A Case for More Power…………

On several occasions over the last 10 years I have had customers with fan drive motor vibration issues that looked like a classic case of unbalanced. Instead we found ‘Lack of Horsepower’ as the cause. And all were powered through a Variable Frequency Drives (VFD). What I’m going to discuss in this column isn’t a common occurrence but it’s one that should be noted and remembered. You could be a maintenance hero if you recognize and correctly identify this issue; it’s all about Horsepower, or the lack thereof. And one more quick point before we get started, all discussions below are based on a VFD being programed to accommodate Pumps and Fans; Volts per Hz. Why? Because pumps and fans are centrifugal and torque requirements are exponentially related to loading, load here is not linier; maybe. VFD’s can also be programed for a constant torque mode. However, this mode is generally reserved for conveyors, compressors, anything with a high starting load, anything that is positive displacement, etc… If we were to use constant torque in this application, we would lose those efficiency gains that had us install the device to begin with. It’s all about profitability; remember.

To start out our conversation, when I buy an induction motor, it is rated to produce a specific Horsepower (HP) at a specific speed and at a specific Full Load Current.

When I introduce a VFD and change the motor’s speed, I also change the HP and what the motor sees as far as maximum current. And the question I have is; “why do people forget this?”

Fan Details – On fans where I have seen this issue, all were belt driven (AMCA Arrangement 9), backward incline or radial blade centrifugal fans. And on the backward incline fans, the blade angle was fairly steep, approaching that of a simple radial blade (90°). And if one knows his fans, you should remember that several fan manufacturers (Twin City predominantly) at one time published a general caution on radial fans; ‘under certain conditions, the fan wheel break HP can exceed the motor HP’. WOW!!!! Can this be true? Yes it is true and the condition they described in their caution was at maximum flow, not minimum. Our VFD related issue is at the other end of the spectrum, low to medium flows. Say what?

Radial fans generally have a pretty steep fan curve; they produce a constant flow over a fairly large pressure range. As we increase the blade angle, providing more and more of a backward angle (trailing blade angle; from 90° toward 120°) we lengthen the fan curve to provide a greater flow variation over the same pressure range. Just by looking at the typical fan curves it is pretty easy to see that it takes a pretty good pressure swing to affect the flow through a radial fan. Whereas a backward incline fan, modulating pressure gives me a pretty good variation in flow.

In HVAC systems, we have two mechanisms that allow us to adjust flow; speed modulation and pressure modulation. Pretty simple, right? Maybe.

When we adjust fan speed, to determine changes in flow, we reference our affinity laws;

·         speed changes cause proportional changes in flow,

·         and changes in speed are proportional to HP cubed.

When we adjust fan pressure, to determine changes in flow, we need to reference our fan curve and our affinity laws;

·         changes in fan pressure are proportional to HP squared,

·         I need a specific fan curve to relate the pressure changes to flow changes,

·         and changes in flow are proportional to HP cubed.

In them olden days, before the advent of Pulse Width Modulated Variable Frequency Drives (VFD, cheap speed control), we had fixed speed fans where we had to open and shut dampers to change flow, e.g. we modulated system pressure. Basically, when we modulate pressure across a fixed speed fan, we are changing the system curve; we are changing total pressure across the fan. To lower the flow we increase total pressure across the fan (shut a damper) and to increase flow we decrease total pressure across the fan (open a damper). The problem with this method, we lose efficiency the more we close or the more we open the control damper. When we use fan speed (VFD) to modulate flow, we are actually sliding the fan curve up and down the system curve. This gives us a greater range of flow control and we are managing efficiency by keeping the fan at its best efficiency point, most of the time anyway.

When we procure a motor for our fan, it doesn’t matter if we plan to use speed or pressure modulation; a motor is always selected for the application with enough HP to handle the maximum expected (designed) flow. With a pressure modulated system, I don’t worry about total pressure causing a motor overload condition when reducing flow because of the difference between HP cubed as compared to HP squared; we can always make pressure as long as we sized our motor based on expected system flow. However, if I try to flow more through the system (increase flow) than what the motor can handle, we enter a condition of motor overload, not enough HP to push air through the system. And believe me, that motor will do everything it can to push air through the system. It’s going to maintain speed or die trying. This is true for both speed and pressure modulated systems.

