Ship Vibration: Hull, Propeller & Engine Excitation

Ship vibration is the cyclic mechanical oscillation of a vessel’s hull, machinery spaces, and accommodation structure, driven by periodic forces from the propeller, the main engine, and wave excitation. It costs the industry money in three concrete ways: accelerated fatigue damage to hull structural details, equipment malfunction from resonant responses, and charter-party disputes or vessel rejections when accommodation vibration exceeds class comfort-notation limits. A slow-speed two-stroke bulk carrier built without vibration analysis in the 1990s often had to retrofit mass dampers at a cost of USD 200,000 to 600,000 per installation; modern newbuilds use finite element analysis from the preliminary design stage to avoid that outcome.

This article covers the physics of hull-girder vibration, the principal excitation sources, the ISO 6954:2000 habitability standard, classification society comfort notations, and the countermeasures available at design and in service. The closely related topic of propulsion-shaft torsional vibration, which is governed by a separate regulatory framework (IACS UR M68), is covered in the engine torsional vibration analysis article. That article should be read alongside this one: many ships experience both hull-girder vibration from the propeller and shaft-train torsional vibration from the engine simultaneously, and separating them diagnostically is a key skill.

Vibration classification: global modes and local modes

Ship vibration divides into two categories with different physics and different regulatory treatments.

Global (hull-girder) vibration involves the entire ship deforming as an elastic beam. The hull girder supports a set of natural frequencies corresponding to bending and torsional modes. The lowest is the 2-node vertical bending mode (where the hull bends like a simply-supported beam with the ends and midship as nodes), with a natural frequency that typically falls between 0.5 Hz and 2.5 Hz for vessels above 100 m. Higher modes follow at approximately 2x, 3x, 4x the fundamental. Horizontal bending and torsional modes sit at different frequencies, often coupling in open-deck vessels such as container ships. When a propeller blade-rate or engine excitation frequency coincides with a hull-girder natural frequency, the response involves the entire ship and can be measured at the stern, at midship, and at the bow simultaneously.

Local (panel and superstructure) vibration involves individual structural panels, decks, bulkheads, stiffeners, and the superstructure responding in their own resonant modes at frequencies from 5 Hz to 50 Hz and above. These are independent of the global modes and cause the buzzing, rattling, and high-frequency vibration felt in individual cabins or on specific deck areas. Local resonances can exist even when global vibration levels are acceptable.

The hull-strength and longitudinal bending article covers the static and quasi-static loading of the hull girder that provides the structural context for the dynamic (vibration) problem described here.

Excitation sources: propeller

The propeller is the dominant vibration source on most single-screw merchant ships. It produces two distinct classes of excitation: surface forces and bearing forces.

Blade-rate pressure pulses and surface forces

As each blade rotates through the non-uniform velocity field behind the hull, the lift it generates varies periodically. The pressure fluctuations radiate outward and produce a time-varying pressure on the hull surface above and around the propeller. This is the blade-rate pressure pulse, and it excites the hull at the first blade-rate frequency and its harmonics:

where $Z$ is the blade count and $n$ is shaft speed in revolutions per minute. The second harmonic occurs at $2 f_{BR}$, the third at $3 f_{BR}$, and so on. For a 5-bladed propeller at 100 RPM, $f_{BR1} = 8.33$ Hz, $f_{BR2} = 16.67$ Hz.

The magnitude of the pressure pulse depends on propeller loading, blade geometry, and the hull-propeller clearance. Classification societies require a minimum tip clearance of 20 to 25 percent of propeller diameter ($D$) for single-screw vessels. Below this threshold the pressure pulse amplitude increases inversely with tip clearance, and at clearances below 15 percent of $D$ the pulse can be 3 to 5 times larger than at 25 percent. The ITTC Recommended Procedures 7.5-02-01-03 (Propulsion/Propeller Open Water Test) and the DNV class rules both set explicit minimum clearance requirements as a design check.

Cavitation greatly amplifies blade-rate excitation. When the pressure on the suction side of a propeller blade drops below vapour pressure, a cavitation sheet forms and then collapses as the blade rotates into higher-pressure water. The collapse produces a short-duration pressure spike that is broadband in frequency content, much higher in amplitude than the non-cavitating blade-rate pulse, and essentially impossible to predict accurately without high-fidelity CFD combined with model-scale tests. ABS guidance notes (2022 edition) state that a full-scale cavitation-induced pressure pulse can be 5 to 10 times larger than the non-cavitating value when sheets are large. This is why propeller design for low vibration requires controlling cavitation extent, not just clearance.

