Saturday, 3 October 2026

Fans and Pumps: Blade Pass, Vane Pass, Cavitation and Flow-Related Vibration

From Mechanical Maintenance to Vibration Analysis - Part 9

In Part 8 - Resonance Diagnosis, we learned that a normal forcing frequency can become severe when it excites a natural frequency. Fans and pumps create important forcing frequencies of their own, especially blade-pass and vane-pass frequencies.

These machines also interact continuously with air, gas or liquid. A spectrum therefore reflects more than the rotor and bearings. It can contain evidence of inlet restrictions, poor operating point, recirculation, turbulence, cavitation, impeller or blade condition, and structural response.

Blade pass and vane pass are forcing frequencies, not automatic fault diagnoses. Their meaning comes from amplitude, harmonics, sidebands, waveform, operating point and process condition.

Begin with speed and component geometry

Before interpreting a peak, record the actual shaft speed and count the blades or vanes. Convert speed from revolutions per minute to hertz:

Running frequency (Hz) = rpm / 60

Then calculate the relevant pass frequency:

Blade-pass frequency (BPF) = number of blades x running frequency
Vane-pass frequency (VPF) = number of impeller vanes x running frequency

A fan with 8 blades operating at 1,200 rpm has a running frequency of 20 Hz and a blade-pass frequency of 160 Hz. A pump with 6 vanes operating at 1,480 rpm has a running frequency of 24.67 Hz and a vane-pass frequency of approximately 148 Hz.

The pass frequency normally exists because each blade or vane repeatedly moves past a stationary part of the housing. Pressure and flow vary during every pass. A visible peak is therefore expected on many healthy machines. What matters is whether it has changed and what other evidence accompanies it.

What makes a pass-frequency peak important?

A blade- or vane-pass peak deserves investigation when one or more of the following occur:

Its amplitude rises significantly from a comparable baseline.
Harmonics of the pass frequency increase.
Running-speed sidebands appear or grow around the pass frequency.
The response becomes strongly directional or changes with flow, damper position or load.
The time waveform becomes distorted or contains repeating bursts.
Pressure, flow, sound, temperature or power changes at the same time.

Possible causes include dirty or damaged blades, worn vanes, rotor or housing eccentricity, a loose impeller, diffuser or inlet problems, non-uniform blade spacing, flow restrictions, operation away from the intended duty point, or resonance near the forcing frequency. The spectrum alone rarely separates all of these possibilities.

Sidebands and harmonics add diagnostic information

Sidebands indicate modulation. If peaks around BPF or VPF are spaced by 1X running speed, something that changes once per revolution may be modulating the pass event. Possibilities include eccentricity, uneven clearance, a loose impeller, one damaged area, or another once-per-revolution flow disturbance.

For example, a six-vane pump at 24.67 Hz has VPF near 148 Hz. Peaks near 123.3, 148 and 172.7 Hz form 1X-spaced sidebands. This is evidence of modulation; it does not identify the physical cause by itself. Inspect the impeller, clearances and casing, compare radial directions, review phase where useful, and check whether the pattern changes with operating point.

Harmonics of BPF or VPF can result from a non-sinusoidal, strongly distorted or impacting force. Their presence may be normal for some designs, especially if stable. Trend comparable conditions and compare similar machines before deciding that the amplitude is abnormal.

Fan vibration: combine aerodynamics with mechanical checks

Fan vibration can be influenced by rotor unbalance, alignment, bearings, belts and looseness, but it may also respond strongly to the air system. Useful checks include:

Blade condition: inspect for dust buildup, corrosion, erosion, cracks, bent blades and missing balance weights.
Inlet condition: check filters, screens, elbows, dampers and obstructions that can create uneven flow into the wheel.
Outlet and duct condition: look for restrictions, abrupt transitions, flexible-connection problems and unstable damper operation.
Operating point: record flow, pressure, damper position, motor load and speed. A changing process may change vibration without a new mechanical defect.
Structure: compare BPF and its harmonics with natural frequencies. An aerodynamic force can become severe when resonance amplifies it.

On variable-speed fans, collect vibration and process data across a controlled speed range when the procedure permits. If BPF moves with speed but the response becomes large only near one fixed frequency, resonance becomes a strong competing hypothesis.

Pump vibration: start with the operating point

A centrifugal pump should be evaluated with its pump curve and system condition. The best efficiency point (BEP) is the flow at which a particular pump operates most efficiently for a given speed and impeller diameter. The manufacturer normally defines an allowable operating region around it.

Operating too far from the intended region can create unstable hydraulic forces, recirculation, pressure pulsation, higher radial load, noise and vibration. The exact acceptable region is pump-specific; do not turn a general percentage from a training chart into a universal limit.

Always record the operating state with the vibration reading:

Suction and discharge pressure.
Flow rate and valve positions.
Liquid level and temperature.
Speed, motor current or power, and pump configuration.
Whether strainers, filters or parallel pumps have changed.

A route measurement taken at a different flow condition may not be directly comparable with the previous trend.

Cavitation: bubble formation and collapse

Cavitation begins when local liquid pressure falls low enough for vapour bubbles to form. As the bubbles move into a higher-pressure region, they collapse. Repeated collapse can generate noise, random impacts, vibration and material damage.

Common contributors include insufficient available net positive suction head, a restricted suction line or strainer, low tank level, excessive liquid temperature, poor inlet geometry, air ingress, excessive speed, or an operating point that demands more flow than the suction system can supply.

Typical evidence may include:

A gravel-like or crackling sound.
Random high-frequency bursts in the time waveform.
A raised broadband noise floor or broad spectral humps.
Changes at vane-pass frequency and its harmonics.
Unstable suction pressure, flow or discharge pressure.
Loss of performance and, with prolonged exposure, impeller pitting.

These symptoms are not unique to cavitation. Air entrainment, rubbing, a damaged bearing, process solids, resonance and sensor-mounting problems can produce similar evidence. Confirmation requires process checks and, where possible, inspection.

NPSH is a system check, not a vibration label

The pump manufacturer specifies net positive suction head required (NPSHr) for a defined test condition. The system provides net positive suction head available (NPSHa). Reliable operation requires sufficient margin between the available and required values under real operating conditions, according to the manufacturer's guidance and the applicable engineering standard.

A vibration analyst does not confirm adequate NPSH from a spectrum alone. Work with operations or engineering to verify suction pressure, vapour pressure at the actual liquid temperature, static head, suction-line losses, speed and flow. Never throttle a pump's suction valve as an improvised diagnostic test unless an approved procedure specifically permits it.

