Marine Propulsion Shafting and Stern Tube

Marine propulsion shafting is the mechanical spine connecting the main engine to the propeller. On a typical 300-metre bulk carrier, the shafting line runs 40 to 65 metres from the engine output flange to the propeller hub, passing through the engine room, a watertight shaft tunnel, and finally the stern tube where the shaft exits the hull underwater. The shafting must transmit full continuous power (up to 27 MW on a modern large bulk carrier) without fatigue failure, support the propeller’s several-tonne mass through bearing reactions, and seal the underwater hull penetration against seawater ingress and lubricant release to the sea.

The IACS Shaft Diameter Calculator applies the IACS UR M68 formula discussed in this article. Readers interested in the engine coupling and alignment side of this train should also see Engine Alignment and Bedplate Flexure; for the power reduction gearing on medium-speed-engine ships, see Marine Reduction Gears; and for the propeller itself, see Marine Propeller.

The shafting train: components and their roles

A conventional single-screw installation consists of four functional elements arranged in series: the thrust shaft and thrust bearing, the intermediate shaft or shafts with their plummer blocks, the propeller (tail) shaft, and the stern tube housing the aft and forward bearings.

Thrust shaft and thrust bearing

The thrust shaft carries the axial collar that transfers propeller thrust to the ship’s hull. On every modern slow-speed two-stroke main engine (MAN B&W, WinGD), the thrust collar and the tilting-pad thrust bearing are integrated inside the engine itself, in the lower crankcase, so a separate thrust shaft does not exist. The engine bedplate transmits the thrust directly to the double-bottom structure. On some medium-speed installations coupled through a reduction gear, a standalone thrust bearing sits in the shafting between the gearbox and the first intermediate shaft; the Mitchell tilting-pad design is standard, with each pad pivoting on a balance ring to maintain a hydrodynamic oil film at loads of 200 kN to over 1,000 kN.

Thrust bearing pad temperatures are monitored by embedded thermocouples, typically one per pad on large bearings. DNV class rules (Part 4 Ch 4 Sec 2) require an alarm at 85 degrees C and a slowdown trigger at 95 degrees C for lubricated tilting-pad thrust bearings.

Intermediate shafts and plummer blocks

Intermediate shafts span the length of the shaft tunnel between the engine room aft bulkhead and the stern tube. Most single-screw commercial vessels carry one to three intermediate shaft sections, each 5 to 12 metres long, with forged integral flanges at both ends bolted together. Each section rests on a plummer block (also called a steady bearing or line-shaft bearing): a split journal bearing housing supported on a raised seating welded to the tunnel structure. Plummer block bearings carry only radial loads; they react shaft self-weight and the small bending from misalignment. Bearing type is typically white metal (babbitt) lined, lubricated by a splash ring drawing from a self-contained oil sump. Typical plummer block span is 5 to 8 shaft diameters; exceeding 10 shaft diameters between supports makes the shaft prone to lateral (whirling) resonance within the operating speed range.

Alignment of the intermediate shaft directly affects the load distribution on the aft stern tube bearing. A shaft train that slopes toward the propeller can lift the aft stern tube bearing above its designed load share, accelerating wear. IACS UR M68 and the individual society rules therefore specify allowable reaction loads at each bearing as part of the alignment calculation, not just dimensional tolerances on gap-and-sag measurements. The detailed treatment is in Engine Alignment and Bedplate Flexure.

Thrust bearing: Michell/Kingsbury tilting-pad design and load path

On standalone thrust bearing installations (medium-speed-engine ships with reduction gears, some older two-stroke plants with a separate thrust block forward of the engine), the Michell tilting-pad (Kingsbury in North American terminology) is the universal design. The bearing consists of a collar forged integral with the thrust shaft, rotating between an ahead-side pad ring and an astern-side pad ring. Each ring typically carries six to eight pads. Each pad is free to tilt on a pivoting land or balance ring, and this tilting generates a converging oil wedge that supports load hydrodynamically at very high efficiency. The coefficient of friction for a well-designed tilting-pad thrust bearing is 0.001 to 0.003, versus 0.005 to 0.010 for a flat-collar (Michell-plate) design.

The thrust force from the propeller in the ahead condition at full power can reach 1,000 to 2,500 kN on a 15,000 to 27,000 kW installation. This entire axial load travels through the thrust collar, through the ahead pads, through the pad housing, and into the chocks welded to the bedplate, which in turn connects to the double-bottom stiffener structure directly below. The transmission path is effectively the entire mid-ship girder taking compression. Astern pads take loads during crash stops and reversing manoeuvres; the astern load is typically 40 to 60% of the ahead design value for fixed-pitch propeller ships.