Bearing forces

The propeller also transmits thrust and torque fluctuations directly through the shaft to the thrust bearing and from there to the ship structure. These are called bearing forces or shaft forces. For a single fixed-pitch propeller operating in uniform inflow, the net thrust and torque fluctuate at the blade rate. In non-uniform inflow (which is the reality behind a ship’s hull), the individual blade loadings vary as each blade sweeps through the wake field, producing alternating forces and moments at the blade rate and harmonics. These forces propagate through the shaft and excite the engine structure as well as the hull; the shaft torsional-vibration aspect is covered in engine torsional vibration analysis.

The propeller-induced bearing forces on a 5-bladed propeller can be decomposed into: alternating thrust $F_T$, horizontal force $F_H$, vertical force $F_V$, and moments $M_H$ and $M_V$ about horizontal and vertical axes through the shaft centerline, all varying at $f_{BR}$ and its harmonics. The non-uniform wake is ship-specific and requires a wake survey at model scale or CFD to quantify accurately.

Excitation sources: main engine

The main engine contributes three classes of vibration excitation, all related to the combustion cycle.

Free forces and moments

A reciprocating engine has unbalanced rotating and reciprocating masses that produce free forces and free moments in the vertical and horizontal planes. For a two-stroke slow-speed engine (the dominant type on bulk carriers, tankers, and container ships), the dominant unbalanced excitation is:

First-order moments (at engine RPM): produced by the first-order unbalanced couple from the crank arrangement. For in-line engines, MAN B&W / WinGD publish the first-order vertical and horizontal moments explicitly in the engine project guide for each cylinder count. A 6-cylinder engine with no built-in balance weights produces a significant first-order vertical moment; 7 and 8 cylinder engines have smaller first-order unbalances but non-zero second-order contributions.

Second-order forces and moments (at 2x engine RPM): particularly relevant for two-stroke engines with long stroke-to-bore ratios. The second-order unbalanced vertical force can be large in 4 and 5 cylinder engines. MAN’s project guide tables for the G-type and S-type engines list, for each cylinder count, whether first- and second-order balance masses are standard, optional, or not available.

External moments from the guide forces (the horizontal transverse forces that the connecting rod transmits to the crosshead guide): these are the H-type moments that act in the horizontal plane and tend to excite transverse hull vibration or cause “rocking” motion.

The firing frequency and its harmonics represent the combustion-related torque pulses already covered in the torsional vibration context, distinct from the free forces above.

Makers of slow-speed two-stroke engines (MAN Energy Solutions, WinGD) supply a vibration data sheet with each engine specifying the unbalanced forces and moments for the specific cylinder configuration. These are the input to the global vibration analysis. For four-stroke medium-speed engines (Wartsila, MAN L/V, Caterpillar, Bergen), the reciprocating unbalances are typically higher order and better self-balanced by the crankshaft arrangement, but the mounting must be checked for the engine’s natural frequencies against the firing frequency.

Engine on elastic mounts

Medium-speed four-stroke engines in ferries, cruise ships, and offshore vessels are often installed on elastic (resilient) mounts to isolate the engine vibration from the hull. The isolation efficiency depends on the ratio of the excitation frequency to the mounted natural frequency of the engine. For effective isolation the mount frequency should be at least 2 to 3 times below the lowest firing harmonic that matters. At below this ratio the mount amplifies rather than isolates. DNV rules and Lloyd’s Register class rules require that the mounted natural frequency be confirmed against the operating speed range for all elastic-mount installations.

Hull-girder natural frequencies

Global vibration modes

The hull girder in free-free condition (free at both ends, i.e., floating) supports a series of natural frequencies. The principal ones that concern the vibration designer:

2-node vertical bending ($f_{V2}$): the lowest vertical bending mode. The bow and stern move in opposite phase with a node near each quarter-length and a second node does not exist in this counting convention: the “2-node” name refers to two points of zero displacement along the length.

3-node vertical bending ($f_{V3}$): the second vertical bending mode, at roughly 2 to 2.5 times $f_{V2}$.