Recirculation and turbulence can resemble other faults

Internal recirculation can occur when flow separates and circulates within the impeller or casing, often during operation away from the intended region. It may produce low-frequency pulsation, broadband energy, pressure fluctuation and changes near VPF.

External turbulence can result from obstructions, sharp elbows close to the inlet, abrupt area changes, partially closed valves, dirty filters or disturbed inlet flow. In fans and pumps it may appear as broad, unstable low-frequency energy rather than a clean mechanical order.

A practical distinction is responsiveness to process change. If vibration follows flow, pressure, valve or damper position while shaft speed remains nearly constant, a flow-related mechanism becomes more likely. Make only approved operational changes and remain inside the manufacturer's limits.

Do not confuse cavitation with a bearing defect

Evidence	Cavitation or flow issue	Rolling-element bearing defect
Relationship to process	Often changes with flow, suction condition or liquid temperature	May change with load, but normally follows bearing geometry and speed
Waveform	Random or irregular high-frequency bursts	More regularly repeating impacts may be present
Spectrum	Broadband rise, humps and possible VPF changes	Calculated defect families and harmonics may appear, especially in enveloped data
Supporting checks	Pressure, flow, NPSH review, sound and impeller inspection	Envelope spectrum, lubrication condition, ultrasound and bearing inspection

The strongest diagnosis explains both vibration and process behaviour. If the evidence does not separate the alternatives, state the uncertainty and recommend the next discriminating test.

Worked diagnosis: six-vane process pump

A six-vane centrifugal pump operates at 1,480 rpm, or 24.67 Hz. The calculated vane-pass frequency is approximately 148 Hz. The latest measurement shows a higher VPF peak, a raised high-frequency noise floor and irregular bursts in the acceleration waveform. Operators also report a gravel-like sound and unstable discharge flow.

The analyst compares the data with the previous baseline and checks the operating condition:

Speed is unchanged.
Suction pressure is lower than during the baseline measurement.
Liquid temperature is higher.
Flow has increased after a process change.
The suction strainer differential pressure is above its normal range.

Interpretation: the vibration and process evidence support a suction-side flow problem with cavitation as the leading hypothesis. The VPF increase alone would not be enough to reach that conclusion.

Action: operations returns the pump to an approved stable condition and the responsible team inspects the suction path and strainer. Engineering verifies the system's NPSH margin. After the confirmed restriction is corrected, suction pressure, flow, sound and vibration are remeasured under the same operating state. The broadband energy and VPF amplitude return near baseline.

How System 1 and other software help

Condition-monitoring software can calculate and trend BPF or VPF, display spectra and waveforms, create narrow-band alarms, trend broadband high-frequency energy, compare multiple measurement locations, and correlate vibration with speed, flow, pressure, temperature and load.

Online systems are especially useful when a flow problem is intermittent. Time-synchronised process and vibration trends can show whether vibration changes before, after or at the same time as the process condition.

Software still cannot guarantee an exact diagnosis. Blade count, vane count, actual speed, sensor position, frequency range, sampling, mounting and process tags must be correct. The analyst must test competing explanations and close the loop after action.

A practical fan-and-pump workflow

Verify the measurement. Confirm sensor position, direction, mounting, units, Fmax and actual speed.
Calculate forcing frequencies. Mark 1X, harmonics, BPF or VPF, bearing frequencies, belt frequencies and electrical components where relevant.
Record the operating condition. Include flow, pressure, valve or damper position, temperature, level, load and configuration.
Compare with a valid baseline. Use the same speed, load and process state whenever possible.
Read spectrum and waveform together. Look for harmonics, sidebands, broadband energy, modulation, impacts and instability.
Inspect the machine and flow path. Check blades, impeller, clearances, filters, strainers, ducts, piping, supports and visible damage.
Rank competing causes. Separate mechanical, hydraulic or aerodynamic, structural and measurement explanations.
Choose a discriminating test. Correlate process tags, use an approved operating change, review phase or resonance data, and apply other condition-monitoring methods.
Verify after correction. Repeat vibration and process measurements under the same condition.

Final takeaway

BPF and VPF are calculated forcing frequencies, not automatic fault names.
Amplitude change, harmonics, sidebands and operating condition give the peaks meaning.
Cavitation commonly produces noise, random impacts and broadband high-frequency energy, but it must be confirmed with process evidence.
Fans and pumps should be diagnosed as complete machine-and-process systems.
A good report states the evidence, uncertainty, recommended check and verification result.

Coming in Part 10

Part 10: Electric Motor Vibration - Line Frequency, Pole Pass, Rotor Bars and Electrical/Mechanical Separation. We will connect spectral patterns with motor load, current data, speed and mechanical checks.

Discussion question: Have you seen a pump or fan repaired mechanically when the main cause was actually the operating condition or flow path?

References and further learning

Mobius Institute, Vibration Analysis Category II Course Manual, Chapter 16: Pumps, Fans and Compressors.
Emerson Process Management, Basic Vibration Analysis - Course 2031, Chapter 1: Introduction to Vibration, including pump-pass frequency and the fault guide.

Educational note: The patterns and examples in this article are learning guidance, not universal alarm limits, NPSH calculations, operating procedures or acceptance criteria. Follow approved site safety procedures, pump and fan manufacturer requirements, process limits and qualified engineering judgement.

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Wednesday, 16 September 2026

Resonance Diagnosis: Natural Frequencies, Bump Tests and Run-Up/Coast-Down Analysis

From Mechanical Maintenance to Vibration Analysis - Part 8

In Part 7 - Gearbox Faults, we saw that gear-mesh vibration can be amplified when a forcing frequency approaches a structural natural frequency. This is resonance: a condition that can make a modest forcing force produce severe vibration.

Resonance is easily misdiagnosed. A large 1X peak may be blamed on extreme unbalance, a blade-pass peak on a fan defect, or gear-mesh vibration on damaged teeth. The forcing frequency is real, but the structure may be multiplying its response.

A natural frequency belongs to the structure. Resonance occurs only when a forcing frequency excites that natural frequency.

Natural frequency and resonance are not the same

Every shaft, bearing housing, baseplate, foundation, pipe and support has natural frequencies. They are properties of the system's mass, stiffness and damping. A structure can have many natural frequencies, each associated with a particular pattern of movement called a mode shape.

A natural frequency may remain quiet for years. Resonance begins when a periodic force approaches it closely enough to excite the mode. Common forcing frequencies include:

Shaft running speed and its harmonics.
Fan blade-pass and pump vane-pass frequencies.
Gear-mesh frequency and its harmonics.
Reciprocating forces, electrical frequencies and flow pulsation.
Impacts, looseness and forces transmitted from nearby equipment.