Thermocouple monitoring is standard on large thrust bearings. DNV class rules (Part 4 Ch 4 Sec 2, 2024) mandate one thermocouple per pad (or per pad pair on compact designs) on installations above 2,250 kW, with bridge alarm at 85 degrees C and automatic slow-down at 95 degrees C. The oil film in the ahead-pad tilting zone runs at 40 to 55 degrees C under normal full-load conditions; a reading above 70 degrees C at sustained full power is a signal of either insufficient oil flow (filter clogging, pump failure) or pad surface damage. LR Rules for Ships (Part 5 Ch 7, 2024) require the same alarm thresholds and additionally specify that the oil supply pressure must be interlocked with the engine start-permit circuit so the main engine cannot be started without confirmed lubrication flow to the thrust bearing.

Thrust bearing axial clearance (the free travel of the shaft before thrust load is taken up by pads) is set at 0.20 to 0.35 mm total on large installations. This clearance is measured during dry-dock with a dial gauge against the thrust collar face. Exceeding 0.5 mm total clearance, or any measurable pad wipe (babbitt transfer to the collar), triggers pad renewal before the bearing resumes service.

Propeller shaft (tailshaft)

The propeller shaft is the aft-most shaft section, running inside the stern tube from the stern tube forward seal gland to the propeller hub. It is also called the tailshaft in survey parlance. It is dimensioned larger than the intermediate shaft because it carries the combined effect of torque transmission, propeller bending (the propeller’s weight acting at the overhang), and the asymmetric inflow loading that creates a vertical bending moment at the aft stern tube bearing.

The propeller fits the tailshaft on a machined taper, with interference fit achieved by hydraulic oil injection between the taper bore and the shaft surface (the keyless hydraulic (Pilgrim) mounting method). Keyless fitting distributes contact stress across the full taper interface rather than concentrating it at a key; this eliminates a stress-concentration site prone to fretting fatigue and is now universal on commercial newbuilds. The taper ratio is typically 1:15 (radial) per IACS and class norms. Withdrawal is reversed: oil injected through the same annular groove forces the hub off the taper.

Propeller-to-shaft connection: keyless taper, oil-injection fitting, and contact pressure

The keyless taper fit replaced the older keyed arrangement because a keyway is a stress concentrator at precisely the highest-torsion location on the shaft. On a keyed fit, the key transmits the full torque through a small rectangular contact area; fretting fatigue cracks initiate at keyway ends and have caused propeller shaft failures in service. IACS UR M68 (2023) and the major classification societies no longer accept new keyed propeller fits on commercial vessels above 750 kW.

On the keyless fit, the hub taper bore is lapped onto the shaft taper to achieve at least 70% contact area (confirmed by blue-ink transfer test). The interference between bore and shaft is defined by the push-up length: the axial travel of the hub along the taper from the hand-tight position to the final seated position. Per SKF and Wartsila installation manuals, the target push-up length for a 440 mm shaft diameter (shaft taper 1:15) is typically 3.5 to 5.0 mm, which develops a radial contact pressure of 30 to 55 MPa across the taper interface. This contact pressure, combined with the friction coefficient of the metal-to-metal interface (approximately 0.12 to 0.15 dry), generates the frictional grip that transmits both torque and axial thrust.

The oil-injection procedure (Pilgrim nut method, or the equivalent SKF hydraulic nut) applies high-pressure oil (250 to 350 bar) through a drilling in the forward end of the shaft taper, feeding an annular groove machined in the hub bore. This film reduces friction to near zero during the push-up stroke, allowing the hub to travel axially without galling the taper surfaces. A hydraulic nut or Pilgrim ring at the aft face of the hub provides the push-up force; the stroke is measured with a calibrated gauge and the target push-up length is specified in the installation drawing to ±0.1 mm. Once the hub reaches the target position, oil injection ceases, the oil film dissipates, and the hub locks onto the shaft through metal-to-metal contact. The forward end of the shaft has a threaded nut and tab washer to prevent aft travel in service.

Final verification of the fit uses a torque-moment check: with the class surveyor present, the propeller is rated by a torque wrench to confirm that the theoretical breakout torque (calculated from the contact pressure, taper geometry, and friction coefficient) is not less than 1.5 times the maximum shaft torque. DNV Rules (Part 4 Ch 4 Sec 1, 2024) specify this safety factor of 1.5 as the minimum for the keyless fit to be accepted without additional mechanical retention. ABS Rules for Building and Classing Steel Vessels (Part 4 Ch 3, 2024) specify the same factor.

