28 Posts

The Power of Water

This is a great example that shows how powerful water can be in destroying a bearing, after only 1 week, and also highlights that when you perform vibration analysis the normal Velocity data should never be forgotten about.

This case history comes from a great friend of mine Matthew Plant linkedin.com/in/mathew-plant-46000456 .

Matt collected the vibration data and performed the analysis and recommendations on his findings to his client.

 

Asset:

The asset is an Automotive Dynamometer test system in an altitude test facility, the bearing supports the dynamometer rolls. The unladen (no vehicle) rolls shaft weighs 3 tonnes and speed is variable from 0 to 720 rpm (0-250kmh). There is a SKF 22228CCKW33 installed at both shaft ends, the bearing in question however is location end, all radial loads are within spec.

 

Background:

This forms part of a routine maintenance condition based monitoring program. The client reported activation of the facility water sprinkler systems and a service inspection was scheduled to ensure no asset was damaged due to the water sprinkler activation.

 

Vibration Survey:

All the data is the survey before the incident and the survey after the incident. The data was collected one week after the incident due to de- contamination works.

On analysis of the vibration data the following points were noted;

Velocity Data:

Figure 1 is the overall Velocity trend. The overall Velocity increased from 0.137 mm/s RMS to 0.602 mm/s RMS. Even still low this was an increase of 440%

Fig 1:

 

Figure 2 compares the before and after incident Velocity spectra’s. This clearing indicates a change in the bearing condition after the incident. The top green plot is after the incident on site.

Fig 2:

 

Figure 3 is the Velocity spectrum and this show activity that is dominated by the bearing outer raceway defect frequency.

Fig 3:

 

PeakVue Data:

Figure 4 is the PeakVue Max Peak trend from before the incident at 2.019g’s and after the incident at 5.623 g’s

Fig 4:

 

Figure 5 compares the PeakVue spectra’s from before (Blue plot) and after the incident (Green plot).

What can be seen is a 3.566 Order and harmonics. This 3.599 Order is the fundamental defect frequency for the SKF 22228CCKW33 installed. You can also note that 2XBSF is the highest frequency.

Fig 5:

 

Acceleration Data:

Figure 6 compares the before (Blue plot) and after (Green plot) of the raw Acceleration time waveform. This also indicated a high increase in the acceleration impactive data (note crest >5).

Fig 6:

 

Figure 7 compares the before (Blue plot) and after (Green plot) Acceleration spectra’s. This also shows an increase in the friction and impactive levels.

Fig 7:

Vibration Analysis Summary and Recommendations:

Due to the high increase in all vibration parameters and defect frequencies evident for the bearing outer raceway and rolling elements it was advised to replace the bearing.

What was the ‘Alarm bell’ for the analysts was the Velocity data.

 

Root Cause

Obviously water ingress was the instigator in the corrosion; however it was noted that the SKF SNL housings should withstand wash down. Further inspections pointed to the cap lifting eye being absent allowing water to enter the enclosure through the W33 lubrication groove.

 

Bearing Inspection:

On inspection the damage due to the bearing after running one week after the water incident was highly evident.

Outer Raceway.

Outer Raceway.

Inner Raceway.

 

Bearing Replacement:

The new bearing was then installed using the hydraulic nut drive up method.

 

A reliable plant is a safe plant

…..an environmentally sound plant

….. a profitable plant

……a cost-effective plant

Introduction:

Fluid film bearings are mainly monitored with proximity probes. It is often stated that “you can’t detect early defects in fluid film bearings with normal vibration techniques (Velocity, Acceleration or a bearing condition unit)”. But in fact you can detect the effects of a fluid film bearing deteriorating with a normal accelerometer.

Under abnormal circumstances metal to metal contact might occur, leading to occasional high-frequency noise that can be detected with normal vibration equipment. The following case study is a great example of this and also using lubrication analysis as part of a maintenance program.

This case history covers a Production facility Extraction Fan which has been monitored as part of a site wide Condition Based Monitoring program. The drive of this program is to integrate condition monitoring techniques and to drive the maintenance program.