2-node horizontal bending ($f_{H2}$): the lowest horizontal mode, typically between 0.8 and 1.5 times $f_{V2}$ depending on hull form.

Torsional mode: in open-section vessels (container ships, some ro-ro), the first torsional mode can couple with horizontal bending and produce combined torsional-horizontal vibration, with natural frequencies sensitive to the open-deck torsional stiffness. The open-deck torsion calculator computes the warping and St. Venant torsional stiffness that governs this mode.

Schlick formula

The Schlick formula (1884, still used in preliminary design) estimates the 2-node vertical natural frequency:

where $I_{mid}$ is the midship section second moment of area about the horizontal axis (m$^4$), $\Delta$ is the ship displacement (tonnes), $L$ is the length between perpendiculars (m), and $C_S$ is an empirical constant (typically 1.28 $\times$ 10$^5$ to 1.64 $\times$ 10$^5$ in consistent SI-compatible units from the Tasai 1965 refinement). The formula was derived from measurements on steam vessels and should be treated as a preliminary check only. A finite element model is required for detailed design.

The Schlick formula result should be compared to the first blade-rate $f_{BR1}$ and first engine excitation frequency $f_{E1}$ (engine RPM in Hz). If $f_{V2}$ falls within 15 percent of either, a detailed resonance check is mandatory.

Local resonances

Deck panels, bulkhead plating, and the superstructure can have local natural frequencies in the 10 to 50 Hz range that coincide with higher blade-rate harmonics or with equipment excitation. These are harder to predict by hand calculation and require finite element analysis of the local structure. The 2-node hull-girder frequency does not predict where a panel will resonate. A panel with large unsupported area and thin plating may have a natural frequency of 18 Hz, which coincides with the third blade-rate harmonic of a 5-bladed propeller at 108 RPM. The design fix is either stiffening (raising the panel frequency) or adding mass (lowering it), chosen to detune from the excitation.

ISO 6954:2000 habitability standard

ISO 6954:2000 is the primary international standard for vibration habitability on passenger and merchant ships. It replaced the 1984 edition, whose absolute velocity limits proved difficult to apply uniformly; the 2000 edition moved to frequency-weighted acceleration.

Measurement and evaluation method

Vibration is measured with accelerometers at defined positions: the deck at the centre of the occupied area in each space, with the measurement point at foot level (standing position) or at seat level (seated position), per ISO 20283-2:2016. The measurement duration is a minimum of 60 seconds per location, at each operating condition (sea speed, reduced speed, manoeuvering). The measured acceleration time histories are frequency-weighted using the $W_d$ filter (ISO 2631-1:1997), which peaks at 4 to 12.5 Hz to represent human sensitivity to whole-body vibration. The RMS of the weighted signal is the evaluation quantity.

Acceptance criteria

ISO 6954:2000, clause 5.3, defines three zones:

The lower limit of 72 mm/s$^2$ and upper limit of 214 mm/s$^2$ apply to the 1 to 80 Hz band. These are the limits for crew accommodation and working spaces. ISO 6954 does not differentiate by space type within these two limits; the two-tier structure is the entire normative content for habitability classification. Class societies have built tighter multi-tier comfort notations on top of these limits.

ISO 20283 series

ISO 20283 is a companion set of measurement standards:

• ISO 20283-1:2016: general guidelines for measurement on ships.
• ISO 20283-2:2016: measurement of structural vibration (accelerometer placement, reporting requirements).
• ISO 20283-3:2008: measurement of noise and vibration for habitation purposes.
• ISO 20283-4:2012: measurement of vibration of ship structures for fatigue analysis.

Surveyors conducting comfort-notation trials for DNV, Lloyd’s Register, or ABS cite ISO 20283-2 as the measurement protocol.

Classification society comfort notations

Class society comfort notations go beyond the two-tier ISO 6954 structure and impose specific acceleration limits, measurement protocols, and design review requirements. The major societies’ notations are compared in the table below.

Comparison of comfort notations

The DNV COMF-V notation is the most frequently cited in newbuilding contracts for cruise ships and high-comfort ferries. DNV publishes requirements in the Rules for Classification of Ships, Part 6 Chapter 24 (Special Equipment and Systems: Comfort Class). The VC (Vibration Comfort) notation, accessible via the DNV VC notation checker, covers vessels seeking verification without the full COMF programme. ABS issues its VSA notation after a trial measurement programme; the ABS VSA calculator supports eligibility checks.