In a pressure modulated system, when I reduced flow, when I shut the damper, I have excess HP because there was no speed change for the lower flows; less work, less power, no overload. And when I open the damper, I am loading the fan (need more power) because of the additional flow; more work, more power, more potential for overload. Pressure modulation is not very efficient but it is effective. Kind of like a sledge hammer in a jewelry shop.

Another concept we need to introduce at this point is ‘Induction Motor Slip’. Induction motors (simple squirrel cage motors) power 95% of industrial equipment, especially fans. For an induction motor to work there needs to be a speed differential between the field rotation and the actual rotor rotation. This speed differential is called motor slip. As we load the motor and demand more HP, e.g. more torque, motor slip increases. A good running motor operating at its Name Plate Rating will operate anywhere between 1 to 5% slip. And even if we install a VFD to control motor speed there is still motor slip or the motor would not run. If we increase motor load, increasing motor slip past 5%, we begin to overload our motor; windings heat up, the motor runs rough, we get high torque pulses as the rotor tries to catch the rotating field. We see this, especially the torque pulses, as a high vibration with all of the energy at the motors fundamental frequency, the X1. On a fixed speed installation, with a simple across the line motor starter, there are thermal overloads that protect the motor from this condition. On a VFD, the overload setting is a programed set point, based on the motor’s full load (and speed) Name Plate Rating. But, does this programed overload set point protect the motor when running at slower speeds? Maybe not!!!

Back to the fans - In reality, radial fans will give me a flat HP curve (not a lot of change in flow) and in backward incline fans, we will see more of a vertical HP curve (lots of flow differential). Remember, there is a relationship between speed, flow and HP.

So what happens to motor power (HP) when I slow a motor down, especially with a VFD programed for Volts per Hz? Speed is lost, voltage is reduced, motor current is limited and HP is lost. As long as flow is decreased proportionally, there shouldn’t be an issue. Unless our fan characteristics; e.g. the HP require to drive the fan at various flows, is FLAT. Then we have a condition of motor overload. Why? Because with the motor operating at a lower speed and with limited current (due to the VFD) we can only produce so much HP. And if flow has not been reduced enough to match motor HP, we are in a condition of motor overload. And how can I tell? Motor slip has increased and if the slip is significant, we will see a high motor vibration. But what about my motor overload protection programed into the VFD, why doesn’t that trip the motor? The reason; when I program the overload set point based on motor full load performance (Name Plate Rating) and I have an overload condition at lower speeds, with limited voltage and current, I never reach the overload set point. Electrically the motor is perceived as ‘all good’ except it’s beating itself to death because of excessive slip. Oh My!!!!!

And no, unfortunately the ‘Slip Compensation’ function included as a programing feature of most VFD’s does not correct this issue. This feature does not work for process control (pressure or flow control), it was intended for process ‘speed’ control (conveyers, etc…).  

And the fix; the easiest way correct this problem is to throttle down on a damper. Yes that’s correct; introduce some back pressure into the system. This forces the fan to run faster at all flows and the faster the speed, the more torque and HP is introduced. As an alternative, since the problem is at the bottom end of fan performance and if you never operate in this region, just add a note to your SOS that cautions the users about this condition and limit operation below certain speeds.

So why not just reprogram the VFD to Constant Torque mode? Isn’t this the correct setting for this type of issue? It could be the answer if I were to see excessive starting current and excessive load through a larger portion of the RPM rang. Since this issue is at the bottom end of system performance, under normal conditions, as filters load and the original facility HVAC balance is maintained, we probably wouldn’t notice the issue. So there is no reason to set the VFD up outside of the manufacturers recommendations; fans and pumps = volts per Hz, conveyors and compressors = constant torque. Following the manufacturer’s recommendations gives me the best return on my VFD investment. But if I change filters (a clean system) or maybe change HVAC configuration without checking the effect on my fan, the issue could sneak up and bite us in the buttocks. So how would you check for this condition? I use a strobe tachometer to check shaft speed relative to what the synchronous frequency should be. My acceptance here; Motor Slip shall not to exceed 5%.

And because you actually read this little white paper, you can break out your strobe tachometer to immediately identify excessive motor slip and then suggest closing down on a damper to correct the vibration issue; you get to be the Maintenance Hero. Gosh, fixing a X1 vibration indication without balancing the rotor; WOW!!! 

