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Equipment is being built lighter and cheaper. This means that resonance has become more of a reliability problem with equipment. Most engineers,  CM analysts, mechanics and managers are not aware of how resonance may be affecting their equipment. Resonance frequencies will excite any vibration occurring at or near the same frequency. This can include misalignment, unbalance, bearing faults or other defect frequencies. This will cause your equipment to fail more quickly as well as other unwanted effects.
You should to be aware of and document the resonance frequencies affecting your equipment. Many methods can be used to determine the resonance frequencies in your equipment and a good vibration analyzer will have the resources to help you do so. Methods such as an impact test (bump test), negative averaging, startup, coast down, etc. can be used to identify resonance frequencies. Additionally, the vibration analyst should look for signs of resonance-related issues when doing routine equipment analysis. For example, always look at the amplitude ratios between horizontal and vertical vibration measurements. Ratios of 3 to 1 or higher (horizontal versus vertical) are an indication of resonance issues in the equipment being monitored.
What should you do once a resonance problem is known and the unwanted reliability consequences understood? It is important to keep equipment operational speeds away from these critical frequencies by at least 20-30 percent. Actions can be taken to shift these critical frequencies and minimize and/or eliminate their negative effect on your equipment reliability. The primary methods are:

  • Add Mass: Adding mass will lower the resonant frequency.
  • Add Stiffness: Adding stiffness raises the resonant frequency.
  • Damping: Dampens the vibration to keep it from becoming a destructive force

Make sure you understand the consequences that resonance has on your equipment. Not understanding and addressing equipment resonance will lead to unwanted and costly reliability issues.
 

by Trent Phillips

WATER/WASTE PROCESSING • December 2012

Production plant analyzes resonance anomaly; looks at condition monitoring program as a profit center

Sometimes in industry,  mechanical “circumstances” change. When it happens,  a machine train identical to other machine trains can suddenly become atypical. This was exactly the case for Process Water Supply Pump A, whose behavior was very similar to that of its sister pump trains, until something changed. In this article we discuss a problem that was abruptly encountered, the methods used to investigate it and the solution devised.
One of four identical pump trains mounted to a common piping system experienced a catastrophic motor (75 hp, 4 pole) failure. The motor could not be saved, and a new motor was purchased and installed. After installation, the pump was started with the new motor. High vibration caused the installers to immediately shut it down. The new motor had been laser aligned to the pump; therefore the alignment was not suspect; therefore vibration data was taken.
Read entire article “Where is that vibration coming from”.
Thanks to Roger Earley with LUBRIZOL for sharing this case study with us.