The response does not need an exact frequency match. The width of the resonant region depends strongly on damping. A lightly damped structure has a narrow, sharp response with high amplification. Greater damping normally lowers the peak and spreads the response over a wider frequency band.

Mass, stiffness and damping control the response

For a simple mass-spring system, natural frequency increases with stiffness and decreases with mass. The practical relationships are:

More mass: generally lowers natural frequency.
Less mass: generally raises natural frequency.
More stiffness: generally raises natural frequency.
Less stiffness: generally lowers natural frequency.
More damping: generally reduces resonant amplification.

These relationships guide corrections, but real machines have multiple coupled modes. A brace, support or added mass can solve one resonance and move another mode into a forcing-frequency range. Structural changes therefore require engineering review and post-modification testing.

Clues that should make you suspect resonance

Observation	Why resonance is possible	Next test
One unusually large spectral peak	A normal forcing frequency may be amplified	Identify the forcing frequency, then perform a resonance test
High vibration in one direction but much lower in another	Structural stiffness and mode shape are directional	Map amplitude and phase across the structure
Amplitude rises sharply only in a narrow speed range	A speed-related order may be crossing a natural frequency	Run-up or coast-down with a tachometer
A broad hump or raised noise floor surrounds peaks	Several components may be amplified within a resonant band	Bump test or frequency-response measurement
Repeated cracked welds, pipes or supports without another clear cause	Amplified cyclic stress may be driving fatigue	Inspect, map motion and test the suspected structure

These observations justify a test; they do not prove resonance. Unbalance, looseness, misalignment, soft foot, hydraulic excitation and poor measurement technique can produce similar clues.

Amplitude and phase through resonance

As a speed-related forcing frequency approaches a natural frequency, amplitude rises. At the resonant region it reaches a maximum, then falls as the forcing frequency moves above the mode.

Phase provides the stronger confirmation. In a simple single-mode response, phase changes progressively by approximately 180 degrees while passing through resonance, with approximately 90 degrees of lag near the natural frequency. Actual plant data can be distorted by other modes, measurement location, phase wrapping, speed changes and poor tachometer signals, so interpret the complete amplitude-and-phase trend rather than one phase reading.

A Bode plot displays amplitude and phase against speed or frequency. A resonant peak combined with the expected phase transition is much stronger evidence than amplitude alone.

Test 1: the bump or impact test

A bump test introduces a short impact into a stationary structure and measures the frequencies at which it rings. The impact contains energy over a range of frequencies; the structure responds most strongly near its natural frequencies.

Safety comes first. Perform the test only under an approved site procedure. Isolate equipment where required, confirm that stored energy and process hazards are controlled, and never strike rotating, hot, pressurized, fragile or safety-critical components. Select a safe impact point and a suitable hammer or soft mallet that will not damage the surface.

Practical collection sequence

Define the forcing frequency and the suspected direction of movement.
Mount an accelerometer firmly on the structure in that direction.
Select an Fmax high enough to include the suspected mode and use a rectangular/uniform window where the instrument procedure requires it.
Use peak-hold averaging or a triggered impact setup. Low line resolution can help capture a short event quickly; refine the test later if nearby modes must be separated.
Make practice impacts to set the gain without clipping, then collect several consistent single impacts.
Repeat in other directions and at other positions. Avoid judging the structure from a nodal point, where the mode may show little movement.
Compare the response peaks with actual forcing frequencies and with historical tests.

The Mobius Category II manual offers example starting settings such as peak-hold averaging, about 400 lines or fewer, multiple averages and pre-triggering where available. These are instrument-dependent setup guides, not universal requirements. A calibrated impact hammer and frequency-response function provide better control because both input force and structural response are measured.

Test 2: run-up and coast-down analysis

During a controlled run-up or coast-down, shaft speed changes and the machine's orders sweep through a range of frequencies. If an order crosses a natural frequency, the response increases and then decreases. A waterfall plot shows the moving order as a diagonal ridge and the resonant region as a high-amplitude zone near a fixed frequency.

Use a reliable tachometer and collect spectra quickly enough to capture the transition. Order tracking can extract 1X or another order while speed changes and display amplitude and phase on a Bode or polar plot. This is especially useful when the machine starts or stops slowly enough for the mode to respond.

Run-up and coast-down tests must follow the machine manufacturer's limits and site operating procedure. Do not hold a machine near a suspected critical speed merely to improve the plot. Some flexible-rotor machines are designed to pass through critical speeds rapidly.

ODS and modal analysis answer different questions

An operating deflection shape (ODS) uses amplitude and phase measured while the machine operates to animate its motion at a selected frequency. It shows how the machine is moving under the present forces. It does not automatically identify every natural mode.

Modal analysis uses a measured input from an instrumented hammer or shaker and measures the structural response, normally with the machine stopped. It identifies natural frequencies, damping and mode shapes more directly. For complex structures or high-consequence modifications, specialist modal testing or finite-element analysis may be required.

Worked diagnosis: fan vibration near running speed

A belt-driven fan operates at 1,490 rpm, or 24.8 Hz. The fan outboard bearing measures 9.2 mm/s horizontally but only 1.8 mm/s vertically. The spectrum is dominated by 1X. Balancing reduces the calculated unbalance force but the vibration remains high.

The analyst records these findings:

A stationary bump test shows a strong horizontal response at 25.4 Hz.
A controlled coast-down shows the 1X amplitude rising sharply near 25 Hz and falling below it.
The 1X phase changes progressively through the high-amplitude region.
Movement mapping shows the fan base swaying horizontally, with the largest response near the unsupported side.

Conclusion: fan 1X is exciting a horizontal structural natural frequency. Residual unbalance supplies the forcing force, but structural flexibility supplies the amplification.

Action: an engineering review identifies a suitable brace and verifies loads, clearances and foundation condition. After installation, the natural frequency moves away from 1X. The bump test, coast-down and steady-state measurements are repeated under comparable conditions to confirm lower vibration and ensure that no new forcing frequency has been approached.

Correcting resonance without creating another problem

There are four broad strategies:

Change the forcing frequency. Adjust operating speed or process pulsation where the machine and process design allow it.
Move the natural frequency. Change stiffness or mass through an engineered structural modification.
Reduce transmission. Use correctly designed isolation where appropriate.
Add damping. Dissipate energy and reduce amplification.

A separation of roughly 15 to 20 percent between a forcing frequency and a natural frequency is commonly used as a screening guide in the reviewed training materials. It is not a universal acceptance limit. Follow the equipment manufacturer's criteria, applicable standards and qualified engineering analysis.