Below are a couple of ‘Tales From the Crypt’ that show examples of this uncommon occurrence and more importantly what we did to correct the issue.

·         Story 1 – 75 HP Belt Driven Overhung Fans, Richland Washington

At the Hanford nuclear reservation 242A Evaporator, the main exhaust ventilation system was being replaced by two new exhaust trains that included a pair of 75HP overhung fans. My mission was to provide vibration acceptance of the fans during commissioning. In this circumstance, the new fans and filtration systems were installed but not yet tied to the facility. This allowed testing without worrying about radiological contamination. Good idea; we thought!!!

The commissioning procedure had included closing down on the filter intakes to simulate HVAC system losses as if the filter trains were tied to the building. However, it was a missed step. When I first saw the fans operate, the VFD’s were set at 40 Hz or about 1200 motor RPM. A quick round of vibration data on the fans and motors showed no issues at 40 Hz. So we began testing. The test called out collecting vibration data on the motors and the fans from 20 Hz through 60 Hz from the VFD. Fan vibration acceptance was in accordance with AMCA 204 and since the AMCA does not cover motors, we used general criteria from the ‘outdated’ ISO 2372 standard for Class II equipment, not to exceed .18 IPS RMS; a solid C. We basically stated in our acceptance documentation that we knew the ISO was outdated. However, the .18 is representative of the other similar equipment at the facility. Remember, acceptance is not too high as to pose a safety concern and not too low as to be operationally limiting.

So we lower fan speed to 20 Hz and after flow has steadied out, the motor started bucking. I was getting .5 to .7 IPS RMS on the motor yet maybe .3 to .35 on the fan bearings. Then as we ramped up the speed, things got better (could meet the criteria) around 40 Hz. The spectrum analyzer I was using showed all of the vibration energy at X1. So what was the problem?

We secured from the testing and regrouped over several meetings. We sent the crews out to check simple items; nuts and bolts, belt alignment and VFD condition. Everything checked out. I even went back and completed a simple bump test to verify there was no resonance condition. And balance, bah humbug, since the motor smoothed out as RPM increased, there was no issue. Then, as I walked by the filter inlets, I noted that the dampers were wide open. And I thought to myself, “didn’t we have a step to throttle down on these to simulate building losses”, why were they open? And was this at least part of the reason for the high vibration?

So we repeated the test, this time with the dampers partially closed and gee whiz buckaroo, all of the vibration levels at all measurement points came within the criteria.

So I went back and reviewed the fan curves and got with Allen Bradley on their VFD. What we found was with wide open dampers (low total pressure across the fan), below 30 Hz to the motor, the fan wheel break HP required was about 10% greater than what the motor could provide. WOW!!! And as we spooled up the fan, the motor finally caught up with the fan as far as the HP to drive it.

I used this anecdote to describe the condition; Basically, we were trying to accelerate our Ferrari in 6^th gear from 20 to 100 MPH. In 6^th gear, it bucks and jerks and stalls, I can’t make enough power to push it smoothly until I get to around 40, then the engine rev’s are high enough to make enough power to accelerate like a banshee.

When we completed final acceptance testing of the fans after the hot tie-in, we left provisions in our procedure to allow adjustment of the building balance dampers if the condition re-appeared. Luckily, there were enough losses from the building to keep total pressure across the fan high enough for everything to work well.

Lesson Learned; if I were to have included a check of motor slip at the bottom end of the RPM range, we would have caught this issue a heck of a lot sooner. Using a strobe tachometer and inspecting for this condition is now part of my ‘low hanging fruit’ checks. Mounting bolt tightness, electrical (balanced current/voltage between phases, VFD) checks, inspecting the structure (resilient mounts), cracks, precision alignment and  motor slip are cheap and easy and rule out the most common issues that cause vibration.

·         Story 2 – 50 HP Belt Driven Overhung Fans, Asan Korea

At a customer’s facility in Asan Korea, they have over 40 similar manufacturing lines, each with a very high temperature processes at the front end. Each pair of the lines has 3, 50 HP AMCA arrangement 4 fans that provided cooling (and backup) for their respective high temperature process. All of these fans are powered through VFD’s and generally run at 50 to 60% of the motor Name Plate speed. These systems had been installed and operational for approximately 10 years. My mission for this customer was providing PdM and CBM training.