Survey programme

For all comfort notations, a sea trial vibration survey is conducted:

1. Full speed ahead in calm water (Beaufort 3 or below), measuring at all designated spaces.
2. Reduced speed (typically 80 percent and 60 percent of MCO), to map the speed-dependent response and identify resonances.
3. Maneuvering (crash stop, turning circle), for offshore vessels and those with azimuth thrusters.

Any space that exceeds the notation limit requires the shipyard to remedy, re-measure, and re-certify before the notation is granted. Typical remedies applied post-trials include adding mass to deck panels (shifts frequency), fitting local stiffeners, and installing tuned vibration absorbers.

Blade number selection and resonance avoidance

The choice of blade number is one of the earliest vibration design decisions. The goal is to place the first blade-rate frequency in a window that avoids both the 2-node hull-girder frequency and the main engine harmonics.

For a slow-speed two-stroke plant at 80 to 105 RPM:

• 4 blades: $f_{BR1}$ = 5.3 to 7.0 Hz. This range is close to the 2-node hull frequency of large vessels (0.5 to 1.5 Hz for VLCCs and ULCCs, up to 2.5 Hz for smaller vessels). The fourth harmonic at 21 to 28 Hz can intersect local panel resonances.
• 5 blades: $f_{BR1}$ = 6.7 to 8.75 Hz. The most common choice for bulk carriers and tankers. The first harmonic at 5 blades is generally above the 2-node hull frequency and below the main engine firing frequency.
• 6 blades: $f_{BR1}$ = 8.0 to 10.5 Hz. Used on container ships and LNG carriers, where the higher ship speed brings higher RPM and different frequency windows.

There is no universally optimal blade count; it depends on the ship’s size, speed, engine type, and structural natural frequencies. The designer maps the relevant frequencies at the concept stage and selects the blade count that leaves the largest detuning margin.

The detuning margin recommended by classification society guidance (DNV rules, ABS Guide for Ship Vibration) is 20 percent separation between the blade-rate frequency and the nearest hull or structural natural frequency:

If this criterion cannot be met by blade-count selection alone, hull stiffening, mass redistribution, or a change in propeller RPM (via a different gearbox ratio or engine model) is considered.

Hull stiffening and countermeasures

When the preliminary design shows a resonance risk, four families of countermeasure are available.

Increase global stiffness. Adding longitudinal structure amidships raises $f_{V2}$. The cost is structural weight; the benefit is permanent detuning. Typical modifications are doubling of the keel box, addition of inner-bottom longitudinals, or thickening of the sheer strake.

Add mass. Adding ballast, structural mass, or equipment mass at specific locations lowers the hull frequency and shifts mode shapes. This is less common as a deliberate design tool because added mass hurts deadweight; it is more useful as a retrofit.

Tuned mass dampers. A tuned mass damper (TMD) is a secondary mass-spring-damper system attached to the structure, tuned so that its natural frequency equals the target excitation frequency. At resonance, the TMD absorbs energy from the primary structure. Marine TMDs for superstructure vibration typically use masses of 1 to 8 tonnes on steel springs or hydraulic actuators. They are narrow-band devices: a TMD tuned to 8.0 Hz provides little benefit at 8.5 Hz. Several post-delivery retrofits on container ships (in the 2000s and 2010s) used TMDs in accommodation superstructures to cure resonant responses to the 5th and 6th engine harmonics.

Active vibration control. Electromagnetic or hydraulic actuators driven by a feedback controller can cancel vibration at specific points. This is established technology on certain warships and research vessels, where the cost is justified. It is rarely used on commercial merchant ships.

Propeller-related countermeasures. If the blade-rate frequency is the cause, and the frequency window is fixed by machinery constraints, options are: (a) reduce pressure pulse amplitude by increasing tip clearance (requires aft-body redesign); (b) change to a skewed propeller, where blade entry into the wake is staggered over a longer arc, reducing the peak pressure pulse; (c) fit a controllable-pitch propeller (CPP) to allow RPM optimization in service; (d) add a wake-equalizing duct or upstream stator to make the inflow more uniform. Item (b), highly skewed propellers (35 to 45 degrees of skew), is now standard on newbuildings for cruise ships and ferries.