This fan unit had a motor that is a direct drive to a fan shaft, the fan shaft has a white metal fluid film bearing. The fan is a standard overhung centrifugal of about 2.5 meters diameter.

This data was collected over an extended period by myself and Ian Graham.

 

Vibration Analysis:

Ian Graham flagged this reliability risk very early in the program for having a considerable 1 Order impact present in the PeakVue™ data (see Figure 1).

Figure 1 shows the initial Vibration data collected on the fan bearing with the 1 Order event present.

 

 

Vibration Trend Analysis:

Figure 2 is the PeakVue™ Trend of the bearing as it was nursed through until maintenance could be conducted.

This showed the initial level, reduction in levels after an oil flush, then a period of monitoring until another oil change and a bearing inspection.

 

 

Lubricant Analysis:

As the initial fan data had a dominant 1 order event and was being monitored on a monthly basis, we needed to determine whether the event was consistent or deteriorating, and what possible causes were.

Oil samples were then taken (see image 1), on visual inspection the oil was in a very poor contaminated condition. The lab report (see image 2) stated a serious concern with particulate matter contamination and high levels of Sodium, Iron, and Tin.

The indications of Tin suggested probable sleeve (Babbitt) wear of the bearing.

Image 1 is the Fan DE bearing oil sample:

Image 2 shows the initial oil sample report and diagnosis with emphasis on the high Tin levels:

 

Maintenance Actions:

This fan was a critical component of the facilities production process, however with a plant upgrade planned for the very near future, the decision was made to closely monitor the deterioration of the assembly rather than to rectify this potentially expensive piece of equipment.

The temporary measure of an oil flush and change was conducted immediately, with a visual inspection planned for the next shutdown.

 

Maintenance Inspection:

During the shutdown the bearing housing was split and the bearing shells separated. The damage to the bearing was very extensive with ‘scalloping’ of the Babbitt material evident in the direction of rotation forming a build-up at the end of the lower shell. The cause of this is most likely the failure of the lubricants ability to sustain an adequate oil wedge between the shaft and the bearing.

Image 3 shows the Fan DE white metal bearing upon inspection:

Image 4 shows the Lower half shell of the white metal bearing with a piece of ‘free floating’ Babbitt that was found in the sump:

 

Conclusion:

In conclusion, with the company’s utilisation of all the available Condition Monitoring technologies and tools, they were able to monitor and be consistently and accurately informed of the state of deterioration of the bearing. This allowed them to implement a rolling program of temporary measures to stave off what was essentially an unserviceable critical machine until the Factory upgrade was conducted.

.

A reliable plant is a safe plant

…..an environmentally sound plant

….. a profitable plant

……a cost-effective plant

 

 

_________________________________________________________________________________________

7th August 2018: Additional data as requested by JUAN CARLOS URQUIOLA

Below is the Velocity data from September 2017 and June 2018, there is a difference in the number of an amplitude of the 1 order as associated harmonics.

 

Below is the Acceleration high frequency spectrum, this shows two mounds of activity that has side bands of 1 order and 2 orders.

 

What happens when recommendations are not followed – “when things are left to burn”.

How often have you performed a reliability survey and issued a report of findings and recommendations to reduce the risk of unplanned system failure… and the client does not follow the recommendations.

This is one example of an infrared thermal imaging survey that highlights the importance of following the recommendations and also that a thermal survey should be performed by an experienced/qualified reliability technician who does not just rely on the thermal camera to rush round the site but also uses the human senses and experienced to assess system condition.

 

 

Initial Survey:

One panel unfortunately had Perspex in the way of the cable terminations, so this could not be surveyed with thermal imaging. Through the perspex cover it was noticed that the cable sheath has split, probably due to excess heat and exposing the copper cable.

This was reported on the day to the site supervisor, and in writing in the report. Site confirmed that they were going to schedule in repair at the soonest opportunity due to the high unknown probable risk.

This is the thermal image of the panel, note no readings as infrared energy doesn’t pass through Perspex.

This is the digital image of the panel showing the Perspex cover and damaged cables.

