inherent reliability as determined by the physical design but also improve it, if the original inherent reliability is below business expectation.