Do not add a brace, mass or isolator simply because it appears convenient. Confirm the mode shape, check structural loads and piping strain, preserve alignment, and test across all operating speeds and important harmonics after the change.

How System 1 and other software help

Condition-monitoring platforms can correlate vibration with speed and load, store run-up and coast-down data, display waterfall and Bode plots, track orders, compare phase, and trend narrow frequency bands. Modal and ODS packages can animate structural movement and calculate frequency-response functions.

Software organizes evidence; it does not make an exact diagnosis automatically. The analyst must verify the tachometer, sensor direction, mounting, operating condition, test repeatability and physical meaning of every peak and phase change.

An eight-step resonance diagnosis workflow

Identify the forcing frequency. Calculate 1X, harmonics, blade/vane pass, gear mesh, electrical and process frequencies.
Confirm the symptom. Compare directions, locations, operating states and historical trends.
Check measurement quality. Verify sensor mounting, Fmax, resolution, phase reference and tachometer signal.
Form competing hypotheses. Include unbalance, looseness, misalignment, soft foot, hydraulic forces and transmitted vibration.
Select a safe test. Use a bump test, speed variation, run-up/coast-down, ODS or modal analysis according to risk and machine availability.
Combine amplitude and phase. Look for an amplitude maximum and a progressive phase transition through the suspected mode.
Design the correction. Change force, stiffness, mass, isolation or damping only after engineering review.
Close the loop. Repeat the special test and normal route measurements across the operating range.

Final takeaway

Natural frequencies exist in every machine and structure.
Resonance is the amplified response created when a forcing frequency excites a natural frequency.
A large spectral peak alone does not prove a severe mechanical fault.
Directionality, speed sensitivity, bump tests, waterfall plots and phase changes build stronger evidence.
Structural modifications must be engineered and verified across the complete operating range.

Coming in Part 9

Part 9: Fans and Pumps - Blade Pass, Vane Pass, Cavitation and Flow-Related Vibration. We will connect spectra and waveforms with operating point, pressure, flow, recirculation and mechanical condition.

Discussion question: Have you encountered a machine that was repeatedly balanced or aligned before resonance was identified as the real amplifier?

References and further learning

Mobius Institute, Vibration Analysis Category II Course Manual, Chapter 17: Natural Frequencies and Resonance.
Emerson Process Management, Basic Vibration Analysis - Course 2031, Chapter 10: Resonance.

Educational note: The diagnostic patterns and setup examples are learning guidance, not universal alarm limits, test procedures or structural-acceptance criteria. Apply approved site safety procedures, manufacturer requirements and qualified engineering judgement.

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Tuesday, 1 September 2026

Gearbox Faults: Gear Mesh Frequency, Sidebands and Evidence

From Mechanical Maintenance to Vibration Analysis - Part 7

In Part 6 - Rolling-Element Bearing Faults, we learned that a calculated fault frequency is evidence, not an automatic replacement decision.

The same discipline is essential for gearboxes. Gear-mesh vibration exists even when the gears are healthy, its amplitude changes with load, and several shafts may modulate the same mesh frequency. The analyst's job is not simply to find gear mesh frequency. It is to explain what is changing around it.

Gear mesh frequency tells you where tooth contact occurs. Sideband spacing often tells you which shaft is modulating that contact.

Start with a gearbox map, not the spectrum

Before collecting data, sketch the drive train. Record the input, intermediate and output shaft speeds; the number of teeth on every mating gear; the gear type; bearing locations; measurement points; normal load range; lubrication system; and direction of rotation.

This map allows you to calculate each shaft speed and each gear-mesh frequency. Without it, a high-frequency peak may be mistaken for a gear mesh, bearing frequency, motor-bar frequency, blade-pass frequency or structural resonance.

Calculate gear mesh frequency correctly

Gear mesh frequency (GMF) is the rate at which teeth enter mesh:

GMF = number of teeth x rotational frequency of that gear

The same GMF must be obtained from either member of a mating pair. If one result differs, the tooth count, shaft speed or gear pairing is wrong.

Worked calculation

A 24-tooth input pinion rotates at 2,400 rpm, or 40 Hz, and drives a 72-tooth gear.

Input GMF = 24 x 40 = 960 Hz.
Output shaft speed = 960 / 72 = 13.33 Hz, or approximately 800 rpm.
Output GMF check = 72 x 13.33 = approximately 960 Hz.

A multi-stage gearbox has a separate GMF for every mating pair. Work through the train one stage at a time and keep the shaft names consistent.

Why GMF alone is not a defect verdict

Every tooth pair generates a small force variation as contact moves through engagement, rolling and sliding. GMF may therefore appear in a healthy gearbox. Its amplitude is also strongly influenced by transmitted torque, gear design, tooth profile, stiffness, backlash, lubrication, resonance and the sensor transmission path.

A higher GMF amplitude at a higher load does not automatically mean deterioration. Compare measurements at similar speed, load, temperature and process condition. Trend the same point, direction, sensor mounting and acquisition setup.

Observation: GMF increased from 2.0 to 3.5 mm/s.

Question: Did gear condition change, or did torque, speed, alignment, lubrication, resonance or measurement setup change?

Better evidence: comparable operating data plus GMF harmonics, sidebands, waveform impacts, trends and corroborating inspection findings.

Sidebands are modulation evidence

When the amplitude or frequency of a gear-mesh signal changes periodically, the FFT produces peaks on both sides of GMF. If the modulating frequency is f_m, sidebands may appear at:

GMF - f_m, GMF + f_m, GMF - 2f_m, GMF + 2f_m ...

A damaged or eccentric gear rotates once per shaft revolution. Its contact stiffness or tooth load may therefore vary at that shaft's 1X frequency, producing GMF sidebands spaced by that shaft speed.

Measure the spacing; do not judge the cluster by appearance alone. In a gearbox with 40 Hz input speed and 13.33 Hz output speed:

Sidebands spaced by 40 Hz point toward modulation associated with the input shaft or pinion.
Sidebands spaced by 13.33 Hz point toward the output shaft or gear.
Both spacings may indicate that both gears, their alignment, or a shared load path is involved.

The number and amplitude of sidebands often become more useful than the central GMF amplitude. They still do not identify the physical failure mode by themselves; they identify a repeating modulation that must be connected to the machine.