During my initial classroom sessions, I generally ask many, many questions to get a feel for what the engineers and technicians know, or don’t know. Plus I have them bring whatever monitoring equipment they normally use for the collection of PdM data. One of the engineers showed up with an older SKF CMVA vibration analyzer. After the initial training session he shared with me a spectrum of a motor with vibration amplitudes substantially higher than similar equipment. It happened to be on one of the 50 HP cooling fans. From his data, he kept showing the high X1 peak and telling me the motor needed to be balanced; could I show them how to complete a trim balance using his analyzer.

In these cases I never drop an opportunity to include additional training; in the field on their equipment is premium ‘hands on’ time. However, I am always skeptical on why there is a need to balance equipment without an underlying reason. So I asked the engineer; Is this a new motor? No. Have you done any work on the motor? No. Have you checked the mounting nuts and bolts? No. Have you checked belt alignment? No. Have you verified the VFD is in good working order? Deer in the headlights, No. Do you have a strobe tachometer? Yes.

O.K., so after the inquisition, I hope you (the reader) know what I’m about to suggest; let’s take this fan out of service, get the backup on line and go after the low hangin’ fruit; check nuts and bolts, belts, alignment and electrical issues before we get into a balance. Checking a VFD for correct operation was new to them and actually worthy of yet another article to come (future writing). And last but not least in our new repertory of low hangin’ fruit checks; check the motor slip with a strobe tachometer.

So as we discuss the upcoming activities with their operations supervisor, and as we were discussing the VFD checks, the need for the strobe tachometer came up. The operations supervisor asked “why are we using a strobe tachometer?” My reply, “motor slip caused by an overloaded motor at lower speeds”. Then I went on to explain the motor speed, motor HP and how fan wheel break HP can exceed motor HP at reduced speeds. And when I stated damper positioning can cause the issue, it came up that on this particular fan, 2 years prior, the fan outlet damper actuator was removed and the damper had been locked wide open. It was originally only 80% open. Plus they had to slow the fan to establish the same flow (approximately 40% speed). And from the vibration technician, the high vibration on this unit was discovered soon after. WOW!!! Could it be? A low HP condition due to excessive flow? Yep, and the root cause; introduction of error!!!

As an exercise in how to address all vibration problems as they came up, we went through all of the basic checks, nuts and bolts, electrical, VFD, alignment, etc…, and what did we find? 7% slip on the motor at a motor speed of 720 RPM (40% speed) and all of the system dampers wide open, the motor was overloaded yet the current was not excessive, no overload trip. So then we started to throttle the fan inlet damper (vortex damper) and ramping up fan speed to maintain system flow. Once we got to about 25% closed on the inlet damper and made our final speed adjustments (final speed 1000 RPM, 55% speed), motor slip was around 2% and overall vibration decreased from 10 mm/s RMS to less than 2 mm/s RMS, a very significant improvement. WOW!!! Correcting a X1 indication and no balancing.    

Two lessons from these stories; motor slip on VFD’s at lower speed is a tell tail for motor loading issues and If I see excessive load combine with previous work and then a jump in a CBM  indication, you can be sure at least 99% of the time you have found the cause your issue.

Maintenance, what a Concept

MMJennings

No Spectrum Analyzer, Oh My!!!………….

In my last blog entry, I kind of left you hanging. Yes, there is a ‘cost’ benefit to starting out with an analog based vibration monitoring program but what do you lose when a digital device has the potential to do so much more. A digital vibration analyzer providing visual indication of both time domain (AC signal) and frequency domain (spectrum of contributing frequencies) is a valuable tool for any facility. And if you have a technician that can actually use the device to its full capability, then you have a winning combination. However, without an analyzer or more importantly, a competent technician, it can be very difficult to ‘troubleshoot’ vibration problems. Or is it? The sales guys will always insist you need the best, a super wamo spectrum analyzer, to monitor and troubleshoot, Or do you?