by Ana Maria Delgado, CRL

Thank you for joining us for our Webinar The Field Balancing Mine Field by Greg Lee. We hope you found the presentation to be valuable and very informative. If you missed our Webinar,  you can view the recorded version at any time. Watch now!
Here are Greg’s answers to your questions:
Q: Do you have experience balancing cooling towers?
A: Yes.  Cooling towers are interesting because there are a number of causes for vibration.  One very dangerous condition that can look like unbalance is a cracked hub.  This can lead to a catastrophic failure of the hub,  allowing the blades to break free and wreak havoc on anything near.  I once saw the result of a hub failure that caused the gearbox to break through the wood mounting frame and fall into the water tank.  The motor was still running with a 12-inch piece of jack shaft flailing around.
With cooling towers, it is especially important to run a complete vibration analysis before attempting to balance.  The customer in the example above had another cooling tower cell with the same cracked hub problem.  We caught that one before the failure using the VIBXPERT II vibration analyzer.  It was showing a high 1× radial vibration as one would expect from unbalance.  In addition it was showing a high 1× axially, as large as the radial.  The spectrums also showed high 5×, 20× and 25× frequencies as the blades bobbed up and down as they passed the 4 main gearbox supports and the jack shaft.  This is derived from 5 blades times 4 supports for 20× and 5 blades times 5 (4 supports and 1 jack shaft) for 25×.
Q: What about a multiplane, multipickup balance? i.e. Nuclear rotor train, 4 rotors, 8 bearings?
A: I am not sure what your exact question is, but, multiple rotors in a single train can be complex to balance.  If the cross effect from plane to plane is large, the complexity grows exponentially.  I worked with a field engineer that balanced a long train of 4 generators and 2 steam turbines using the trial weight field balancing method.  It took him a week to balance this system of rotors and bearings.
For something as complex and expensive as you describe, I would bring in the OEM or a company that specializes in Nuclear Turbine applications.  Because of the number of bearings, they would most likely use a 16- or 32-channel dynamic data recorder versus trying to use a typical two-channel field balancer.
Q: To select a trial weight, is there a ratio between the machine weight to trial weight to get the correct change in phase or amplitude?
A: Many of the companies that produce field balancing equipment have developed proprietary formulas to calculate how much trial weight to use and where to place the weight.  The intent of these formulas is to obtain a 30% change in amplitude and/or a 30% change in phase angle.  It should be understood that these derived trial weights are guides, not an absolute.  In most cases, these formulas take into account rotor weight, speed and the amount of initial unbalance.  The instrument then calculates the suggested trial weight and its position.
There are a number of “Trial Weight Formulas” used.  For example the United States Department of the Interior Bureau of Reclamation recommends that the “trial weight should be approximately equal to the weight of the rotating parts divided by 10,000.” Most of the time in field balancing the weight of the rotor is not known or is at best a rough guess.  In these cases it is advisable to look at the correction weights previously placed on the rotor and use these as a guide.
Q: How do you determine if it is hydraulic imbalance instead of mechanical?
A: By hydraulic unbalance, I take it that you are referring to internal hydraulic forces in a pump.  Fans can experience similar interference from wind.  Unbalance will manifest itself at exactly 1 times the running speed.  The unbalance vibration amplitude will be exhibited primarily in the radial direction.  If you see a lot of axial vibration (50% or more of the radial) then you likely have additional problems that are not balance related.
For hydraulic problems, look for an additional frequency equal to the number of vanes times the running speed.  Hydraulic instability in a pump is often seen in spectrums as low frequency broad-banded vibration below the running speed.  Often, hydraulic problems are accompanied by cavitation.  There are specific Shock Pulse measurements which will help you identify cavitation.
Q: On a balance stand, ideally you would want to be at running speed. Most of the time this is not possible due to size/mass etc.  How much difference does it make if you can only run at slower speeds such as 30% of operating RPM?
A: First, it is important to understand that balanced is balanced at any speed.  For an object to be balanced, the rotational centerline and the mass centerline must be the same.  This will hold true at any speed.  Because of this, it is not necessarily true that the best stand balance is at running speed.  For clarification, please refer to the first few slides of the presentation.
Now, to the answer:  In general, as long as one is away from the rotor’s critical resonance speeds, it is fine to balance a rotor at a speed lower than running speed.  No percentage rule is necessary. Just stay away from critical speeds.  With that said, there are some differences between balancing machines.  There are two primary types of balancing machines; a Soft Bearing Machine and a Hard Bearing Machine.  Both types have advantages and disadvantages.