Read the complete pattern

Observed pattern	Possible explanation	Useful next check
Stable GMF with few small sidebands	Normal tooth contact or stable load-related response	Compare with baseline at the same load
GMF sidebands spaced at one shaft speed	Eccentricity, runout, localized tooth damage or load modulation on that shaft	Inspect waveform, phase, gear contact and shaft runout
Higher GMF harmonics with sideband families	Increasing nonlinearity, misalignment, wear, looseness or severe contact disturbance	Check alignment, backlash, mounting, load and oil debris
Once-per-revolution impact in acceleration waveform	Cracked, chipped or broken tooth; localized contact defect	Relate impact period to the responsible shaft and inspect teeth
Broadband high-frequency energy	Impacts, poor lubrication, wear debris, looseness, bearing activity or resonance	Use waveform, oil analysis, envelope data and local comparisons

These are hypothesis patterns, not universal fault rules. Gear geometry, load, transmission path and structural resonance can change the appearance significantly.

Use the acceleration time waveform

Gearbox waveforms are naturally busy because many teeth are engaging. A localized damaged tooth can add a stronger pulse once per revolution of the shaft carrying that tooth. Acceleration usually reveals these short impacts more clearly than velocity.

Set the waveform duration long enough to include several revolutions of the slow shaft. A very short record may show tooth-mesh impacts but hide the slow once-per-revolution modulation. Look for:

Repeating impacts at the input, intermediate or output shaft period.
Amplitude beating that agrees with the measured sideband spacing.
Random bursts that may indicate looseness, debris or intermittent contact.
Peak-to-peak growth even when the velocity RMS changes little.

Measurement setup can make or break the diagnosis

GMF and its harmonics can be far above the frequency range used for routine motor data. The Emerson course recommends an Fmax of approximately 3.5 x GMF where practical so higher harmonics and adjacent sidebands remain visible. SKF gives a similar practical starting point of about 3.25 x GMF. Treat these as setup guides, not severity limits.

Also consider:

Resolution: the frequency spacing between lines must be much smaller than the slowest shaft speed you need to separate.
Sensor: use an accelerometer with adequate high-frequency response.
Mounting: stud mounting normally preserves high-frequency content better than a hand-held probe or loose magnet.
Location: measure near each bearing that supports a gear shaft, in relevant radial directions and axial direction for helical gears where appropriate.
Sampling: avoid aliasing and retain enough waveform samples to capture impacts.
Operating state: record speed, load, direction, oil temperature and transient events.

A single spectrum may require both a wide frequency view and a high-resolution zoom around GMF. The wide view finds harmonics and resonances; the zoom separates closely spaced sidebands.

Variable speed requires order-based thinking

When speed changes, GMF and shaft-related sidebands move. A fixed-frequency trend can miss the peak or combine different operating states. Use tachometer-referenced orders, speed-synchronous sampling, waterfall plots or narrow speed/load bands where available.

During run-up or coast-down, a gear-mesh harmonic may cross a structural natural frequency and grow dramatically. That amplitude increase may be resonance amplification rather than sudden tooth deterioration. A waterfall plot helps separate a speed-following order from a fixed natural frequency.

Do not confuse gear, bearing and process activity

Several sources can occupy the same high-frequency region:

Rolling-element bearing frequencies and their harmonics.
Motor rotor-bar or stator-slot frequencies.
Fan blade-pass, pump vane-pass or compressor lobe frequencies.
Variable-frequency-drive switching activity.
Structural resonance excited by normal gear mesh.
Impacts from looseness, coupling problems or a nearby machine.

Gear mesh is synchronous with shaft speed and tooth count. Bearing defect frequencies are normally non-integer orders and may include cage- or shaft-related sidebands. Calculate all credible sources before assigning a label.

Worked diagnosis: two-stage conveyor gearbox

A conveyor gearbox operates at steady production load. The input shaft runs at 29.5 Hz, the intermediate shaft at 8.2 Hz and the output shaft at 1.9 Hz. The first-stage GMF is 590 Hz. The analyst observes:

GMF remains similar to the historical baseline.
Sidebands around GMF have increased and are spaced at 8.2 Hz.
The same spacing appears around 2xGMF.
The acceleration waveform shows amplitude modulation every 0.122 seconds, approximately one intermediate-shaft revolution.
Oil debris has increased slightly, while bearing envelope trends remain stable.

Observation: a growing first-stage mesh sideband family is modulated at intermediate-shaft speed.

Leading hypothesis: a developing tooth-contact problem associated with the intermediate-shaft gear, such as localized wear, eccentricity or alignment-related load variation.

Competing explanations: load fluctuation at 8.2 Hz, gear resonance, shaft runout, support-bearing clearance, loose mounting or an incorrect shaft-speed calculation.

Recommended actions: repeat data at comparable load; verify shaft speeds and tooth counts; collect high-resolution spectra and longer acceleration waveforms; check phase and runout where practical; review alignment, backlash and lubrication; inspect oil debris; and schedule a borescope or tooth-contact inspection according to risk.

An eight-step gearbox diagnosis workflow

Map the train. Record every shaft speed, tooth count, gear pair, bearing and measurement point.
Record operating condition. Capture speed, torque or load, direction, temperature and lubrication state.
Calculate all frequencies. Include shaft orders, each GMF, GMF harmonics, bearing frequencies, blade/vane frequencies and relevant electrical components.
Verify measurement quality. Confirm sensor, mounting, Fmax, resolution, waveform length and tachometer signal.
Read the pattern. Compare GMF, harmonics, sidebands, broadband energy and the acceleration waveform.
Measure sideband spacing. Connect the spacing to the input, intermediate or output shaft instead of guessing from peak height.
Test competing causes. Review load, alignment, backlash, bearings, resonance, lubrication, looseness, torsional effects and transmitted vibration.
Close the loop. Inspect gears and oil, document the actual failure mode, correct the root cause and repeat measurements under comparable conditions.

The damaged gear may not be the root cause

Common contributors to gear distress include incorrect or contaminated lubricant, inadequate oil delivery, water ingress, excessive or cyclic load, shaft misalignment, soft foot, incorrect backlash, bearing clearance or failure, housing distortion, poor installation, overheating and resonance.

Replacing one damaged gear without checking its mating gear and the cause can create a poor contact pattern and repeat failure. Gear-set, bearing and alignment decisions should follow the gearbox manufacturer's guidance and an inspection of the complete load path.

How System 1 and other software help

Condition-monitoring platforms can store gearbox kinematics, calculate shaft and mesh frequencies, place sideband cursors, trend selected bands, create waterfall plots, correlate vibration with speed and load, and alarm on changing patterns.

Software still depends on correct tooth counts, shaft speeds, sensor locations and operating context. It can display a sideband family instantly; the analyst must decide whether the spacing is physically credible and what test will distinguish the leading hypothesis from its competitors.

