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6:00 am
April 1, 2009
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Tracing Fan Vibration to Flexible Soil

In this case, the cause of machinery health problems really did start from the ground up.

The call came in to Mechanical Solutions, Inc. (MSI) from an electric power gen company. Excessive vibration was plaguing two newly installed Induced Draft (ID) fans at one of the company’s flagship generating stations. The two units, each driven by an electric induction motor rated well in excess of 5000 HP, were operated in parallel at a constant speed with dampers that controlled the downstream flow of air. Elevated ID fan vibration levels at one times (1x) the running speed had been reported by the station. In fact, this vibration had been substantial enough to exceed the fan trip levels during startups time and again. Unfortunately, traditional vibration analysis techniques had not been successful in helping to resolve what initially appeared to be a straightforward machinery unbalance and/or misalignment problem.

Testing
To get to the bottom of the fan vibration issue, MSI performed a combination of operating forced response testing and impact modal testing. The purpose was to collect vibration data at locations throughout the fan systems, including the fan bearing housings, bearing supports, bearing sole plates, concrete pedestals and concrete floor pads.
fan-vibration_msi

The impact modal testing was conducted—safely—while the ID fans were running, so as to determine the natural frequencies and the mode shapes while the bearings were energized— which accounted for the important sleeve bearing stiffnesses of the units. An instrumented impact hammer was used to excite each bearing housing in the axial, vertical and horizontal directions, while a set of accelerometers and two multiple-channel spectrum analyzers recorded the response data throughout each fan system. The amount of force put on the fan bearings, supports and foundations at each frequency was transmitted by the piezoelectric crystal contained within the head of the impulse hammer. This input force was divided into each of the acceleration responses to determine a frequency response function (FRF) between the locations/directions of the hammer impacts and the locations/directions of the responses at the accelerometers. The logarithm of each FRF was plotted versus the frequency—which allowed both the low and the high response modes to be inspected with equal clarity. The peaks in the FRF plots represented the natural frequencies of the fan-pedestal-floor structural systems. The impact modal testing also was used to determine the mode shapes of the fan systems at each natural frequency of vibration. Data for each of the impact modal tests was acquired at approximately 50 locations in the three orthogonal directions on the bearing housings, bearing supports, bearing sole plates, concrete pedestals and concrete floor pads.

 

A specialized operating forced response vibration technique also was utilized to record test data in three orthogonal directions throughout the fan systems. This allowed the data to be subsequently processed to produce detailed, animated operating deflection shapes (ODS) of the fan systems. The data was collected under maximum load conditions, where the maximum forced response was present in the fans. Each mode shape and ODS animation displayed the relative motion, i.e. the amplitude and the phase, at each measurement location on the structure at a selected frequency. The animations were beneficial because they illustrated the relative motions of the various system components in an exaggerated fashion, which encouraged the efficient identification of the root cause(s) of the vibration problem. Still images from the ODS animations for the two ID fans are presented in fig. 1.

Diagnosis
The collected test data confirmed both of the ID fans had excessive vibration amplitudes that occurred in the horizontal direction at the locations of the outboard bearing housings and outboard support pedestals. The vibration spectra showed several harmonics of the fan running speed, and the highest peaks occurred at 1x and 2x the running speed. The ODS test results displayed horizontal rocking motions of the inboard and the outboard pedestals in both of the ID fan installations (see fig. 1). There also was clear evidence of looseness of the bearing assemblies in both of the fans (e.g. the housing, support and soleplate, especially in the outboard bearing assemblies).

The ODS data for ID Fan A showed in-phase horizontal rocking motion that was driven by the rotor motion at both the inboard and outboard support pedestals. flexibility of the base mat or soil underneath the pedestals allowed the pedestals to develop this side-to-side rigid body motion.

The test data for ID Fan B illustrated a more pronounced rigid body side-to-side or “rocking” motion at the outboard bearing pedestal, because of 1x running speed excitation driven by the fan’s rotor. Again, flexibility of the base underneath the pedestals allowed this horizontal motion to occur. ID Fan B’s rotor also described a relatively large horizontal orbit in the outboard bearing due to the motion of the bearing housing. Though the relative displacement of the shaft in the outboard bearing was relatively small and within operating specifications, the absolute motion relative to ground was several times greater. ID Fan B showed approximately 20% more deflection at the outboard pedestal in the horizontal direction than did ID Fan A, and 50% less deflection at the inboard pedestal in the same direction. Based on this information, MSI concluded that the rotor critical speed had shifted downward toward the running speed due to the high flexibility of the outboard support pedestal. This shift caused the fan to operate at resonance with the running speed.

Further, an independent structural natural frequency of the outboard pedestal was identified near the running speed. The rotor critical speed and the structural resonance interacted with each other, and were the likely cause of the modulated orbits of ID Fan B’s rotor. Rotordynamic analysis showed that the lateral stiffening of the outboard pedestals would detune the 1x resonance sufficiently to decrease the amplitude of the vibration responses to acceptable levels.

Solution
Ultimately, the root cause of the excessive vibration was the flexibility of the soil that was beneath the base mats of the supporting pedestals of each ID fan. It was recommended that the horizontal stiffness of the outboard bearing pedestal and support assemblies—e.g. the base mats and the underlying soil—be increased substantially. This modification would de-tune the offending resonance condition of the rotor critical speed and the pedestal structural natural frequency to create sufficient margin versus the fan running speed. In addition, it was suggested that all of the bearing housing and support assemblies be tightened and stiffened as much as practical to minimize the overall vibration of the fans. It was also noted that efforts to reduce the vibration levels by improving the balance or the alignment of the units beyond the manufacturer’s recommended parameters would at best provide a limited and temporary solution to the problem. Exceptional balance and alignment levels generally cannot be maintained in plant rotating machinery on a practical basis—especially in cases when the rotor is exposed to fly ash that will accumulate easily during the routine operation of the machine. The root cause of the vibration problem, soil flexibility, would have been extremely difficult and very costly to trace without the benefit of the specialized and well-proven troubleshooting approach that was implemented in this case. On the other hand, a design audit by a qualified firm—before the ID fans were installed at the facility—could have been exploited to avoid this puzzling vibration problem altogether.

Maki Onari is manager of Turbomachinery Testing and Eric Olson is director of marketing for Mechanical Solutions, Inc. (MSI), headquartered in Whippany, NJ. MSI provides consulting and R&D services such machinery design, analysis and testing on a wide range of equipment, including electronic systems and all types of rotating, reciprocating and turbomachinery, among others, for end-users and OEM clients around the world. Telephone: (973) 326-9920.

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