Vibration monitoring in service

Permanent vibration monitoring has become standard on vessels with comfort-class notations. Accelerometers are placed on machinery foundations, shaft bearings, accommodation deck structures, and at the stern. Data is logged continuously and analyzed for frequency-domain signatures.

The main engine room automation and monitoring systems on modern vessels (covered in marine engine room automation and monitoring) include vibration as one of the parameters in the condition monitoring programme. ISO 10816-6 governs the machinery vibration severity criteria: the ISO 10816 vibration severity zone calculator applies those zones to measured RMS velocities on pumps, fans, and auxiliary engines.

For the propulsion shaft, the shaft torsional critical speed calculator checks whether the shaft train torsional critical speed falls within the barred speed range defined by IACS UR M68 and the engine maker. This check is required for every newbuild by class.

Propeller cavitation monitoring using hull-mounted hydrophones (listening for cavitation noise) and hull vibration accelerometers (detecting the broadband shock from cavitation collapse) is used on some naval and research vessels. Commercial application is growing but not yet universal.

Propeller-induced stern vibration: the dominant complaint source

On single-screw merchant ships, vibration complaints from the accommodation aft almost always originate from propeller blade-rate excitation of the stern structure. The stern is close to the propeller, it has the highest pressure pulse, and accommodation cabins are often cantilevered above the propeller aperture on bulk carriers and tankers. This combination means that even when global hull vibration is acceptably low, the aft accommodation can violate ISO 6954 limits through local resonance of the stern panels.

The propeller excites the stern at $f_{BR1}$ and $2 f_{BR1}$. If a deck panel above the propeller has a natural frequency near $f_{BR1}$, that panel resonates. The practical check is: calculate panel natural frequency from standard plate vibration formulas, compare to $f_{BR1}$, and if within 15 percent, add a stiffener to raise the panel frequency or a small mass to lower it.

The hull-propeller clearance is the most effective single parameter controlling aft accommodation vibration. For a given propeller loading, reducing clearance from 25 percent of $D$ to 20 percent of $D$ roughly doubles the pressure pulse amplitude. A 15 percent clearance, sometimes seen on full-form vessels where the hull lines constrain the propeller position, can produce pulse amplitudes 4 to 5 times larger than at 25 percent, putting aft accommodation above the ISO 6954 upper limit with a standard propeller design.

Wave-induced hull vibration: whipping and springing

Beyond propeller and engine excitation, the hull girder is excited by wave forces.

Whipping is a transient vibration excited by wave impact (slamming) at the bow. A single bow slam injects energy into the hull girder across a range of frequencies; the hull responds in its free-free natural modes (2-node vertical, 3-node vertical) and the vibration decays over several cycles. Whipping matters for fatigue: the superimposed dynamic stress can be 20 to 40 percent of the static wave-bending stress at the slam event. The IACS Common Structural Rules for Bulk Carriers and Tankers (CSR-BC&OT, 2024 edition) requires a whipping fatigue check at structural details in the midship region.

Springing is the steady-state resonant vibration of the hull girder excited by the second-order wave force. For vessels with a 2-node natural frequency below 1 Hz (ULCCs, large bulk carriers, and ore carriers above 300 m), the wave encounter frequency at certain sea states can coincide with the hull natural frequency. The hull then vibrates continuously at its natural frequency, superimposed on the rigid-body wave-induced motion. Springing is most significant in head or bow seas at moderate to high Beaufort numbers. It has been observed on ore carriers of the Valemax class (400,000 DWT) and several Capesize vessels. The practical consequence is additional fatigue cycles at the natural frequency that are not captured by standard spectral fatigue analysis.

These topics connect to the seakeeping domain; see seakeeping and ship motions in waves for the wave-loading side of the problem.

The difference from propulsion-shaft torsional vibration

This article focuses on hull structural vibration: the elastic deformation of the ship itself, excited primarily at the propeller blade rate and engine free forces. Torsional vibration of the propulsion shaft train is a different problem with a different governing framework.