Final takeaway

GMF is tooth count multiplied by the rotational frequency of the gear.
GMF can be present in a healthy gearbox and is strongly affected by load.
Sideband spacing often identifies the shaft modulating the mesh.
Harmonics, waveform impacts, trends and oil evidence strengthen the diagnosis.
Correct Fmax, resolution, waveform duration and sensor mounting are essential.
A damaged gear is evidence of a failure mechanism, not necessarily its root cause.

Coming in Part 8

Part 8: Resonance and Natural Frequency Testing. We will examine resonance symptoms, impact testing, run-up and coast-down data, phase changes and practical ways to separate a forcing frequency from a structural response.

Discussion question: When you diagnose a gearbox, which evidence has been most useful - GMF trend, sideband spacing, acceleration waveform, oil analysis or visual inspection?

References and further learning

Emerson Process Management, Basic Vibration Analysis - Course 2031, Section 6: Gear Defects.
SKF, Spectrum Analysis, Section 10: Gears.
Fluke Reliability, Condition Monitoring with VibGuard IIoT Helped Two Customers Avoid Over $1.7M in Damage and Downtime.

Educational note: The patterns are simplified learning guidance, not universal fault rules, alarm limits or remaining-life predictions. Apply approved site procedures, gearbox-manufacturer guidance and qualified engineering judgement to real machinery.

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Thursday, 20 August 2026

Rolling-Element Bearing Faults: Frequencies, Enveloping and Evidence

From Mechanical Maintenance to Vibration Analysis - Part 6

In Part 5 - From Spectral Peaks to Fault Hypotheses, we learned to separate an observation from a diagnosis and test competing explanations.

Now we apply that discipline to one of the most important and frequently misunderstood subjects in vibration analysis: rolling-element bearing faults.

A peak near a calculated bearing frequency is evidence. It is not, by itself, permission to replace the bearing.

Bearing signals may be weak, high-frequency and affected by load, lubrication, mounting, sensor position, structural resonance and speed variation. A reliable conclusion therefore combines the correct bearing geometry, good measurements, trending and physical evidence.

Why a local bearing defect creates vibration

When a rolling element repeatedly passes over a pit, crack or spall, it produces a short impact. That impact excites high-frequency resonances in the bearing, housing and sensor mounting path. The impacts repeat according to the geometry and speed of the bearing component involved.

The raw time waveform may show a train of impacts, while the FFT and envelope spectrum help reveal the repetition rate. Because bearing geometry is not normally synchronized to an exact integer multiple of shaft speed, many bearing frequencies appear at non-integer orders such as 3.58X or 5.42X.

The four calculated bearing frequencies

Frequency	Component represented	Typical supporting clues
BPFO Ball Pass Frequency Outer race	Rolling elements passing a fixed outer-race location	BPFO harmonics; amplitude may be strongest close to the loaded outer-race zone
BPFI Ball Pass Frequency Inner race	Rolling elements passing a defect rotating with the inner race	BPFI harmonics with possible 1X running-speed sidebands
BSF Ball Spin Frequency	Rotation of a ball or roller	BSF or its harmonics with possible cage-frequency modulation
FTF Fundamental Train Frequency	Cage rotational frequency	A low, sub-synchronous component; possible cage, lubrication or load-zone involvement

Use the manufacturer and exact bearing model whenever possible. Calculated frequencies depend on the number and size of rolling elements, pitch diameter and contact angle. Actual frequencies may shift slightly because of slip, load and operating condition, so software cursors should allow a sensible tolerance rather than demanding perfect alignment.

A practical frequency example

Consider a motor running at 1,800 rpm:

Shaft speed = 1,800 / 60 = 30 Hz.
Suppose the bearing database gives BPFO = 3.58X, BPFI = 5.42X, BSF = 2.35X and FTF = 0.40X.
BPFO = 3.58 x 30 = 107.4 Hz.
BPFI = 5.42 x 30 = 162.6 Hz.
BSF = 2.35 x 30 = 70.5 Hz.
FTF = 0.40 x 30 = 12 Hz.

If the envelope spectrum contains peaks near 107.4, 214.8 and 322.2 Hz, BPFO and its first two harmonics become a reasonable outer-race fault hypothesis. The analyst must still verify the bearing identity, actual speed, trend, sensor position and competing impact sources.

Why enveloping helps detect early damage

Early bearing impacts may contain little energy compared with normal shaft vibration. In a conventional velocity spectrum, strong 1X and 2X components can dominate the display while the bearing signal remains hidden in the high-frequency noise floor.

Envelope analysis, also called demodulation, typically:

Measures acceleration at a suitable sample rate.
Uses a high-pass or band-pass filter around a high-frequency resonance excited by the impacts.
Rectifies the filtered signal so the ringing transients can be followed.
Applies a low-pass stage and calculates the envelope spectrum to expose the lower-frequency impact repetition rate.

The result may reveal BPFO, BPFI, BSF or FTF before those components become obvious in a normal velocity spectrum. Enveloping does not replace the waveform or conventional spectrum; it adds another view of the same machine behaviour.

Envelope setup matters

The filter band and sampling rate determine what the instrument can reveal. The high-pass or band-pass filter should reject strong lower-frequency machine vibration while retaining a resonance band excited by short bearing impacts. The envelope-spectrum Fmax must then be high enough to display the bearing frequencies and several harmonics.

As a practical starting point, the Mobius Category II manual recommends an envelope-spectrum Fmax of roughly 3.5 to 5 times the calculated BPFI. If the bearing is unknown and has approximately 8 to 12 rolling elements, it suggests starting near 15X to 20X running speed. These are setup guides, not alarm limits; adjust them for the bearing, machine and instrument.

Choose the resonance band carefully. Motor-bar and variable-frequency-drive switching activity, gearmesh, reciprocating-machine impacts, rotating looseness and cavitation can also generate high-frequency energy. A strong envelope reading therefore proves that repetitive impacts or modulation are present, not automatically that the bearing is damaged.

Conventional velocity: useful for overall machine condition and lower-to-mid-frequency faults such as unbalance, misalignment and looseness.

Acceleration: sensitive to higher-frequency energy and impacts.

Envelope spectrum: useful for identifying the repetition rate hidden inside high-frequency impact energy.

Read harmonics and sidebands as modulation clues

A single peak close to a calculated frequency is weak evidence. Confidence improves when a physically meaningful family appears.