Shaft material and IACS UR M68 diameter rule

Material specification under IACS UR M68

IACS UR M68 (2023 edition) specifies that propulsion shafts must be made from forgings approved under the same UR. Two grade families are defined: carbon steel grades (SC3, SC4 series) and alloy steel grades (SA3, SA4 series). Minimum tensile strength ranges from 400 N/mm2 (SC3) to 700 N/mm2 (alloy grades). Most intermediate and propeller shafts on ocean-going commercial ships use SC4-grade carbon steel (UTS 410 to 490 N/mm2), whose combination of machinability and fatigue life at 10^7 cycles meets the design life without the cost premium of alloy steel. Ice-class vessels and offshore vessels subject to heavy shock loads shift to alloy steel, which offers superior notch toughness below zero degrees C.

Chemical composition limits under M68 include maximum sulfur of 0.035% and phosphorus of 0.035%, with lower limits mandated on cleanliness-critical propeller shafts. Ultrasonic testing to UR W2/W7 and magnetic particle or liquid penetrant testing after final machining is mandatory. The class society surveys the material at the forge and again at the shipyard before installation.

The IACS UR M68 shaft diameter formula

The formula gives the minimum outer diameter for a solid shaft. The factor $k$ encodes the shaft position and fitting arrangement: IACS UR M68 Table 1 sets $k = 1.00$ for an intermediate shaft on a rigid coupling, rising to $k = 1.22$ for the propeller shaft outboard of the aft stern tube bearing (where bending from propeller weight and asymmetric inflow adds to torsion). For a hollow shaft with bore diameter $d_i$, the solid-equivalent diameter is corrected by $(1 - (d_i/d)^4)^{-1/3}$, which raises the required outer diameter to deliver the same torsional section modulus as the solid design. A typical bore ratio of $d_i/d = 0.50$ requires the outer diameter to be about 4.5% larger than the solid equivalent.

Worked numbers (carbon steel SC4, UTS = 450 N/mm2): a 15,000 kW main engine turning at 105 rpm gives $P \cdot k / N = 15000 \times 1.00 / 105 = 142.9$ for the intermediate shaft and $15000 \times 1.22 / 105 = 174.3$ for the propeller shaft. Applying the formula, intermediate shaft minimum diameter is approximately 430 mm; propeller shaft minimum is approximately 460 mm. These round up to standard forging sizes in 10 mm increments. On a large container ship (60,000 kW, 94 rpm), propeller shaft diameter approaches 700 mm.

Surface finish on bearing contact zones must reach Ra 0.4 to 0.8 micrometres after grinding, per DNV and IACS requirements, to ensure correct bedding in white-metal bearings.

Stern tube: construction and bearing arrangement

The stern tube is a steel cylinder welded into the aft hull structure, typically running from the aft peak bulkhead aft through the stern frame and terminating just aft of the propeller boss aperture. Its wall thickness is governed by hull structural requirements (the stern tube is load-bearing as well as functional) and by the bearing housing bore dimensions. Two bearing seatings are machined at each end: the aft (outboard) seating, which sits underwater near the propeller, and the forward (inboard) seating, which sits in the engine room aft void space.

Aft stern tube bearing

The aft bearing carries the dominant share of propeller shaft loads. On a typical 15,000 kW installation, the propeller weighs 25 to 45 tonnes in air (roughly 22 to 40 tonnes net in water), and the bending moment from this mass acting at the propeller hub overhang puts a reaction load of 80 to 150 kN on the aft bearing under normal trim and draught conditions. Propeller inflow asymmetry (the propeller rotating in a wake field that varies with depth from the hull) adds a cyclic bending component at blade-rate frequency; this is the primary source of fatigue loading on the propeller shaft.

Oil-lubricated white-metal bearings lined with tin-base babbitt (ISO 4381 grade SnSb12Cu6Pb or similar) are the dominant aft bearing type on large commercial vessels. The bearing housing is a steel sleeve bored to accept the babbitt liner; oil is supplied through a circumferential groove cut into the bearing bore and connected to the stern tube oil system. Typical operating clearance is 0.0015 to 0.002 times shaft diameter, giving 0.65 to 0.9 mm on a 440 mm shaft. Load capacity is 0.6 to 1.0 MPa mean pressure on the projected area. Bearing length-to-diameter ratio for the aft stern tube bearing is typically 3.0 to 4.5, meaning the aft bearing is 1.3 to 2.0 metres long on a 440 mm shaft.