Shaft torsional vibration is the twisting oscillation of the coupled system formed by the engine crankshaft, intermediate shaft, propeller shaft, and propeller. It is governed by the engine’s cyclic firing torque and the torsional stiffnesses and inertias of the shaft components. IACS UR M68 (formerly M37) sets the allowable torsional stress and requires calculation of critical speeds for every installation. The shaft torsional critical speed calculator implements the UR M68 checks. Full treatment of torsional vibration analysis, including the barred-speed range, the detuning vibration damper, and the measurement protocol at sea trials, is in engine torsional vibration analysis.

The connection between the two: torsional vibration of the shaft produces torque fluctuations that are transmitted to the propeller and thence to the hull as bearing forces at the same blade-rate and engine-harmonic frequencies. A poorly tuned shaft with a critical speed near the blade rate will amplify the torque fluctuation and increase the hull excitation force, worsening hull vibration. This coupling means that solving a hull vibration problem without also checking the shaft torsional behaviour can lead to incomplete remedies.

Finite element analysis in vibration design

Modern vibration design uses finite element (FE) models at three levels of fidelity:

Global hull model (all-tanks or dry hull): a beam or grillage model that captures the mass distribution and stiffness of the full ship. Used to find the natural frequencies and mode shapes of the hull girder (2-node vertical, horizontal, torsional). Build time: days for an experienced analyst. Accuracy: 10 to 20 percent on natural frequencies, acceptable for early detuning checks.

Structural model of the aft ship: a full 3D plate-and-stiffener FE model of the stern structure, engine room, and aft accommodation, typically extending from the propeller aperture to the engine room forward bulkhead. This model resolves local panel natural frequencies and the global stern mode. Loads applied are the propeller pressure pulses from CFD or empirical sources. Build time: weeks. Accuracy: 5 to 10 percent on local frequencies when calibrated against measurement.

Accommodation superstructure model: a detailed FE model of the deckhouse structure, with plating, stiffeners, and cabin bulkheads. Used to find the modes that could be excited by engine harmonics or superstructure natural frequencies. Applied forces are the engine free forces and moments from the maker’s data sheet. Build time: days to weeks, depending on the model.

The principal commercial software packages used in the marine sector are NASTRAN (MSC Software / Hexagon), ANSYS Mechanical, Abaqus (Dassault Systemes), and the class-society tools DNV Sesam Hull (formerly NAUTICUS Hull) and NAPA Steel. Classification society plan approval review assesses the vibration FE model and the forced-response analysis for newbuildings requesting comfort notations.

Elastic mounts, floating floors, and isolation

Vibration isolation uses resilient elements to break the mechanical transmission path between source and receiver.

Resilient engine mounts (anti-vibration mounts, AVMs) are the standard approach for four-stroke medium-speed engines. A set of rubber or steel-spring mounts supports the engine at multiple points. The mounted system has six rigid-body natural frequencies (three translational, three rotational); all six must be below the firing frequency by at least a factor of 1.7 to maintain isolation rather than amplification. DNV COMF rules require documentation of the isolation efficiency across the operating speed range.

Double-resilient mounting (engine on AVMs on a floating steel subframe, the subframe on secondary mounts) provides a second attenuation stage and is used on cruise ships and naval vessels. The second stage reduces structure-borne noise and vibration at medium frequencies (100 to 2,000 Hz) by a further 10 to 20 dB compared to a single-stage mount.

Floating floors in accommodation spaces are a receiver-side treatment. A screed or prefabricated floor panel is installed on resilient pads over the structural deck. The isolation begins at the natural frequency of the floor-on-pad system (typically 8 to 15 Hz) and increases with frequency above that. Floating floors are effective for noise and high-frequency vibration but do not attenuate the low-frequency hull-girder vibration below 5 Hz.

Flexible pipe connections between the engine and ship piping (exhaust, sea water) prevent vibration from bypassing the AVMs through rigid pipe connections.

Vibration in specific vessel types

Different vessel categories encounter distinct vibration challenges.

Container ships: open-deck torsional modes couple with horizontal bending at natural frequencies that can coincide with the 5th and 6th engine harmonics. Torsional hull vibration is the dominant concern; see hull form design and the containership torsion context in engine torsional vibration analysis.

Cruise ships and ferries: diesel-electric propulsion with azimuth thrusters eliminates the direct diesel-mechanical shaft path, removing the dominant source of low-frequency vibration. The remaining sources are the generator sets (in flexible-mounted engine rooms), the azimuth thrusters (blade rate at pod RPM), and wave excitation. COMF notation is standard on new cruise ships; DNV COMF-V Level 1 (the most stringent, 40 mm/s$^2$ in passenger cabins) is the market expectation for upper-segment vessels.