Harmonics: repeated impacts can generate BPFO, 2xBPFO, 3xBPFO and higher multiples.
1X sidebands around BPFI: an inner-race defect rotates through the bearing load zone, so impact amplitude may be modulated once per shaft revolution.
FTF sidebands around BSF: rolling-element damage may change as the element moves around the bearing and contacts the inner and outer raceways.
Rising noise floor: more distributed surface distress, contamination or poor lubrication can create broadband high-frequency energy.

Sideband spacing matters. Do not call every cluster a bearing fault; measure the spacing and relate it to shaft speed, FTF or another credible modulating frequency.

High-frequency vibration normally attenuates more rapidly through a structure than low-frequency vibration. Compare equivalent high-frequency readings at nearby bearings: a signal concentrated at one housing provides useful location evidence, while the same low-frequency pattern at several points may be transmitted from another source.

A simplified progression of bearing deterioration

Bearing damage does not always follow a neat universal sequence, but analysts often see a progression similar to this:

Earliest indication: high-frequency energy or the envelope trend rises while conventional velocity changes little.
Localized race defect develops: a bearing-frequency family and harmonics become visible, often first for one race.
Damage spreads: more harmonics and sidebands appear; rolling-element or cage-related activity may join the race frequencies, and impacts become clearer in acceleration and the waveform.
Advanced deterioration: broadband energy and the noise floor rise, discrete peaks may become less distinct, clearances and internal geometry may change, and conventional velocity can become significant.

Do not misinterpret falling discrete peaks as recovery. Severe loss of internal geometry can reduce the clear transmission of individual frequencies while broadband energy and random impacts continue to increase.

Do not use this sequence as a countdown clock. Remaining life depends on the size and number of defects, the components involved, loss of internal geometry, rate of progression, load, speed, lubrication, service time and experience with genuinely comparable machines. No two cases are identical.

Do not confuse the damaged component with the root cause

A vibration pattern may identify where damage is occurring without explaining why it occurred. Possible contributors include:

Insufficient, excessive, incorrect or contaminated lubricant.
Water or solid-particle contamination.
Misalignment, unbalance, belt forces or excessive external load.
Incorrect fits, loss of internal clearance or installation damage.
Electrical discharge or fluting in motors and variable-frequency-drive applications.
Resonance, looseness or impacts transmitted from a nearby machine component.

Replacing the bearing without correcting the cause may simply restart the failure cycle.

Worked example: motor drive-end bearing

A motor runs at 1,780 rpm, or approximately 29.67 Hz. The correct bearing model gives BPFI = 5.45X. The analyst observes:

Envelope peaks near BPFI and 2xBPFI.
Sidebands around BPFI spaced at approximately 1X running speed.
The envelope trend has doubled over six weeks at comparable load.
Conventional velocity has increased only slightly.
The drive-end bearing temperature is stable.
Lubrication history shows a recent change in grease quantity.

Observation: a growing, speed-related impact family aligns with BPFI and includes 1X sidebands.

Leading hypothesis: developing inner-race-related distress.

Competing explanations: an incorrect bearing number, transmitted impacts, electrical activity, lubrication-related friction or a mounting problem.

Recommended actions: confirm actual speed and bearing identification; review the grease type, quantity and procedure; inspect motor current and grounding evidence where relevant; increase measurement frequency; and plan inspection or replacement according to risk and trend rate. After maintenance, inspect the removed bearing and document the failure mode rather than recording only “bearing replaced.”

A seven-step bearing diagnosis workflow

Verify the bearing and operating condition. Record model, speed, load, temperature and lubrication state.
Verify the measurement. Use a repeatable point, firm sensor mounting, suitable frequency range and adequate sampling.
Review multiple displays. Compare velocity, acceleration, envelope spectrum, waveform and trends.
Overlay the correct frequencies. Look for harmonics and meaningful sideband spacing, allowing for reasonable slip.
Compare locations and history. Check nearby bearings, directions, baselines and comparable machines.
Test the root-cause and false-positive hypotheses. Review lubrication, contamination, load, alignment, fits, electrical switching, gearmesh, cavitation and other impact sources.
Verify the intervention. Inspect the removed bearing and repeat measurements under a comparable condition.

High-frequency methods are deliberately sensitive. Unless risk, trend rate or a low-speed application justifies earlier action, avoid overhauling a machine from one envelope spectrum alone. Seek confirmation in repeatable trends, the conventional spectrum, waveform, lubrication condition or another independent technique.

How software helps - and where judgement remains essential

System 1 and other condition-monitoring platforms can store bearing data, calculate or import defect frequencies, apply frequency markers, trend amplitudes, display waveforms and spectra, and generate alarms. These features improve consistency and make developing patterns easier to see.

However, the software cannot guarantee that the installed bearing matches the database, that the speed is correct, that the sensor is well mounted, or that the peak comes from the bearing. The analyst still has to connect the signal to the machine and select the next discriminating test.

Final takeaway

BPFO, BPFI, BSF and FTF are geometry-based diagnostic frequencies, not automatic failure verdicts.
Early bearing damage is often easier to see in high-frequency acceleration and envelope data than in overall velocity.
Harmonics, sidebands, waveforms and trends provide stronger evidence than a single peak.
A damaged bearing is not necessarily the root cause of the failure.
The best diagnosis combines vibration, operating condition, lubrication history and physical inspection.

Coming in Part 7

Part 7: Gearbox Faults and Sideband Analysis. We will examine gear mesh frequency, harmonics, sidebands, modulation and how to separate gear damage from shaft-speed and bearing-related activity.

Discussion question: Which bearing evidence do you trust most in your plant - envelope trends, defect-frequency patterns, temperature, lubrication findings or inspection results?

References and further learning

Emerson Process Management, Basic Vibration Analysis - Course 2031, Section 5: Rolling Element Bearings.
Mobius Institute, Category II Vibration Analysis Manual, Chapter 13: Rolling Element Bearing Analysis.
SKF, Spectrum Analysis.
SKF Evolution, Vibrations Help Find Faulty Components.
Fluke, How Demodulation Helps You Detect Bearing Faults Early.
Fluke, Smarter Bearing Monitoring.

Educational note: The patterns are simplified learning guidance, not universal fault rules, alarm limits or remaining-life predictions. Apply approved site procedures, bearing-manufacturer guidance and qualified engineering judgement to real machinery.

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Tuesday, 18 August 2026

A Horrible Noise That Disappeared: Diagnosing a Loose Motor Cooling-Fan Part

A strange mechanical noise does not have to remain present for the fault to be real.

In a Bently Nevada success story from an offshore oil and gas platform, a technician heard a “horrible noise” from a dissolved-salts pump motor during routine vibration data collection. The noise disappeared shortly afterward, but the measurements taken during the event preserved the evidence. What followed is a valuable lesson in combining human observation, spectral analysis, waveform interpretation and focused inspection.