 

 

Unplanned Failure:

This was not inspected/repaired and the panel caught on fire. This caused shutdown of the plant and a huge costs to the company in downtime and reputation due to unfilled orders to their customers.

Images of the failed component.

 

Repair:

Image of the repair. Here you can see the burn fire marks on the back panel.

 

 

Conclusion:

Sometimes we try our best to ensure our clients do the right thing for reliability on their plant. Unfortunately they don’t always action what we recommend, not matter how much we try to convince them. In this instance all we can do is keep spreading the word of how important it is to know the condition of your system and then to actually action any risks. This in turn will reduce the risk of unplanned failure.

A special thanks to James Pearce for sharing his experience.

 

Recently I saw a post from Terrence OHanlon of Reliabilityweb.com, that I feel summed up Reliability.

A RELIABLE plant is a SAFE plant

…..an ENVIRONMENTALLY SOUND plant

….. a PROFITABLE plant

……a COST-EFFECTIVE plant

This month’s blog is to promote the thinking that when drive trains are aligned they should be aligned to the bearing tolerances and not the coupling tolerances. In addition how many people receive an alignment report with a soft foot check? We have found that some companies allocate their employees a laser alignment kit tell them what buttons to press and send them in the field. Without proper training and mentoring how will these employees learn correct Precision Alignment? Without correct training they will not know how to fix problems if they don’t understand fully what they are doing.

This month’s blog shows the importance of Precision Alignment including soft foot check and that the users of laser alignment equipment should be properly trained and mentored in Precision Alignment.

This survey was conducted by a great friend of mine and recent VCAT 3 Certified Seasoned Analyst James Pearce. linkedin.com/in/james-pearcevibrationanalysis

 

Background:

We were called to investigate an apparent increase in vibration levels after a high pressure hot water pump was replaced with a new pump end and a reconditioned drive motor. The operator felt that it was not running as smooth as the old pump set.

 

Instrumentation:

For this survey James used the CSI 2140 Dual channel Machinery Health Analyser. Data analysis was carried out using the CSI AMS Machinery Health manager software V5.61.

 

Methodology:

Vibration data including Velocity, Acceleration and bearing condition unit PeakVue was collected from each bearing location as close as possible to the source. Where applicable additional data including high resolution vibration data was collected.

 

Executive summary:

There are elevated directional Velocity vibration levels when running at 2680 RPM (Low speed). This is due to a coincidence of a system natural frequency being excited by a motor Soft Foot condition.

 

Maintenance Recommendations:

  • Check/inspect condition of the foundation, looking for looseness and any deterioration in the base plate.
  • Perform precision alignment that must start with a soft food check and soft foot elimination. Followed by precision laser alignment.
  • If these actions do not resolve the issue then stiffening of the base may allow for improved precision alignment and may move the system resonance out of the running speed range.

 

Analysis Summary:

  • There are increased directional Velocity vibration levels at the motor when running at 2680 RPM.
  • After reviewing the vibration data it was decided to perform further checks and the motor holding down bolts was loosened one time when in operation, this is to check for distortion of the motor rotor to stator air gap. During this test it was and found that the Velocity amplitude reduced. The amplitude reduced to its lowest level when the motor non-drive end foot bolt (bolt closest to pump #1) was loosened (see figure 1&2). This indicates there is a soft foot issue.
  • In addition an overall vibration coast down test & resonance bump test was performed. This data confirmed a natural frequency at 2X 2680RPM (see figure 3).

 

HPHW Pump #2 Motor Non-Drive End

The motor has elevated directional Velocity vibration levels. By loosening one motor fixing foot bolt at a time, the Velocity amplitude reduced. The amplitude reduced to its lowest level when the motor non-drive end foot bolt (bolt closest to pump #1) was loosened.

Figure 1 compares the Velocity spectra when running at 2680RPM, for the one order levels, as found state (4.332mm/sec RMS) & where the amplitude decreased the most after the motor foot bolt was loosened (2.651mm/sec RMS).

Fig 1:

 

HPHW Pump #2 Motor Non-Drive End

Figure 2 is a photo of the motor indicating which foot bolt was loosened which resulted in the best decrease in amplitude.