So let’s look at this problem from a common sense point of view. If I went into a facility cold, with no knowledge of previous work and collected simple analog Low Frequency vibration data (overalls, filter out RMS value), typically I would find 15 to 20% of similar equipment types on the high side of an established normal. And of that 15 to 20%, you may see up to 5% of those items that have any serious issues. And if I supplement the Low Frequency data collected with a round of High Frequency data, based on my experience, I would typically find less that 5% (closer to 3%) of similar equipment types with indications of potential bearing issues

And before I leave the discussion of common sense, I need to make one more point; most vibration issues that are of a serious nature, the ones that are destructive, the ones that are the reason for having a Shutdown List, are usually (better than a 100% probability) cause by design misapplication or installation error. WOW!!! These are the issues that are hard to troubleshoot and even harder to fix. However, once identified and corrected you should not have to deal with these again. These are generally issues that are a One Time Fix.

So let’s do the math on our initial vibration survey; if I had 60 small frame (ANSI frame) pumps, let’s say all between 60 and 100 HP, and 20% exhibit higher than normal values (Alarm status), than I have 12 pumps on my Watch List. 5% of those pumps, or one pump, would exhibit vibration readings that would warrant a Shutdown status. And, I would have an additional 2 pumps that indicate a potential bearing issue. In reality, of the 2 pumps identified with potential bearing issues, most likely one or both pumps would be in the group of 12 Watch List pumps. However, for this discussion, we’ll assume that there are 2 addition pumps that need attention. Final count; I have one pump out of 60 that requires immediate attention and I have 13 pumps that should be placed on a Watch List. Of the 13, 2 are bearing issue and 11 other issue.  

So, do I need to spend the money on a spectrum analyzer to troubleshoot on my one Shutdown pump? I’m not sure that I do. Why? Between the cost of the analyzer and the training to use it, I could buy and install a new pump, maybe even two. And, keep in mind there is a good chance that the vibration issue on my one Shutdown pump is a One Time Fix. What if I just called an expert, a consultant (a vendor) and had him look at my pump(s); that would be way more cost effective than a digital analyzer and the technical training. Plus if the vendor did not provide an outcome to my satisfaction, I don’t have to pay him. I have to pay my technicians, no matter if they correctly identify the issue or not. Remember, with a program, I now have performance criteria (acceptance criteria) and I can include that criteria in a contract and hold that vendor’s feet to the fire; e.g. if he advertises he’s an expert, then his accounts receivable is based his performance and of course, my equipment improvement.

The important part of the above message; you now have a program using cost effective tools to categorize equipment into good, mediocre and poor performers. And based on performance criteria that you as the equipment owner took the responsibility to develop, you can now use that criterion to establish vendor contract requirements. WOW!!! There we go again, running our Maintenance Department like a business.

On the two pumps with potential bearing problems; just replace the bearings, its cheap insurance. Do you really give a rat’s butt weather it’s an inner races or outer race defect? That’s what a spectrum analyzer could tell you. I think not; you just care that you have an identified bearing issue that needs attention. And now that I have established Low and High Frequency vibration performance criteria, I can use the criteria to ensure my rebuilds are effective through post maintenance acceptance testing, e.g. we are holding ourselves accountable (like the vendor) to do error free work. WOW!!! There we go again, running our Maintenance Department like a business.

So now that I have a vendor contracted to sort out my one Shutdown status pump and I am replacing bearings on my two pumps with potential bearing issues, what about the 11 others that are in an Alarm condition and that I have placed on a Watch List? How do I troubleshoot those without a spectrum analyzer? Statistical Probability baby! It’s all documented on the internet. You say what?

Common Causes of Vibration…………..

If you have been in this business long enough, especially over the last 15 years, you have probably read just about everything there is to read on the internet associated with vibration troubleshooting. If you haven’t, you’re a fool. The internet is rich with comments, videos and technical articles that focus on vibration monitoring from around the globe. However, I do concede, there is also a lot of trash out there.

Anyway, one item you will see pop up every now and then on the information super highway is a chart that shows the most common causes of equipment vibration. I have developed my own variation based on my experience in power production, DOE nuclear remediation projects and manufacturing segments. And unlike the other talking heads and their internet variations, my top cause is alignment, not balance. The following is what I present in my training courses.