A transient sound may disappear before an inspection begins, but correctly captured vibration data can preserve the mechanical signature of the event.

The operating context

The machine was considered moderately critical and was monitored periodically using a portable Scout 220 data collector. Offshore technicians collected the readings, uploaded them to a System 1 server and requested remote diagnostic support from Bently Nevada Machinery Diagnostic Services engineers.

This arrangement matters. It shows how a condition-monitoring program can connect three important capabilities:

People near the machine who notice abnormal sound, smell, heat or movement.
Portable measurements that preserve the spectrum and time waveform.
Diagnostic expertise that can interpret the data and guide a targeted inspection.

What the vibration data showed

The measurement taken while the noise was present contained a single prominent peak at approximately 1210 Hz, accompanied by 1X and 2X running-speed sidebands. The unusual vibration appeared only at the motor non-drive end and was absent from the later measurement taken after the noise stopped.

The 1210 Hz component did not match an expected forcing frequency for the machine. This prevented the analyst from simply assigning it to a normal rotating or electrical source. Instead, the spectrum and waveform suggested that repeated impacts were exciting a local resonance near 1210 Hz. The 1X and 2X sideband spacing showed that the high-frequency response was being modulated at running-speed-related intervals.

Measured observation: A 1210 Hz peak with 1X and 2X sidebands appeared at the motor non-drive end during the audible event.

Interpretation: Periodic impact or rubbing was likely exciting a structural resonance.

Location hypothesis: A loose rotating part or contact between rotating and stationary parts near the overhung cooling fan.

Discriminating action: Open the fan cover and inspect the cooling-fan assembly.

Why sidebands were important

A sideband is a spectral component that appears on either side of a carrier frequency. In this case, the carrier was the resonance near 1210 Hz. Regular variation in the amplitude of that vibration produced sidebands separated by the modulating frequencies.

The key point is not merely that sidebands existed. Their spacing connected the high-frequency resonance to a once-per-revolution and twice-per-revolution mechanical event. That relationship supported a hypothesis involving a rotating component repeatedly contacting, striking or changing load as the shaft turned.

Sidebands should always be interpreted with the waveform, machine speed, measurement location and equipment geometry. They describe modulation; they do not identify the damaged component by themselves.

Why the noise was so noticeable

The abnormal frequency was high enough to be heard clearly by the technician. Baker Hughes noted that it was near the frequency range where human hearing is particularly sensitive. This helps explain why the event sounded severe even though it was temporary.

Human senses remain useful screening tools in maintenance. An experienced technician may recognize a change before an alarm is triggered. However, the safest workflow is to treat sound as an observation, capture objective data and avoid approaching or opening moving equipment until it is isolated under the approved procedure.

The inspection confirmed the diagnosis

Because the abnormal response was localized at the non-drive end, the diagnostic team suspected the overhung cooling fan. They recommended removing the fan cover and looking for abnormalities.

The inspection found part of a broken retaining ring loose inside the cover. The ring belonged to the fan assembly and had separated. The technician replaced it, and the machine returned to service less than 48 hours after the abnormal measurement.

According to the source case study, failure to detect the problem could have allowed the fan to separate from the rotor and destroy the motor. The estimated replacement cost of the motor was approximately €40,000. The larger value, however, also included avoided secondary damage, unplanned downtime and operational risk.

A practical diagnostic workflow

Record the human observation. Note what was heard, where it was strongest, when it began and whether operating conditions changed.
Preserve event data. Save the spectrum, time waveform, speed, load, measurement direction and timestamp. Do not overwrite an abnormal reading with a later normal one.
Compare locations. A response isolated to one end or one direction can greatly reduce the search area.
Identify measured facts. List the dominant frequency, sideband spacing, waveform features and differences from the baseline.
Develop more than one hypothesis. Consider loose parts, rubbing, impacts, bearing faults, aerodynamic effects, electrical sources and structural resonance as appropriate.
Use machine geometry. Ask which component near the measurement point could physically produce the observed periodic event.
Choose a focused inspection. Inspect the suspected area under proper isolation instead of dismantling the entire machine.
Verify after repair. Repeat comparable measurements and confirm that the abnormal sound and vibration signature are gone.

Lessons for a condition-monitoring program

Intermittent faults deserve immediate attention. Disappearance of the symptom is not proof that the defect corrected itself.
Keep the abnormal dataset. Event measurements may contain evidence that routine follow-up readings no longer show.
Use both spectrum and waveform. Frequency-domain sidebands reveal modulation, while the time waveform helps show the underlying impact pattern.
Measurement location matters. Localization to the motor non-drive end directed attention toward the cooling fan.
Combine technology with technician experience. The technician’s report of abnormal noise gave essential context to the data.
Convert diagnosis into a specific action. The recommendation was not simply “monitor closely”; it identified the fan cover as the next safe inspection point.
Moderately critical assets can still create major losses. Periodic portable monitoring can prevent expensive failures outside the permanently monitored machine population.

Final takeaway

This case is a strong example of evidence-based vibration analysis. A temporary noise was captured as a localized high-frequency resonance with running-speed sidebands. The pattern suggested impact or rubbing near the motor cooling fan, and a focused inspection revealed a broken retaining ring.

The successful diagnosis did not come from one spectral rule. It came from linking sound, timing, location, spectrum, waveform, machine geometry and physical inspection. That is the habit that turns condition-monitoring data into reliable maintenance decisions.

Source and further reading

This educational analysis is based on the Baker Hughes/Bently Nevada success story Loose Part on Motor Cooling Fan, authored by Håkon Myklestad of Norway MDS. Refer to the original case study for its spectrum, waveform and damaged-component images.

Educational note: This article summarizes a published case and expands on its diagnostic lessons. Actual machinery decisions must follow approved isolation procedures, manufacturer guidance, applicable standards and qualified engineering judgement.

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Friday, 14 August 2026

Thank You for 1,003 Views of the Vibration Analysis Game!

Thank You for 1,003 Views!

We are truly grateful to everyone who has visited and explored the RCM GLOBAL Vibration Analysis Game.

Reaching 1,003 views is an encouraging milestone. Every visit, answer and shared link helps us build a stronger learning community for vibration analysts, maintenance professionals, engineers and students.

If you have not played yet—or would like to test your diagnostic skills again—click below:

▶ Play the Vibration Analysis Game

Please share the game with colleagues and friends who are interested in condition monitoring, predictive maintenance and reliability engineering.

Thank you for learning and growing with RCM GLOBAL!

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