Wear-down is measured with a bridge (wear-down) gauge bridging the stern tube opening. The measurement records the vertical drop of the shaft centerline from a datum on the stern tube structure. Class rules set the wear-down limit at the point where the shaft would contact the lower shell of the bearing housing; this is typically 0.50 to 0.75 mm for white-metal lined aft bearings. Exceeding the limit at the in-port check mandates dry-docking for bearing renewal before the next voyage.

Forward stern tube bearing

The forward bearing is shorter (L/D ratio 1.0 to 2.0) and carries a much smaller load because the shaft is supported at greater bearing spacing from the propeller. Its primary function is to maintain shaft centerline at the forward end of the stern tube, limiting angular deflection that would load the aft seal gland eccentrically. White-metal or cutless-rubber designs are common; some modern installations use composite polymer at the forward position with white-metal aft.

Oil-lubricated versus water-lubricated stern tube bearings

The following comparison summarizes the principal differences:

Water-lubricated bearings eliminate the oil-to-sea interface entirely, which removes both the VGP EAL compliance burden and the risk of oil pollution from seal failure. Thordon Bearings Inc. reports bearing installations on vessels up to 45,000 DWT; above that displacement, the load per projected area begins to approach the limit for their SXL compound in aft stern tube service. Vesconite Hillside and Orkot TLMB materials have comparable published load limits. For vessels where the aft bearing load exceeds the polymer limit, a hybrid arrangement (polymer forward, white-metal aft) is sometimes used.

Stern tube seals

The stern tube is sealed at both ends: the aft (outboard) seal prevents seawater from entering the stern tube and oil from leaking to sea; the forward (inboard) seal prevents oil from draining into the engine room.

Aft (outboard) seal

The dominant aft seal design for oil-lubricated stern tubes is the multi-lip elastomer ring seal, supplied by Wartsila (formerly Lips and Tideman), Simplex-Turbulo (now part of SKF), Kemel, and EagleBurgmann. A standard Simplex S-type aft seal carries four lip rings (each an annular elastomer element energized by a circumferential garter spring) arranged on a brass or stainless carrier ring bolted to the stern tube shell:

• Lip 1 (outermost): faces seawater; keeps water out. This lip is in contact with seawater pressure and runs with a thin oil film fed back from the stern tube circuit.
• Lip 2: backs up Lip 1; also water-side.
• Lip 3: faces aft toward the oil side; the space between Lips 2 and 3 is the “drain space” vented to atmosphere (or a low-pressure monitoring point) so any breach is detected before cross-contamination occurs.
• Lip 4 (innermost): primary oil-side seal, retaining stern tube oil from leaking aft.

The split between water-side and oil-side lips at the drain space is the key diagnostic point: a positive oil pressure (detected by a float switch or level alarm in the drain tank) signals Lip 4 wear; a water presence in the drain signals Lip 1 or 2 wear. Both conditions warrant seal replacement at the next available port call or dry-dock depending on severity.

Shaft liner (also called shaft sleeve or rope guard) covers the tailshaft in way of the seal lips. Liners are chrome-plated or Stellite-coated steel sleeves shrunk onto the shaft. Lip ring wear manifests as grooves worn into the liner surface; when groove depth exceeds 1 mm, the liner must be replaced even if the lips themselves are within tolerance, because the groove defeats the sealing action of a new lip set.

Stern tube oil lubrication circuit and the gravity/header tank

The stern tube oil system on an oil-lubricated installation is a static head circuit, not a pressure-fed pump circuit. A header (gravity) tank is mounted above the waterline in the shaft tunnel or engine room, typically 1.5 to 3.5 metres above the aft stern tube bearing centreline. The height of this tank relative to the stern tube aft seal determines the static oil pressure at the aft seal: the pressure, in bar, is approximately 0.1 times the vertical head in metres. For a header tank 2.0 metres above the aft seal, the oil pressure at the seal is 0.20 bar. The key design condition is that this static head must exceed the local seawater pressure at the aft seal face at all draughts.

For a vessel at maximum draught of 14 metres, the seawater pressure at the aft seal (typically located about 3 metres above the keel) is about 1.1 bar. The header tank, at 2 metres above the bearing centreline and therefore roughly 5 metres above the keel, provides only 0.5 bar oil pressure. This means the aft seal always operates with seawater pressure exceeding oil pressure at the outboard lip face, and the outermost water-side lip (Lip 1 in the Simplex four-lip arrangement) must actively keep seawater out against that differential. The oil-side lips (Lips 3 and 4) must prevent oil leaking to sea against the head pressure; they do so because the net oil pressure at the drain space (between the water-side and oil-side lips) is lower than seawater pressure, so the tendency is always toward water ingressing rather than oil escaping if the outer lips fail. This is the fundamental asymmetry of aft stern tube seal design.