Bulk carriers and tankers: these rely heavily on the single slow-speed two-stroke engine and fixed-pitch propeller. The typical vibration problem is aft-accommodation vibration at the first blade-rate. The Schlick formula is a useful first check; FE analysis is required for vessels under COMF contracts.

LNG carriers: the propulsion plant has changed materially over the past two decades, from steam turbine (almost no vibration issues) to two-stroke dual-fuel (MEGI, X-DF types) with similar vibration characteristics to conventional two-stroke vessels. Membrane cargo tanks are sensitive to global hull-girder deflection under sloshing and wave loads; vibration interacts with the structural integrity programme for the containment system.

Offshore vessels (PSVs, AHTSs, OSVs): dynamic positioning requires the vessel to be nearly stationary with high thruster activity. Azimuth and tunnel thrusters at varied RPMs can produce broadband vibration that is harder to control than fixed-RPM propulsion. The ISO 6954 measurement is taken at multiple DP operating conditions.

Regulatory context

ISO 6954:2000 is a voluntary standard; it becomes contractual when referenced in the building specification or class notation. MLC 2006 (ILO Maritime Labour Convention 2006, Title 3.1, Standard A3.1(6)(e)) requires that accommodation spaces be free from “excessive noise and vibration.” While MLC does not cite ISO 6954 directly, the ILO Guidelines (2009) reference ISO 6954 values as the benchmark for compliance, giving the standard indirect regulatory force on MLC-certified vessels. Flag-state PSC inspectors have cited accommodation vibration above the ISO 6954 upper limit as a deficiency under MLC.

SOLAS regulation II-1/3-1 (passenger and cargo ship construction) requires adequate structural design but does not address vibration levels numerically. The vibration control obligations on commercial ships derive primarily from class rules, building contracts, and MLC.

Limitations

The methods described in this article carry several practical limitations that the practitioner should keep in mind.

The Schlick formula for hull natural frequency is an empirical fit to 19th- and 20th-century steam vessels. It carries a typical uncertainty of 20 to 30 percent. It should not be used as the sole basis for a resonance ruling on a newbuild; it is appropriate only as a preliminary check to decide whether detailed FE analysis is needed.

Propeller pressure pulse prediction by empirical methods (Holden, Gawn) assumes a uniform, non-cavitating flow. Real wake fields behind ships are highly non-uniform and speed-dependent; cavitation onset changes the excitation character entirely. Errors of a factor of 2 to 5 in predicted pressure pulse are not unusual when comparing hand calculations to measured values on delivered ships.

FE model accuracy depends on correctly representing structural mass (including outfit), fluid-structure interaction (added mass from surrounding water reduces natural frequencies by 15 to 25 percent at deep submergence), and boundary conditions at the sea surface. FE models without added mass will over-predict natural frequencies. The virtual mass factor (Prohaska factor) correction is standard in marine structural analysis but requires careful application at different structural scales.

ISO 6954 limits are not universally applicable to all vessel types. The standard’s scope covers “passenger and merchant ships on sea voyages.” Offshore vessels, naval ships, and workboats may use different criteria. ISO 6954 also does not address whole-body vibration in the occupational health sense (ISO 2631-1); that standard applies to seated workers exposed continuously during working hours.

Comfort notation surveys are conducted under specific trial conditions (Beaufort 3 or lower, specified speed range). In-service conditions may differ, and some vessels that pass the trial show elevated vibration in service when operating with different loading, fouled hull, or different propeller pitch settings. The notation is not a guarantee of continuous compliance in all sea states.

See also

Related calculators:

• DNV VC notation checker
• ABS VSA notation checker
• ISO 10816 vibration severity zone
• Shaft torsional critical speed
• Sway-roll coupling natural frequency
• Open-deck torsion: containership

Related wiki articles:

• Engine torsional vibration analysis
• Marine engine room automation and monitoring
• Marine propeller
• Propeller theory
• Hull strength and longitudinal bending
• Marine diesel engine
• Two-stroke marine diesel engine fundamentals
• Seakeeping
• Ship motions in waves
• Hull form design
• Classification society
• Naval architecture coefficients