  • Soft Bearing Balancer
    A soft bearing balancer allows the rotor ends to move freely in the horizontal direction in the balancing stand.  This type of balancer allows the rotor to turn at much slower speeds than the rotor’s operational speed.  The balancing procedure is almost identical to field balancing and a calibration or trial weight is used to test the response of the rotor.  In this way each rotor balance is in essence self-calibrated.  Like field balancing, multiple runs are required and the correction and trim weights are applied until the rotor meets the acceptable criteria.  As long as the speed is above the resonance of the soft work supports, and not at the rotor’s critical speed, the response will be linear and very accurate. Some of the largest steam turbines in the country have been balanced using soft bearing work supports resting on rail road tracks.  These rotors are balanced at speeds around 30 RPM.  If one is concerned about the number of runs in a stand, then a hard bearing machine might be preferred.
  • Hard Bearing Balancer
    A hard bearing machine fixes the rotor ends to the balancing stand pedestals.  This system only requires one run to determine unbalance and correction weights.  A hard bearing stand measures force, rather than motion like the soft bearing machine.  If one knows the force and angle of the unbalance plus the weight of the rotor, a correction can be calculated.  The advantage is that only one run is required to determine correction weights.  However, because the hard bearing machine measures force directly, the accuracy is sensitive to speed. If the speed of a rotor doubles, the force increases by a factor of 4.  Thus the higher the speed, the higher the measurable force and the better the accuracy of the balancing stand.  One may be nervous about running rotors such as fans at higher speeds due to wind forces.  In this case, a soft bearing machine would be better.