Fig 2:

 

HPHW Pump #2 Motor Drive End

Figure 3 is the data from a resonance bump test & overall vibration coast down test, performed at the motor drive end (DE).

The top plot bump test result indicates a system natural frequency that will coincide with twice the running speed (when running at the low speed) and amplify the vibration levels.

The bottom plot amplitude peak from the coast down test also confirms this condition with a peak at 5336 RPM, twice the running speed at the low speed setting.

Fig 3:

Hello. This is a different avenue than my usual posts, this one is a link to a paper I have just written with the great knowledge and help from Dr K.

 

Briefly:

I was introduced to Dr Knezevic (Dr K) of the Micre Akademy though a great mentor in thermal imaging Austin Dunne of the Institute of Infrared Thermography.

Please click on this link to learn more about the great work of the Micre Akademy

Please click on this link to learn more about the Institute of Infrared Thermography

 

About this Paper:

MIRCE Science is a theory for predicting expected functionability performance for a functionable system type. Accuracy of the predictions is governed by the degree of the scientific understanding of the physical mechanisms, and human rules, that govern the motion of functionable system types though MIRCE Space. The main objective of this paper is to address vibration monitoring as one of the possible mechanisms that governs motion of a gearbox through functionability states, which are contained in MIRCE Space. In general, and to illustrate this process through a case study related to heavy gearbox used in Plastics Manufacturing industry, conducted by the author with vibration data collected on site by Ian Graham.

 

Click here for the paper Vibration Monitoring Mechanism Of Motion

 

Acknowledgement:

The author wishes to acknowledge the support received from Dr Knezevic, MIRCE Akademy, Exeter, UK, while preparing this paper. As the “father” of MIRCE Science, Dr Knezevic, has inspired me to view how every day Condition Based Monitoring can have a significant impact on functionability performance of the whole system.  Consequently, I can now understand how many companies are performing Condition Based Monitoring but are not linking this to the business performance of the whole organisation/company. MIRCE Science is the body of knowledge that bring together these two very different but related disciplines, for the ultimate benefit of the user.

This is one I recently finished and thought it would be a great one to share so people know what can be achieved.

 

Background:

We had three pump sets suffering from elevated vibration levels when operated in different combinations. Conventional vibration analysis was performed and this indicated a structural resonant condition.

The pump motors are mounted on a false floor:

and the pump barrels are below the floor:

The pump with the worst motion was on pump 3, the one far away from the edge of the drop. Also this pump has the least structural support under the floor. When ran in certain combinations pump 3 would be excited very badly.

 

The cost effective solution.

I designed a vibration dynamic absorber.

Dynamic Absorbers are often overlooked and not used, they can be seen as a band aid or a last option for some vibration problems. Whereas in some cases they can be the only cost efficient option, and they are very effective.

The Vibration Dynamic Absorber is a unique bespoke item, maintenance free, that is designed to absorb unwanted energy. It is tuned to have the same resonant frequency as the structure to set up an out of phase signal reducing the signal generated by the structure.

 

How did I design these?

For this one it was more of a ‘gut feel’. I looked at the motor and then drew out a design that wouldn’t look out of place when mounted, and that had some adjustment to it when fitted as theory doesn’t always pan out in real life. Then from this I worked backwards to get the correct material dimensions/configuration so it was resonant at the target frequency. I also made some weight configurations so I could cover my target range.

I will be going back in 6 months to see how it fairs. I did consider a round bar and weight but thought that with the rectangular bar you have more control on what way it will be resonant. As once you have performed phase analysis on the motor you then know what way it is moving and can mount the absorber accordingly.

Image of Pump 2 Vibration Dynamic Absorber:

Image of Pump 3 Vibration Dynamic Absorber:

Pump 2 Live Motion Video

 

Pump 3 Live Motion Video

 

Pump 3 Slow Motion Video

 

What am I covering?

On pump 3 I am covering the one problem frequency, 1 Order, but the two arms are of different lengths in terms of the length from the point of pivot (clamping) to the mass. Also the arms are of different dimensions with different mass at the end so they could be tuned to the same frequency.