Maintenance Concepts Table of Common Vibration Sources
Source	Probability of Occurrence
Misalignment (soft foot, coupling, belts and sheaves)	50%
Balance	25%
Resonance (At Run Speed. If VFD involved, will replace Balance as #2 Cause)	15%
Loose Parts or Components (Loose Nuts and Bolts)	5%
Electrical (voltage/current unbalance, air gap, overload, eccentricity. If VFD involved, will replace Loose Parts as #4 Cause)	3%
Bent Shaft, Shaft Dimensional Defects	1%
Blade Pass, Vane Pass, Gear Mesh or Bearing Defect AMPLIFIED by Resonance	1%
Rotor Rub	Less Than 1%
Leprechauns and Gremlins	Less Than 1%

So on the remaining 11 Watch List pumps; I will schedule things I have immediate control over; things I can generally self-perform, things that are in the top 10 as far as causes of vibration. Initially I want to check the easy stuff, the low hanging fruit. I will check mounting bolt tightness, electrical (balanced current/voltage between phases), inspect the structure (resilient mounts, cracks and breaks) and I will schedule precision alignment checks. Precision alignment checks? What the heck is precision alignment? The answer is nothing more than we what have been discussing from the beginning of this blog;

Precision alignment is establishing and documenting a measurable tolerance (acceptance criteria) for flexibly coupled equipment based on equipment speed. And most important, meeting that criteria when completing the alignment.

WOW!!! Not only do I now have Performance Criteria for my dynamic equipment (vibration monitoring), I have an Acceptance Criterion that allows me to meet that performance criteria. Again, the following table is what I present in my training courses for general coupling alignment.

Maintenance Concepts Table of Alignment Tolerances Small Frame Pumps Electrical Motor Driven
Shaft Speed (RPM)	Offset (mils)	Angular (mils/in.)
600	5	2.5
900	4	2
1200	3	1.5
1800	2	1
3600	1	0.5
> 4000	0.5	0.25

These tolerances are based on the Reverse Dial Indicator method of alignment. All commercial laser alignment sets on the market today are based on this method of alignment. If you don’t know or understand this alignment method, you need to research it; use the internet. Below are just a few links that look at different alignment methods, shows when they are applicable and provide instructions on how to complete each alignment method.

http://www.ludeca.com/brochures/Ludeca_A-Practical-Guide-to-Shaft-Alignment.pdf

http://www.lifetime-reliability.com/tutorials/maintenance-management/18_ShaftAlignment.pdf

http://reliabilityweb.com/index.php/articles/rim_face_alignment_method/

So what if your facility does not have the skills or equipment to perform this type of alignment? Based on your established and documented precision alignment acceptance criteria, that you as the equipment owner took the responsibility to develop, you can now use that criterion to establish vendor contract requirements. WOW!!! There we go again, running our Maintenance Department like a business.

And to double check the vendor’s work, besides just meeting alignment tolerance criteria, remember that I now have an established Low and High Frequency vibration performance criteria and I can use that criteria to ensure the vendor’s alignments are effective based on post maintenance acceptance testing. WOW!!! There we go again, running our Maintenance Department like a business.

Based on my 30 plus years in the industry, checking nuts and bolts, correcting any electrical issues (balanced current/voltage), verifying structural integrity and verifying alignment, better than 75% of your Watch List pumps will be brought into an acceptable dynamic (vibration) performance range.

And now we have just 2 remaining Watch List pumps. Here, the cause of vibration may just be a bit harder determined and correct. Based on our referenced probability, there is a good chance of a mechanical balancing issue(s). We will review mechanical balancing in an analog world in future blogs; hint you don’t need a spectrum analyzer here either. Anyway, now it’s really time to call an expert, a consultant (a vendor) so that he can provide support to your maintenance staff. Just remember, in his contract, establish criteria for him to work against. Make it clear he only gets paid if there is an improvement. WOW!!! There we go again, running our Maintenance Department like a business.

Using Reason, Logic, Common Sense and World Class Business Practices, leveraging access to global media as a tool to determine the most probable causes of dynamic equipment degraded performance, establishing specific performance criteria and establishing specific acceptance criteria, I can improved the dynamic condition of my 60 small frame pumps without investing in a spectrum analyzer and the associated training. And, I am left with the advantage of a ‘validated’ and ongoing Condition Based Maintenance (CBM) program at the best price.