The header tank volume is sized to provide 48 to 72 hours of autonomous operation at the expected steady-state consumption rate (0.1 to 0.3 L/day on a new seal, rising to 1 to 3 L/day on worn lips). A low-level alarm in the header tank is the primary early-warning instrument for seal wear; a sudden drop in level (more than 5 L/day) is cause for port inspection. Oil consumption records are a mandatory input for IACS UR M75 condition-based survey.

The circuit also serves oil cooling: oil returning from the aft bearing zone carries heat generated by bearing friction and mechanical losses; this oil rises through the header tank, where it equilibrates with ambient temperature, before returning by static head. No pump is needed for normal operation, which is why stern tube oil system failure only occurs on seal or piping damage, not pump trips.

Air Guard (pressurized air seal)

Several classification societies and seal manufacturers offer an air-barrier (Air Guard) system in which the drain space between the water-side and oil-side lips is pressurized with clean air or nitrogen at 0.1 to 0.3 bar above stern tube oil pressure. This pressure differential prevents both water ingress (the air overcomes any head differential from sea pressure) and oil-to-sea leakage (the air forms a positive barrier on the oil side). DNV and LR now accept the Air Guard arrangement as equivalent to the standard drain-space monitoring for condition-based survey purposes.

The Air Guard arrangement also satisfies the US EPA VGP 2013 zero-discharge requirement in a different way from switching to a water-lubricated system. Rather than removing the oil entirely, the air barrier physically separates the oil from the seawater at the drain space, so any oil that bypasses the oil-side lip exhausts into the pressurized air cavity rather than to sea. The air supply is provided by the ship’s instrument air system at 6 to 7 bar, reduced by a pressure regulator to the controlled 0.1 to 0.3 bar differential. An oil mist detector or drain tank float switch monitors the cavity for any oil breakthrough that would indicate oil-side lip failure. When fitted with Air Guard, the Wartsila and Kemel seal arrangements are documented as achieving zero measured oil-to-sea discharge in EPA-approved third-party tests, which is the condition required by VGP 2013 for an oil-lubricated stern tube system to satisfy the EAL alternative.

Rope guard and the aft seal protection zone

The rope guard (also called the propeller shaft rope guard or zinc guard ring) is a conical or cylindrical steel deflector bolted to the shaft just aft of the aft seal housing. Its function is to prevent ropes, fishing line, netting, and floating debris from wrapping around the shaft and being dragged into contact with the aft seal lips. Rope ingestion is a major cause of aft seal failure in coastal and port service: a single rope turn around the shaft at 90 rpm can exert a contact force of several kilonewtons on the seal carrier, tearing the lip ring from its housing within minutes. The rope guard is a mandatory fitting under most class societies’ stern tube construction requirements; DNV Rules (Part 4 Ch 4 Sec 1, 2024) and LR Rules (Part 5 Ch 7, 2024) both require it on all ocean-going propelled vessels.

The shaft liner (shaft sleeve) in way of the rope guard is inspected at every dry-dock. Chrome oxide or tungsten carbide thermal-spray coatings are sometimes applied to the liner surface in the rope-guard zone to resist abrasion from soft debris contact without the full Stellite or chrome-plating cost.

Forward (inboard) seal

The forward seal retains stern tube oil at the engine room end. It works against a much lower differential pressure (oil head only; typically 0.05 to 0.15 bar) and normally consists of two to three lips. The forward seal is the easier replacement of the pair: it is accessible with the shaft in situ if the intermediate shaft coupling is disconnected.

Environmentally acceptable lubricants and VGP compliance

The US EPA Vessel General Permit (VGP 2013), section 2.2.9, requires all vessels over 79 feet in length operating in US waters to use EALs at all oil-to-sea interfaces, including the stern tube, propeller hub seals, thruster seals, and stabilizer fin seals. An EAL must be:

• Biodegradable: meeting ASTM D6731 (for ultimate biodegradability above 60%) or OECD 301B/306 criteria.
• Minimally toxic: failing to trigger acute toxicity in 96-hour fish or invertebrate bioassay at concentrations below 1,000 ppm.
• Non-bioaccumulating: log Kow (octanol-water partition coefficient) below 3.

Synthetic ester-based oils (ISO VG 100 to 150 for aft stern tubes on low-speed shafts) are the most widely used EAL type. Polyalkylene glycols (PAGs) meet all three criteria but are not compatible with mineral oil residues, requiring a thorough system flush before commissioning. EALs are 2.5 to 4 times the cost of equivalent mineral oils; the trade-off is the elimination of potential clean-water-act violation fines, which the EPA has assessed at up to $25,000 per day per violation on oil-to-sea incidents.