I also did find that the sweet spot was not necessary the point of higher deflection of the absorber and that the three motors all reacted differently.

 

Final Review of actual vs theory:

I have had time to review the final theoretical tuning of the three pumps to actual results. They are all different and no one motor is the same, they all have their own personalities dynamically wise.

Pump 3 had the highest overall vibration, one dominant frequency at 1 Order on pump 3 and this was successfully reduced.

Pump 1 and pump 2 had two frequencies in the data. And both of the vibration dynamic absorbers were tuned to the lower frequency not the one order.

 

Table of final overall levels:

Before After % reduction
Pump 1 Motor NDE (Top) 4.314 2.854 33.84%
Motor DE (Coupling end) 2.092 1.617 22.71%
Pump 2 Motor NDE (Top) 9.95 6.959 30.06%
Motor DE (Coupling end) 4.05 3.012 25.63%
Pump 3 Motor NDE (Top) 27.02 7.59 71.91%
Motor DE (Coupling end) 10.73 5.113 52.35%

 

Pump 1:

Pump 1 actually showed the text book results. The theoretical calculations for the tuned damper was for the lower frequency not the running speed (1520 CPM yes they are on soft start VFD). It actually split the frequency – text book……….beauty!!

 

I have more questions and theories now, this is pretty exciting stuff. Hopefully I can keep this going on other pumps.

Hello all

This month’s blog is slightly different from the usual ones we post. This month is more of an opinion regarding data dogs. We are seeing more equipment suppliers selling VA equipment that they promote as “anyone can use” and you know need to know or have experience to use. That the software will diagnose for you. Or even, just collect the data upload to the cloud and we will tell you if you have any issues.

I feel there are places for this type of program but one thing I dislike is companies sending “data dogs” to collect the data. These are cheap labour sent to press a button and collect the vibration data as fast as they can. This type of VA often gives this service a bad name as they miss diagnose, miss defects or the person in the office performing the analysis just gives the ‘wall chart analysis’ of its either misalignment, imbalance, looseness or resonance.

So much can be gained by a competent engineer or technician attending the asset to collect the vibration data. Most of your analysis should be performed at the machine, not in the air conditioned office!

We also find that there are many facilities/companies that are on the start of their reliability journey that require a person on site to promote and ensure the job is done and followed though correctly.

 

An Example:

The images below back up this point. A great friend of mine, James Pearce, was performing a quarry motor VA survey and while at a motor he sensed an abnormal noise, he tracked it down to the GTU take up conveyor pulley. The GTU is not on the vibration program but when you have an experienced engineer or technician collecting the data walking the plant they also use their other senses to ensure plant reliability.

 

James reported this to site that had a controlled shut down of the quarry immediately to replace the pulley bearings. Site confirmed that they would have not inspected this pulley and it would have catastrophically failed causing a lot of additional hard work. This controlled shutdown cost 3 hours of production. But this saved replacing the pulley shaft as there was no damage to the shaft. If this was left to totally fail this would have cost 9-11 hours production downtime at 2,000 Tons per hour. There is also the possibility that the pulley could have failed in a way that caused damage to the conveyor belt incurring more down time and a lot more costs.

 

 

And here is the video!

 

You can see the bearing there – this should not be glowing red. This bearing had totally failed!

So remember that 5 years of experience is not the same as 1 years of experience 5 times and you can’t analyse what you don’t know or understand.

Ultrasound trending and Vibration Analysis working together.

This is a good example of how condition monitoring technologies work well as integrated technologies.

Through routine in house overall ultrasonic dB trending a change in condition was noted from one of the motor bearings and this was an increasing trend. I was called to verify the asset condition through vibration analysis.

 

Executive Summary:

  • Removal of the motor on condition of the bearing enabled a control change-out and a more cost efficient repair rather than running to failure.
  • The cause of the elevated Ultrasonic levels and the vibration defect frequencies was false Brinelling to the drive end bearing.
  • In addition there appears to be grease compatibility problems result from either mixing incompatible greases, or from ingress of other contaminate, Dry powers absorb the oil causing the grease to thicken.