Where is my Value Added…………

The last step is to document the improvement and show the value added by the improved performance of my 14 small frame pumps. Please don’t expect to see a complete payback for your monitoring equipment, training and program investment. However, showing small gains form this one specific equipment type can improve your credibility with CBM program investment. The easiest way to show value added for common equipment types, like pumps and fans, is to show deferred overhaul (rebuild) schedule; from a 3 year overhaul cycle to a 4 year overhaul cycle, showing the cost deferment over a 10 year period; e.g. instead of 3 overhauls in 10 years, you now have 2. And, if luck is on your side, one of your 14 deficient pumps will be a business critical pump. Here you could model the cost saving through an avoided unscheduled outage; e.g. my $5000 investment just saved you $1,000,000 in production losses. Again, I have several spreadsheet based models that I will share in future blogs.

Nuts and Bolts and Washers………..

Before we leave this article, I want to talk about fasteners that secure dynamic equipment to a foundation. I can include a myriad of horror stories here but I will stick to my three favorites.

·         Story 1 – 60 HP, Gould 3196 Pumps, Tainan, Taiwan

Part of our job scope in Taiwan was to assist the customer in establishing a basic vibration monitoring program. We set up a basic program similar to the one shared above; all analog. On our first round of data collection, I noticed that better than half of their small frame pumps were running a fairly steep vibration level; at the lower end of the scale we were seeing < 1 to 1.5 mm/s RMS at all monitoring points and at the higher end, we were seeing 2.5 to 3.5 mm/s. And there were several readings in excess of 4 mm/s. The facility set their Alarm value ant 2.5 and their Shutdown at 3 mm/s. So after we collected and then reviewed the data as part of the training, I asked the question; was there any work recently completed on the pumps that had higher readings? The answer was yes. During a recent outage, a vendor was hired to check and correct alignment on a group of pumps. Initially all of the pumps ran quite smoothly and then as time passed, they ran rougher and rougher. So I asked if they checked alignment. The answer; “no, the vendor had realigned the pumps during the outage, the pumps are in alignment”.

After I finally convinced then to check pump alignment, we started out with a simple soft foot check. Dial indication on the motor foot, loosen the mounting bolt and see if there is any lift. In this case no motor foot lift, however what really surprised me was how a 135 lb man (Taiwanese men are fairly small individuals) with very delicate hands and arms could loosen the motor mounting bolts with nothing more than a 10” crescent wrench. WOW!!! And when I ask if the bolts had ever been tightened with a torque wrench, I got a blank stare.

Folks I’m gonna’ tell you right here, if you don’t tighten the bolts at or near their yield, they’re going to loosen up over time. And that’s just what happened. The bolts were not tight and the motors shifted over time causing alignment issues and a corresponding higher vibration.

Moral of the story; always torque your mounting fasteners or at lease use a cheater to tighten the heck out of them. A 10“ crescent on a 5/8“ bolt just isn’t going to cut it.  

·         Story 2 - 300 HP,   Richland Washington

At the Hanford nuclear reservation, a cross site transfer system was installed to transfer highly radioactive contaminated waste over several miles from one storage area to another. The project dollar value was in the billions and completion was being held up by high vibration from a pair of 300 HP multi stage horizontal booster pumps manufactured by Sulzer Bingham. Just to relieve the suspense; the high vibration was not caused by an alignment issue, it was a resonance issue and of course a VFD was involved (my substitute #2 cause). However, when we went in to double check the common causes of vibration, the ones we have immediate control over (alignment, loose nuts and bolts, electrical and structural), we found that the washers that were part of the motor mounting were not rated for graded fasteners, e.g. they installed a SAE Grade 2 washers under SAE Grade 5 bolts. And as you torqued the bolts up, the washers would collapse under the bolt heads and would actually ‘extrude’ into the motor feet bolt holes.

We discovered the issue when checking bolt tightness; we torqued the bolts to specific criteria (a torque value for that size fastener) and as the washer pulled through the motor mounting holes, it would affect the alignment. Basically, we could ‘snug up’ the bolts and maintain alignment but when we went to torque the bolts after the final alignment adjustment, we would be outside of our alignment tolerance. The fix; hardened washers - ASTM F436. The hardened washers fixed this problem and allowed us to focus on the real issue, the resonance issue.