Q: What do you suggest if site balancing requires disassembling the pumps to get access to the impeller, Isn’t it worth doing in a balancing machine in the workshop?
A: Of course this depends on a lot of factors.  If one has to disassemble the pump to add or remove weight, it is probably preferable to remove the pump rotor impeller assembly to a balancing machine.
Q: For which machine sizes is site balancing more effective—small and medium machines or heavy duty machines?
A: In general terms, the larger a machine, the more expensive and difficult it is to move.  Thus the strongest case for field balancing is for larger machines.  However, machines like fans can be quite small and easy to access.  Field balancing is not limited to large expensive machinery.  It really depends on the application and the access to insert correction weights.
Q: Is site balancing a valuable condition to ask for during engineering and procurement stages?
A: I would recommend that any piece of new or used equipment being purchased have vibration and unbalance limits included in the specification.  I would refer you to the International Standards Organization (ISO) balancing and vibration standards for an internationally recognized reference for vibration standards.  If you are in the petrochemical industry, I would recommend looking at the American Petroleum Institute (API) specifications for vibration.
Q: Isn’t a coast down needed to find out if the machine is operating above or below critical speed to get a correct balance solution?
A: Yes. It is highly recommended that one identify the resonances of a machine before attempting to balance.  Field balancing at or near the critical speed can cause issues with amplitude and phase measurements.  As a general rule, one should stay approximately 20% away from a shaft resonance when balancing.  Because field balancing is basically a vector ratio problem, the field balancing technique will work fine for rotors running above or below the critical.
Capturing phase and amplitude during coast-down or startup is one of the best ways to identify the resonant frequencies of a rotor.  In the majority of situations, it is preferable to capture a coast down, as the data will not be influenced by the motor torque like it is during a startup.  With phase and amplitude data, one can view Bode and Nyquist plots which graphically identify the resonant frequencies.
Another method is to capture a cascade plot of spectrums as the machine starts up or coasts down.  Once again, this provides a particularly graphic method to identify a rotor’s resonant frequency.
Finally, perhaps the most common method of identifying resonance is a bump test.  This method can be used while the machine is off.  If your analyzer supports negative averaging, one can perform a bump test on a running machine.  The result shows a frequency spectrum where the peaks represent the resonant frequencies of the object being bump tested.
Q: For what size, speed, and HP machine would you recommend the installation of an external balance disc on the rotor to make field balancing and adding weights easier?
A: If it is the type of machine that would go out of balance often, is expensive to remove from service, does not have an easy way to add or remove weight, or is difficult to move to a balancing stand, I would recommend installing balancing disks.
Q: Have you experienced balancing long shafts where maybe 2 planes are not enough?
A: Absolutely.  If a shaft is long and flexible, additional planes may be necessary.  There is no hard and fast rule that states if a shaft is 10 times longer than its diameter, additional planes will be required.  Often, shafts will be supported by more than 2 bearings.  This would generally lead one to balance in more than 2 planes.
Q: I have heard that vibration due to misalignment conditions can be minimized through balancing but that seemed contrary to a remark made during the presentation.  Can balancing be effective in reducing machine response due to misalignment?  Thanks.
A: The first field job I did was for balancing a high speed, direct drive fan.  When I got there, an analysis revealed that there was a high amount of fan unbalance, a large amount of misalignment, and a very loose cork base.  The unbalance contributed to the looseness and the looseness caused the base to flex and all of these contributed to the misalignment.  The looseness contributed to the unit’s ability to vibrate at 1× the unbalance frequency and flex in the frame allowed additional misalignment.  The misalignment also contributed to the base looseness and the amplitude of the unbalance.  Any machine is a system, and, in this case, each condition made the other conditions worse.  They fed each other, but each condition must be corrected to fix the machine as a unit.  For example, if the looseness was corrected first, it would have zero effect on the balance.  By clamping down the base, more of the unbalance force is transferred to the bearings.  If the imbalance is left uncorrected, the bearings will fail early.  The unit still needs to be balanced and balancing will not correct the looseness or misalignment.  Since the balance is in essence a forcing frequency, the looseness may go down in amplitude but the machine is still loose.
In this case, the first problem to fix was the cork base.  It was removed and the fan grouted in, thereby eliminating the looseness.  If this were all we did, we would still have unbalance and misalignment.  So next we aligned the motor to the fan shaft.  Once this was done, the unit was started and the fan balanced.
In a pinch, we could have balanced the fan first.  It is likely that the looseness and misalignment would have been reduced, but would still have been present and feeding each other.  So I would say that balancing might reduce the symptoms of misalignment but not correct the misalignment.  The inverse would be true for correcting the misalignment.
Q: What is the difference between field balancing and using a balancing stand in a motor shop?
A: If the balancing stand is a soft bearing type, very little.  The process and math are the same.  In the field, there is less response linearly in the structures when compared to a soft bearing stand.  Thus, in the field, you can expect to see a little less unbalance reduction when placing correction weights.  This effect in the field is minimal.
If the balancing stand is a hard bearing type, then the shop process is a little different.  A hard bearing system measures force directly.  Knowing the weight of the rotor, the RPM and the force of unbalance, one can calculate the correction.  The results are nearly identical to a soft bearing machine.
Q: We have problems in balancing fans at full operating speed due to operational factors. What percentage of operating RPM should we try to balance at and what problems could we look for not balancing at full operating speed?
A: There is no specific percentage of running speed that will yield better results.  If you can reduce speed, then make sure you are not near a resonance where phase angle and amplitude shift.  A Bode or Nyquist plot taken during a startup or coast down is best for identifying the resonant frequencies of the fan.  Refer to the first few slides of the presentation. When the mass and rotating centerline are the same the rotor is balanced regardless of the RPM.
It is also important to make sure your fan is truly out of balance.  On a belt driven fan, check for sheave eccentricity where the sheave is off center or out-of-round.  This can cause vibration that looks like unbalance.  For instance, look for other influences such as air turbulence, unequal blade pitch, and looseness, to mention just a few.
Q: How do you calculate system lag? And will it change based on RPM?
A: By system lag I assume that you are referring to the balancing system.  In the old days, when we commonly used strobe lights to determine phase angle, there was significant lag in the electronics.  By knowing this, we were able to shortcut the balancing procedure and determine the heavy spot of a rotor.  Often this lag was about 40 degrees between when the heavy spot passed the transducer and the strobe triggered.  With the digital equipment we use today, electronic lag is  virtually eliminated.  For example,  I was balancing a spindle turning 40,000 RPM and was seeing less than 5 degrees of lag.
The easiest way to determine your instrument’s lag is to get a rotor that is balanced, place a weight at a known position, and see your instrument’s result.  By using the field balancing procedure built into today’s modern balancing instruments, lag is automatically compensated for in the balancing procedure.
Q: How to distinguish couple unbalance and quasi static unbalance?
A: Look at the phase angle of each plane.  If they are the same, it is purely static.  If they are 180 degrees opposite from each other, it is pure uncouple unbalance.
Q: What is the maximum level of vibration at which we can perform in-situ Balancing?
A: There is no set amount of unbalance where we cannot perform a field balance.  Of course one must apply a little logic.  If the vibration is so bad that it is causing damage, then it might be wiser to pull the rotor and place it on a balancing stand.
Q: Is there any procedure to perform single-shot balancing rather than 4-run method?
A: For a single plane balance, it requires 2 runs to secure a solution and an additional run to verify the result.  With a two-plane balance it takes 3 runs to secure a solution and a 4th run to verify the result.  Normally, in the field this is the best approach.
On a journal bearing machine where the masses are known and the heavy spot is verified, one can calculate how much weight is needed to reduce the vibration.  This would only take an initial measurement to determine.  However it is rare that we know the precise weight at each bearing, and even this process often takes multiple runs.
Q: Must the 30 -30 rule be followed for on-site balancing?
A: In the words of Captain Barbosa “…the code is more what you’d call guidelines than actual rules.”  The 30-30 rule is under ideal conditions.  I have balanced where I got the phase exactly right so the trial weight change was more like 0 degrees and 15% amplitude change.
Q: If we change angle of blades of cooling towers, will it have any effect on balanced impeller (During balancing let’s say we have 11 degree angle of blade, and then we have to change angle to 7 degrees because of process requirement)?
A: If one blade’s pitch is off relative to the other blades, it will look like unbalance.  However, this condition would be accompanied by a lot of axial force at a frequency of 1× because of the unequal blade pitch.  If it were pure unbalance, then the axial force would be steady and not have a large 1× frequency component.  So, in your example, as long as the blades pitch the same amount and the aerodynamic lift changes equally on all blades, the rotor will still be in balance.
Q: Do quasi static unbalance and couple unbalance have the same solution or something else?
A: You seem to be mixing terms.  When using a two plane balancing technique, the program takes into account the cross effect between planes A and B.  Separating the static and couple balance is possible, but with the accuracy of today’s instruments, it is rarely done.
Q: Is there any effect, if we put the trial weight at 75% RPM and then correction weight at 85% RPM?
A: The context of your question is not clear to me.  I think you are saying, if the speed changes during the initial measurement, trial weight measurement and the correction measurement, will this have an adverse effect.  Yes, changing speeds can create problems with the field balancing vector solution.  Sometimes it is impossible to take measurements at the same speed, and, the more this speed varies, the less accurate the balancing solution will be.
Q: Could you please go over the “no phase balancing” variation that you talked about in more detail?
A: To review the process would require an article.  This process is primarily used on single plane problems.  There are 4 steps required to calculate a solution.

  1. Take an initial amplitude measurement.
  2. Place a known amount of weight at zero degrees and take a second amplitude reading.
  3. Remove that weight and place it at 120 degrees.  Take the third amplitude reading.
  4. Remove that weight and place it at 270 degrees.  Take the third amplitude reading.
  5. This data can be plotted on polar paper to determine a solution.

There is more information on this process available on the Internet.
Q: Rather than try balancing at near critical resonance speed, would it be beneficial to try and stiffen structure to move resonance away from balancing and operating speed?
A: Yes this can be, and is done.  Many times it takes a lot of stiffening or adding mass to significantly shift a resonance.  Changing speed is easier, if possible.
I once balanced a large vertical fan in a 40 foot high tower.  The tower was resonating and causing problems.  We loosened the guidewires going to the top of the structure to decrease the stiffness.  This lowered the resonant frequency and helped us achieve a good balance.
I hope these answers were beneficial to all of you. If you have any additional questions, please feel free to contact Greg Lee, the presenter, directly.

by Yolanda Lopez