 

Failure Mode:

From inspection the primary failure mode as per ISO 15243:2004 is 5.3.3.3 False Brinelling, there is also a secondary failure mode as per ISO 15243:2004 of 5.2.2 Abrasive Wear due to inadequate lubrication.

False Brinelling occurs in the contact area due to micromovements and/or resilience of the elastic contact under cyclic vibrations. Depending on the intensity of the vibrations, lubrication conditions and load, a combination of corrosion and wear can occur, forming shallow depressions in the raceway. In the case of a stationary bearing, the depressions appear at rolling element pitch.

In many cases, it is possible to discern rust at the bottom of the depressions. This is caused by oxidation of the detached particles, which have a large area in relation to their volume, as a result of their exposure to air.

Key Points are:

  • rolling element / raceway contact areas
  • micromovements / elastic deformation
  • vibrations
  • corrosion/wear shiny or reddish depressions
  • when stationary: at rolling element pitch
  • when rotating: parallel “flutes”

Abrasive wear. Most of the time, real abrasive wear occurs due to inadequate lubrication or the ingress of solid contaminants. Abrasive wear is generally characterised by dull surfaces. Abrasive wear is a degenerative process that eventually destroys the microgeometry of a bearing because wear particles further reduce the lubricant’s effectiveness. Abrasive particles can quickly wear down the raceways of rings and rolling elements, as well as cage pockets. Under poor lubrication conditions, the cage may be the first component to wear.

 

Bearing Inspection: Motor Drive End Bearing – FAG X-life NU324-E-TVP2-C3

Image 1 is of the poor grease condition from the bearing.

Image 1:

 

Image 2 is an image of the false Brinelling indetention on the inner raceway.

Image 2:

 

Image 3 is a microscopic image of a false Brinelling depression on the inner raceway. Rust at the bottom of the depressions. This is caused by oxidation of the detached particles

Image 3:

 

Image 4 is a microscopic image of the inner raceway showing the over roll of particles.

Image 4:

 

Image 5 is an image of the outer raceway in the load zone showing the false Brinelling. This is only present in the load zone.

Image 5:

 

Image 6 is a microscopic image of a false Brinelling depression on the outer raceway.

Image 6:

 

Image 7 is a microscopic image of a rolling element. Here you can see the flat spot from the false Brinelling. In addition the ring that is around the inner and outer raceway is due to over roll of particles and poor lubrication condition. Flat spot from the false Brinelling Ring of over roll of particles

Image 7:


 

Vibration Data: 

The comparison below show Fan 1 (in blue) and Fan 2 (in green). This highlights the very high destructive levels of the drive end bearing and that it was close to failure.

 

The PeakVue spectrum plot below confirmed that it was a bearing defect and highest at the outer raceway.

Hi All, this is the last post for 2017 – Enjoy

Background:
We were called to inspect a gearbox as the client had reported an abnormal sound. This was a very large old extruder high torque gearbox with a single input and dual output shafts.

Executive Summary:
Through onsite vibration analysis we were able to pinpoint the shaft that was generating the abnormal noises, this enabled the bearings for the shaft to be pre-ordered so they arrived at the repair shop the same time as the gearbox. This ensured a quick turnaround was completed with minimal production loss.

On Site Initial Assessment:
The gearbox vibrational levels as measured under full load conditions were >20mm/s RMS. This is considered “Vibration Causing Damage” as per ISO 10816-3. The Acceleration Peak to Peak impactions at Gearmesh #1 was excessive at 162G’s. There was also indications of misalignment on the 1st intermediate shaft and considerable looseness present. The 1st intermediate shaft ‘binds’ for 1/4 to 1/3 of a revolution when turned by hand.

Vibration Data:
The Input shaft high frequency Acceleration spectra clearly shows a high 2x gearmesh frequency for the gearmesh 1. This indicates there is misalignment within the gearing setup. The sidebanding at 19.20Hz indicates that it is relative to the 1st intermediate shaft.

The plot above is the Acceleration Spectrum from the Gearbox NDE Horizontal.

The Peak to Peak measurements on the Acceleration Time Waveform below indicates the Acceleration forces are within the 1st Intermediate shafting. The total reading of 162G’s is highly destructive and is impacting at 19.2Hz, the 1st intermediate shaft speed.