·         Story 3 – 125 HP Belt Driven Overhung Fan, Richland Washington

At the Hanford nuclear reservation, the Plutonium Finishing Plant main exhaust ventilation system consists of 8 electric driven 125 HP belt driven overhung fans and 2 back-up steam driven (approximately 75 HP) fans. These fans all take suction on a common plenum and exhaust through a common stack. These fans have been (and still are) in operation since 1943. Many still have the vintage 1943 open drip proof motors that were originally installed. However, this is not to say that these fans operate with no vibration issues. Because of the parallel operation and slightly different fan curves, these fans have a tendency to produce flow oscillations, a pressure disturbance, as if the fans are operating way to the left side of the curve. The result is a high axial vibration. This has taken its toll on these older fans, historically plagued by fan wheel cracks and bearing issues. In 2011, one of these fans failed catastrophically, the fan wheel hub shattered allowing the fan wheel to unwind itself in the housing, and actually shearing the fan shaft bearing bolts from the mounting pedestal allowing the fan shaft to crash into the motor.

Shearing the Fan Shaft Beating Bolts; these bolts are ¾” and the wheel failure sheared them flush with the bearing pedestal. WOW!!!! No matter what the fan wheel or shaft does, the bearings should never be ripped from the mount. The problem, over time with many, many bearing replacements, a spacer plate was added and welded to the bearing deck on the pedestal. Instead of drilling through the spacer and the bearing deck to relocate the new bearing bolt holes, the plates were drilled and tapped and the bolts were installed in blind holes, e.g. the spacer plate (a piece of A36 cold rolled, very soft material) was acting as the nut. And for whatever reason, over the years as new bolts were procured, they were ordered way too long. Instead of reordering the correct length or at least cutting the bolts down, washers were stacked under the bolts for a spacer so they could be tightened. And the bolts installed were 304 stainless steel, a very, very soft material. So between the soft bolting material, extended bolt length and limited thread engagement, when the bolts where stressed they failed and turned a wheel failure into a complete fan failure. This fan has been permanently removed from service.

It is hard to believe this type of ‘cob job’ came from a nuclear facility, but it did. They let their design safeguards (configuration management) degrade and it bit them in the buttock. If you are securing a dynamic component, use SAE Grade 5 fasteners or better. Use hardened flat washers and use lock washers.

The objective of these stories is to highlight the importance of fasteners and dynamic components. If any of the asset managers (responsible owner(s) of the equipment) had thought through failure prevention, they would have their component fastening systems spec’ed out and acceptance criteria documented and defined for torqueing the fasteners.

My recommendation for mounting dynamic equipment;

Maintenance Concepts Motor and Bearing Mounting Fastener Sets
Components	Specification
Bolts	(length, diameter), NC, SAE J449 Gr. 5, Zinc Electroplated IAW ASTM B633
Nuts	(diameter), NC, SAR J995, Gr. 5, Zinc Electroplated IAW ASTM B633
Flat Washers	(diameter), ASTM F436, ANSI B.18.22.1, Zinc Electroplated IAW ASTM B633
Lock Washer	(diameter), CS Lock Washer Wire, ANSI B.18.21.1, Zinc Electroplated IAW ASTM B633

Maintenance Concepts Basic Torqueing  Criteria
Fastener Diameter	Torque Value SAE Gr. 5
1/2”	48 ft-lbs
9/16”	70 ft-lbs
5/8”	96 ft-lbs
3/4”	160 ft-lbs
7/8”	240 ft-lbs
1”	370 ft-lbs

Maintenance Concepts Fastener Installation Criteria

For securing motors, motor mounting plates, mounted bearings or any item supporting a shaft where shaft velocities exceed 1500 FPM, as a minimum, SAE Gr. 5 fasteners sets shall be used. A fastener set includes; bolt, nut, 2 flat washers and 1 lock washer. When installing fastener sets, a minimum of 5 threads and a maximum of 10 threads will extend from the nut. Bind mountings will not be used. All fasteners sets shall be torqued after final adjustments.

For adjustable motor mounting plates generally associated with belt driven equipment, where mounting studs are provided as a component of the mounting plate, SAE Gr. 5 nuts, hardened flat washer (F436) and lock washers will be used to secure the motor to the plate. In lieu of a final torqueing, tightening the nuts to fully crush the lock washers plus ¼ turn is acceptable.

Maintenance, what a Concept!!!!

MMJennings