Several countries’ coastal legislation (Canada, Australia, parts of the EU under MARPOL Annex I Regional provisions) apply EAL requirements equivalent to or stricter than VGP 2013 in their coastal waters, so vessels in international trade find EAL adoption economically rational regardless of US trading pattern.

Alignment loads and the bearing load calculation

Shaft alignment is not simply a matter of keeping a straight line. The propeller shaft droops under its own weight and propeller mass; the engine sits on a flexible bedplate and moves at operating temperature; the hull girder hoggs and sags with wave loading and ballast condition. A correct alignment design specifies target reaction loads at each bearing (including the aft stern tube bearing, the forward stern tube bearing, and each plummer block) such that no bearing is overloaded or lifted off under the design envelope of conditions.

IACS UR M68 (Annex B, Shaft Alignment) requires the designer to calculate the influence numbers matrix (the change in bearing reaction per unit settlement at each support) and to demonstrate that the full operating envelope stays within allowable bearing load and shaft stress limits. The “reasonable alignment” check compares calculated bearing loads against class-permitted minimum and maximum values.

For the aft stern tube bearing, the minimum reaction load must be positive (the bearing cannot be lifted) under all conditions, including the light-ship, worst-trim condition with the hot engine at maximum thermal expansion. Insufficient aft bearing load leads to intermittent metal contact, accelerated wear, and fretting fatigue at the forward stern tube bearing and shaft at that location. LR (Rules and Regulations for the Classification of Ships, Part 5 Ch 7) and DNV (Part 4 Ch 4 Sec 2) both specify minimum aft bearing reaction loads of approximately 20% of propeller weight in water as the floor condition; actual class requirements vary slightly by vessel type and shaft span.

The dedicated Engine Shaft Alignment Sag Calculator computes gap-and-sag measurements at each coupling for a given bearing load solution. The Shaft Torsional Critical Speed Calculator addresses the torsional resonance side of the same shafting design; both are complementary to the alignment work described here. For the torsional vibration analysis itself, see Engine Torsional Vibration Analysis.

Tailshaft surveys and condition-based extended intervals

Traditional 5-year withdrawal survey

Class society rules historically required propeller shaft (tailshaft) withdrawal from the stern tube at every Special Periodical Survey, i.e., every five years. Withdrawal means uncoupling the intermediate shaft, removing the propeller, extracting the tailshaft aft through the stern tube aperture (or forward into the engine room on shorter installations), NDT examination of the aft taper and oil-contact zone, white-metal bearing inspection and measurement, seal renewal, and reassembly. On a typical Panamax bulker, tailshaft withdrawal takes 7 to 12 days in dry-dock, a significant cost and schedule burden.

IACS UR M75 condition-based extended survey

IACS adopted UR M75 in 2014, revised in 2021, providing a framework for extending the tailshaft survey interval beyond five years when condition monitoring evidence satisfies class surveyors that the shaft remains in acceptable condition. Under UR M75, an extended survey to 7.5 years may be approved when all the following criteria are met:

1. Stern tube oil has been analysed at intervals not exceeding six months, with results trending within acceptable limits for water content (below 0.2% by volume in most class guidance), particle count, and wear-metal concentration (iron and non-ferrous metals below class-specific limits).
2. The lubricant is confirmed as an EAL compliant with VGP criteria, which is taken as evidence that corrosion and contamination risk in the bearing zone is lower than with mineral oil.
3. Stern tube seal condition is monitored by continuous or periodic drain-space pressure monitoring and oil consumption records.
4. No abnormal wear-down readings, no bearing temperature excursions, no vibration anomalies, and no seal failure events in the preceding five-year period.
5. The arrangement was presented to the class society at the Special Survey and accepted before the first 5-year period concluded.

When all five conditions are satisfied, the tailshaft remains in situ. The 7.5-year point triggers a further evaluation; if the above monitoring conditions continue to be met, some class societies accept a second extension to 10 years on a case-by-case basis. DNV and LR both publish detailed condition monitoring checklists for their implementation of UR M75.

Oil sampling procedure for UR M75 compliance: samples are drawn from the stern tube drain cock (not the sump suction) using vacuum extraction equipment to minimize air contamination, placed in clean glass vials, and dispatched to an accredited laboratory within 48 hours of sampling. The laboratory reports water content (Karl Fischer titration), viscosity at 40 degrees C (ASTM D445), acid number (ASTM D664), iron (ICP-OES), copper, tin, and lead (ICP-OES), and particle count (ISO 4406). The owner compares results against baseline and class threshold values and files the report with the class surveyor.

Intermediate inspection

At the 2.5-year intermediate hull survey, the stern tube is inspected in situ without withdrawal. The inspection includes wear-down measurement, stern tube seal inspection (lip condition assessment through the drain space), oil consumption records review, oil analysis review, and operational logs review. If wear-down is within 60% of the allowable maximum, no mandatory action is taken. If wear-down is between 60% and 100%, the class surveyor may recommend monitoring at 12-month intervals. If the allowable maximum is reached, the ship is required to dry-dock for bearing renewal before the next voyage.

Shafting vibration modes and the torsional critical speed

The shaft train has three principal vibration modes: torsional, lateral (whirling), and axial.

Torsional vibration is excited by the cyclic torque variations of the main engine. Each power cylinder fires in sequence; the torque curve between firings has harmonics at multiples of firing frequency. A torsional critical speed (TCS) occurs when a harmonic of firing frequency matches a natural torsional frequency of the shaftline-propeller-engine system. At a TCS, shaft twist angles amplify and shear stress in the shaft can reach values far above the mean torsional stress from power transmission. Class rules require that the TCS is either outside the normal operating speed range (with a 10 rpm margin from MCR) or, if it falls within range, that the peak torsional shear stress is demonstrably below the allowable value specified in the rules. The allowable shear stress for torsional vibration on SC4 shaft steel in continuous service is approximately 18 to 21 N/mm2 (class-dependent). The Shaft Torsional Critical Speed Calculator computes the natural frequencies; for the full analysis methodology and the engine-side torque harmonics, see Engine Torsional Vibration Analysis.

Lateral (whirling) critical speeds are the natural frequencies of the shaft treated as a beam on spring supports at each bearing. The first lateral (whirling) critical speed for a uniform shaft section of length $L$ and diameter $d$ between rigid end supports scales as approximately:

where $E$ is Young’s modulus (206 GPa for steel), $I = \pi d^4 / 64$ is second moment of area, $\rho$ is density (7,850 kg/m3), and $A = \pi d^2 / 4$ is cross-section area. Substituting for a 430 mm solid shaft with $L = 8.0$ m gives a first whirling critical speed around 1,850 rpm, which is far above a typical slow-speed shaft running at 90 to 120 rpm. The risk rises on longer spans: the same shaft at $L = 12.0$ m drops to around 820 rpm, which is still comfortably above normal operating range but inside the overspeed range for a runaway.

Class rules set the maximum permitted bearing span not by a formula but by requiring the designer to demonstrate, in the alignment calculation submission, that no lateral natural frequency falls within the service speed range plus a 10% margin. DNV Rules (Part 4 Ch 4 Sec 2, 2024) and LR Rules (Part 5 Ch 7, 2024) both require this demonstration for intermediate shaft spans above 8 metres. For most slow-speed installations, intermediate shaft sections of up to 10 metres between plummer blocks are accepted without analysis when shaft diameter is in the 400 to 600 mm range; spans above 12 metres at those diameters require a formal whirling critical speed calculation submitted to the class surveyor before commissioning.

The propeller shaft presents a different whirling risk because the propeller mass at the overhang acts as a concentrated end mass, lowering the first whirling frequency compared to a simple unsupported span. The aft stern tube bearing and the forward stern tube bearing form the two end supports; the unsupported propeller overhang (the distance from the aft bearing to the propeller hub centre) adds a cantilever term. For a 44 tonne propeller hub at 0.8 m overhang on a 440 mm shaft, the cantilever reduces the first lateral critical speed by approximately 15 to 20% relative to the pinned-end calculation. This is why the aft stern tube bearing length-to-diameter ratio of 3.0 to 4.5 (section on aft stern tube bearing above) is not arbitrary: the long bearing stiffens the aft support sufficiently to keep the lateral critical speed above the operating range.

Axial vibration from propeller thrust pulsations is normally small for fixed-pitch propellers (the dominant force is at blade-rate, i.e., blade number times shaft rpm). On installations with relatively soft thrust bearing mountings (for noise isolation purposes, used on some passenger ships), axial resonance can cause significant fore-aft vibration of the engine itself and requires a detuner or active control damper.

Controllable pitch propellers and hollow shafts

Controllable pitch propeller (CPP) installations require the tailshaft and intermediate shafts to be hollow, because the hydraulic oil supply pipe and pitch feedback rod run through the bore of the entire shafting line from the servo valve unit in the engine room to the hydraulic cylinder in the propeller hub. The bore diameter is typically 100 to 200 mm depending on system flow requirements. This bore reduces shaft cross-section, which is accounted for in the hollow-shaft diameter correction factor in the IACS UR M68 formula. The CPP hub itself contains a hydraulic piston and link mechanism; for the hub hydraulic circuit, see the Controllable Pitch Propeller Hydraulic Hub calculator. Shaft power transmission through a reduction gearbox on CPP installations is addressed in Marine Reduction Gears.

Power transmission losses in the shaftline

The shaft does not deliver 100% of the engine brake power to the propeller. Losses occur at the thrust bearing (0.1 to 0.3% of shaft power), in the intermediate shaft plummer blocks (0.02 to 0.05% per bearing, negligible on most installations), and in the stern tube aft bearing (0.1 to 0.3%, higher at low shaft speeds when hydrodynamic film formation is marginal). Total shafting mechanical losses on a well-maintained installation are typically 0.5 to 1.0% of brake power, meaning the delivered power to the propeller (DHP) is 99.0 to 99.5% of shaft power (SHP). The Propeller Shaft SHP vs DHP Calculator converts between the two using class-recommended loss factors. Losses from gearboxes on medium-speed installations are substantially larger (1.5 to 3.0%) and are calculated separately in Tech Shaft Gearbox Losses.

When an auxiliary alternator (shaft generator) is driven from the main shafting, it draws power from the PTO flange at the engine end of the shaft; the electrical output is calculated by the Shaft Generator Calculator.

Cathodic protection and shaft earthing

The propeller shaft rotates in sea water, creating a potential for galvanic corrosion where the bronze or stainless propeller contacts the steel shaft. Class rules require sacrificial zinc anodes on the propeller hub and, on some installations, bonded anode rings on the stern tube. A shaft earthing brush (shaft grounding device) provides an electrical return path from the shaft to the hull, preventing shaft current build-up that would pit bearing surfaces. The full context of hull corrosion protection is in Marine Cathodic Protection and Hull Coatings.

Limitations

The following caveats define the scope of this article:

IACS UR M68 formula scope. The IACS UR M68 diameter formula produces the minimum compliant diameter for carbon or alloy steel shafts on standard propulsion plants. It does not cover: titanium or composite shafts (emerging on naval and specialist vessels), shafts subject to ice-class UR I loads (separate ice-class rules apply), or shafts on vessels with dynamic positioning in which thruster shaft sizing follows different load scenarios.

Water-lubricated bearing load data. Published specific load limits for Thordon SXL, Vesconite, and Orkot TLMB are taken from OEM data sheets as of 2023. These figures are confirmed under the manufacturer’s own laboratory conditions; actual in-service performance depends on water cleanliness, shaft surface finish, and operational speed range. Class society approval of water-lubricated bearings is required on a case-by-case basis for vessels above ~25,000 DWT at most societies.

VGP applicability. The VGP 2013 EAL requirement applies to vessels in US waters. The EPA issued a final rule in December 2022 reauthorizing the VGP through 2027. Non-US coastal jurisdictions have enacted equivalent requirements, but their exact thresholds vary; verify with flag state or local port authority before entry.

Condition-based survey (UR M75). Extended survey intervals are only available to ships that have continuously met all five qualifying criteria. A single oil-analysis exceedance, seal failure, or classification note on abnormal wear-down disqualifies the ship from the extension; the 5-year withdrawal survey becomes mandatory at the next scheduled interval.

Alignment loads. The alignment load calculation methodology summarized here follows IACS UR M68 Annex B. Each class society (DNV, LR, ABS, ClassNK, BV) has its own specific implementation with slightly different allowable values and weighting matrices. Use the flag-state-specified class rules for any actual design calculation.

See also

• Engine Alignment and Bedplate Flexure: influence coefficients, hot and cold alignment, bearing settlements
• Engine Torsional Vibration Analysis: harmonic torque analysis, critical speed avoidance, detuner sizing
• Marine Reduction Gears: gear train design, loss factors, CPP gearbox arrangements
• Marine Propeller: propeller theory, blade section design, cavitation
• Marine Propeller Pitch and Construction: pitch distribution, material grades, hub design
• Marine Cathodic Protection and Hull Coatings: anodes, shaft earthing, impressed current
• Propeller Theory Deep Dive: actuator disk, wake fraction, thrust deduction
• IACS Shaft Diameter Calculator
• Propeller Shaft SHP vs DHP Calculator
• Shaft Torsional Critical Speed Calculator
• Engine Shaft Alignment Sag Calculator
• Bearing Hydrodynamic Film Calculator
• Tech Shaft Gearbox Losses Calculator
• Shaft Generator Calculator
• Controllable Pitch Propeller Hydraulic Hub Calculator