The Velocity spectrum taken from the NDE of the 1st intermediate shaft shows a considerable amount of run speed harmonics attributed to the shaft speed. This is an indication of looseness.

Cause of Failure:

On inspection the tab washer on the first intermediate shaft outer bearing had failed. In addition the suspected gear on the 1st intermediate shaft was extremely loose. It was found that this shaft had been previously repaired with metal spray and this had failed. On closer inspection the stress raiser appears to be around the keyway, as there was no strengthening welds around the keyway to support the metal spray.

Strip Down Images:

This is an image of the gearbox internal layout.

Images of the failed tab washer found in the bearing cap from the 1st Intermediate shaft.

Image of the key that supported the 1st intermediate gear that was loose.

Metals spray coating that was under the 1st Intermediate shaft gear. This failed initially at the metal spray coating at the keyway.

 

 

G,day all, here is another interesting job I got called to

 

Background:

This pump and motor had a history of reliability failures from bearings, shaft shearing and pipework flanges leaking. This was a pair of pumps on separate base frames but secured to the same concrete floor with a pipework common outlet.

I performed vibration analysis with phase analysis and diagnosed a foundation and structural problems as the root cause.

 

Vibration Data:

Pump Vibration Data:

Figure 1 shows the overall Velocity vibration trend from our first visit and second visit. This is gathered at the motor DE.

From this trend you can see a marked increase in the velocity vibration levels from 6.304 mm/s RMS and 8.388 mm/s RMS.

Fig 1:

 

Fig 2 compares the acceleration time waveforms from the motor drive end bearings

From this comparison you can see the lower levels of pump A (Blue Plot) and the very high impacting from pump B (Green Plot)

Fig 2:

 

Fig 3 is the vibration data from the motor drive end bearing

There is high impacting form the motor bearing and damage to the inner and outer raceway

Fig 3:

 

Pump B – Motor Bearing Inspection

Failure Mode:

From inspection the failure mode as per ISO 15243:2004 is 5.3.3.3 False Brinelling.

False Brinelling occurs in the contact area due to micromovements and/or resilience of the elastic contact under cyclic vibrations. Depending on the intensity of the vibrations, lubrication conditions and load, a combination of corrosion and wear can occur, forming shallow depressions in the raceway. In the case of a stationary bearing, the depressions appear at rolling element pitch.

In many cases, it is possible to discern rust at the bottom of the depressions. This is caused by oxidation of the detached particles, which have a large area in relation to their volume, as a result of their exposure to air.

Key Points are:

  • rolling element / raceway contact areas
  • micromovements / elastic deformation
  • vibrations
  • corrosion/wear & shiny or reddish depressions
  • when stationary: at rolling element pitch
  • when rotating: parallel “flutes”

 

Findings:

  1. Depressions appearing at rolling element pitch indicating damage while the pump was in standby stationary bearing (Image 1)
  2. Indications of oxidation of the detached particles, which have a large area in relation to their volume, as a result of their exposure to air.

 

Bearing Inspection: Motor Drive End Bearing – FAG X-lite NU319E.TVP2

Image 1 is the outer raceway, and displays depressions appearing at rolling element pitch which indicates damage to the bearing when the motor was stationary

Image 1:

 

Image 2 is a close up of the depression at rolling element pitch on the outer raceway, this is from the load side of the bearings and also shows the roll over.

Image 2:

 

Image 3 is a microscopic image of a depression on the outer raceway.

Image 3:

 

Image 4 is an image from the inner raceway, this also displays the depressions at the pitch of the rolling elements.

Image 4:

 

Image 5 is a microscopic image of a rolling element.

Image 5:

 

 

Motion Amplification

Even with this vibration data the client was not convinced so I had to use another technology to show the client how the structural and base was causing them their reliability headache.

 

This first video shows how the pipe work was moving, this was the cause of the stress and strain to the flange joints that led to the leaks

 

This second video is of the base plate, this showed the true motion of the pump and how these failures were being induced.

%d bloggers like this: