Engine Alignment and Bedplate Flexure

Main engine alignment is one of the few engineering tasks on a ship where an error made at installation can remain invisible for years, then destroy a main bearing without warning. A slow-speed two-stroke diesel on a large vessel carries a crankshaft up to 18 m long on 7 to 14 main bearings. Each bearing must carry close to its design share of the combined crankshaft weight and firing loads; a bearing carrying 40% more than its share will fail well before its rated service interval. Crankshaft deflection measurement, the gap-and-sag coupling check, and jacking reaction surveys are the principal tools for verifying that the load distribution is correct at installation and remains so throughout service. The engine-shaft-alignment-sag calculator computes the cold-to-hot thermal sag correction for a specific engine installation.

Why alignment is a structural problem, not just a precision problem

The engine bedplate is a fabricated steel box girder, typically 6 to 12 m long, that sits on chocks bolted to the hull double-bottom tank top. It is not rigid. Under the combined static weight of the engine and the dynamic firing loads from cylinder combustion, the bedplate deflects continuously. The hull double-bottom it sits on is also not rigid: it deflects under the ship’s displacement, bends with cargo loading, and cycles with wave-induced hull-girder bending in seaway conditions.

The result is that the centreline through the crankshaft bearings is never a perfect straight line. It is a curve that changes shape with loading condition, temperature, and time. Alignment work amounts to controlling that curve so every bearing operates within its hydrodynamic film regime.

IACS UR M53, which governs crankshaft design and scantling calculation, treats the crankshaft as a beam supported on elastic foundations (the main bearings). The allowable crankshaft web stresses are derived from a calculation model that assumes the bearings provide support within a specified tolerance band. When the as-installed bearing positions fall outside that band, the actual web stresses exceed the allowable, and the UR’s safety margin is consumed.

Crankshaft deflection: measurement and limits

The deflection meter and what it reads

A deflection meter (K-meter in MAN B&W terminology) is a precision dial gauge or electronic transducer mounted between two adjacent crank webs. With the crankshaft positioned so the throw in question is at bottom dead centre (BDC), the instrument is set to zero. The crankshaft is then rotated, and readings are taken at four standard positions: bottom (BDC, the starting zero), port (90 degrees), top (TDC), and starboard (270 degrees). A final reading at BDC confirms the instrument hasn’t shifted.

The four readings give two deflection values per throw:

where $\delta_V$ is the vertical deflection (positive when BDC reads larger gap than TDC, meaning the crankshaft centre is displaced downward relative to the web ends) and $\delta_H$ is the horizontal deflection. Both values are recorded in hundredths of a millimetre.

Modern electronic K-meters log readings automatically and output the full deflection diagram across all throws, flagging exceedances against the pre-programmed tolerance.

Tolerance limits

MAN Energy Solutions specifies deflection tolerance in its installation manuals as a function of engine stroke. For the ME-C family the limit is:

where $S$ is the stroke in millimetres. For an S80ME-C engine (800 mm stroke) the limit is $\pm 0.12\text{ mm}$; for a G95ME-C (2,900 mm stroke) it is $\pm 0.44\text{ mm}$. These limits apply to the cold afloat condition at even keel.

WinGD specifies comparable limits for the RT-flex and X-DF families. The X72DF (780 mm stroke) carries a tolerance of $\pm 0.10\text{ mm}$ vertically. WinGD installation documents note that vertical deflection limits are typically tighter than horizontal because the dominant hull-girder deformation acts vertically.

Class societies review manufacturer limits and confirm or tighten them in the ship’s alignment documentation package. DNV’s rules for rotating machinery require that crankshaft deflections recorded at final sea trial lie within the manufacturer’s approved limits and that any deviation above 50% of limit triggers a bearing inspection before the vessel enters commercial service.

Pattern interpretation

The spatial pattern of deflections across all throws tells the surveyor more than any single reading. A consistent positive vertical deflection across every throw (BDC larger than TDC) indicates the crankshaft centre is displaced downward throughout its length, meaning the engine sits in a convex bedplate curve: the bedplate’s ends are higher than the middle. This pattern is characteristic of hull hogging, where the double-bottom lifts at mid-length. The opposite pattern, negative vertical deflections across all throws, indicates hull sagging.

A single anomalous throw, where one cylinder’s vertical deflection is two to three times larger than neighbouring throws, points to a local problem: a worn main bearing at that position, a cracked or deformed bedplate cross-girder, or a failed chock under that bearing location. A progressive trend from fore to aft, where deflections increase steadily along the engine, indicates the bedplate is twisted longitudinally rather than simply bowed.

Horizontal deflections are normally small (under 0.05 mm) on a well-installed engine. Large horizontal values suggest the bedplate is laterally off-centre on its chocks, or that hull-girder torsion is significant (relevant to container ships in oblique seas).

Hull-girder deformation and bedplate flexure

Static load-condition shifts

The hull double-bottom deflects differently in full-load, ballast, and dry-dock conditions. For a Panamax bulk carrier the mid-ship double-bottom can rise or fall by 3 to 6 mm from light ballast to full-load displacement. The engine bedplate follows the double-bottom surface. MAN Energy Solutions’ installation guidance for large bore engines states that the algebraic difference between the in-dock cold deflection and the afloat loaded condition should be incorporated into the target deflection values at installation, so that the afloat loaded condition falls at the centre of the tolerance band.

For the alignment calculation to capture this effect, the naval architect provides a predicted hull-girder curvature diagram for the loaded and ballast conditions. The engine builder uses this to compute the expected change in bedplate sag from dock condition to service condition, and the engine is set in dock with an intentional bias that will reduce to near-zero when the ship is loaded afloat.

The magnitude of this correction is typically 0.05 to 0.15 mm for a 6-cylinder engine on a 40,000 DWT vessel. For a 9-cylinder engine on a 180,000 DWT capesize bulker, the correction can reach 0.25 mm.

Hull hogging and sagging

Hull hogging (the ship centre rising relative to its ends) acts like a three-point bending load on the bedplate. The middle bearing positions are forced upward relative to the end bearings. The crankshaft webs at mid-engine show larger positive vertical deflections. The central main bearings carry increased load; the forward and aft bearings are partially offloaded.

Sagging reverses this: mid-engine bearings are depressed, losing load, while end bearings gain it.

Both effects are continuous in service. DNV guidance on shaft alignment (covering both crankshaft and propulsion shafting) notes that wave-induced hull-girder bending can produce instantaneous double-bottom deflections of 0.5 to 2 mm on a large ore carrier in heavy weather, and that this cycling stress must be superimposed on the static misalignment stress when assessing fatigue life of crankshaft webs.

Thermal growth of the bedplate

When the engine warms from ambient to full-load operating temperature, the bedplate expands. Steel’s coefficient of thermal expansion is approximately $11.7 \times 10^{-6}\text{ mm/mm/°C}$. For a bedplate 8 m long that warms from 20°C to 80°C, the longitudinal growth is:

This growth is constrained by the holding-down bolts and chocks, so it does not translate to a simple lengthening. Instead, differential temperature through the bedplate height (the crankcase bottom running warmer than the tank top interface) creates a bowing moment. The bedplate typically rises at mid-length relative to its ends when hot, producing a slight hogging of the crankshaft centreline that partially cancels the hull-sag bias in many loading conditions.

MAN Energy Solutions’ cold-to-hot correction values for ME-C engines account for this, and the engine-shaft-alignment-sag calculator implements these corrections for specific engine configurations.

Bearing load distribution and the alignment calculation

From deflection to bearing loads

The relationship between crankshaft deflection readings and main bearing loads is not direct but requires an elastic beam calculation. The crankshaft is treated as a continuous beam on elastic supports (the bearings), and the deflection diagram is the observed deformed shape. Given the crankshaft stiffness (from the engine builder’s data) and the observed deflections, the bearing reaction forces can be back-calculated.

The standard method, documented in MAN Energy Solutions’ alignment software (MAN CoCoS and the ME Engine Alignment Program) and WinGD’s equivalent, uses influence coefficients: for each bearing, a unit displacement gives a known set of reactions at all other bearings. Superimposing the actual displacement pattern gives the actual reaction set.

The result is expressed as bearing loads in kN, compared against the nominal design loads. Class rules, including Lloyd’s Register Part 5 Chapter 5, require that the alignment calculation demonstrate each bearing carries at least 30% of its nominal design load (to ensure hydrodynamic film formation) and not more than 200% (to avoid white-metal fatigue). Some class societies tighten this: DNV requires at least 40% of nominal for stern tube bearings on long shafts.

White-metal bearing fatigue under overload

White metal (tin-based Babbitt alloy, typically 88% Sn with Sb and Cu additions) is the standard bearing surface for main bearings on large slow-speed diesel engines. Its compressive fatigue limit under cyclic loading is approximately 10 to 14 MPa, depending on alloy and thickness. Main bearing projected areas on modern engines are sized to keep maximum pressure well below this at nominal loads, providing a fatigue life of 30,000 hours or more at the design bearing load.

When alignment error concentrates load on one bearing, the projected-area pressure increases in proportion. A bearing carrying 180% of nominal load is operating near or above the white-metal fatigue limit on every firing cycle. Subsurface fatigue cracks initiate, grow, and eventually produce spalling of the white-metal layer. The bearing then runs metal-to-metal, crankshaft journal scoring follows quickly, and the repair requires an engine lift.

The damage is silent until it is severe. Modern engines with bearing temperature sensors or oil mist detectors will give early warnings, but the underlying cause is the misalignment, not the bearing itself.

Underloaded bearings and film breakdown

The opposite failure is less intuitive but equally damaging. A bearing carrying less than 30% of nominal load lacks sufficient specific pressure to form a stable hydrodynamic oil film. The oil film requires a minimum pressure to prevent metal-to-metal contact during the peak firing cycle. Below the minimum load, the journal runs in boundary lubrication for part of each revolution, producing rapid wear of both the white metal and the journal surface.

The oil analysis programme will show elevated tin (from white metal) and iron (from journal wear) in the bearing drain oil sample. On modern engines with lube oil analysis continuous-monitoring capability, this shows up within a few hundred running hours of alignment going wrong.

Shaft alignment: the propulsion shafting perspective

The coupled system

The marine propulsion shafting system connects the engine output flange to the propeller through intermediate shafts and a tail shaft in the stern tube. The crankshaft alignment and the shafting alignment are coupled: the engine output bearing loads are affected by the force and moment that the intermediate shaft imposes at the coupling flange, and the intermediate shaft’s bearing loads are affected by the engine flange’s position and angle.

IACS UR M68 prescribes minimum shaft diameters based on rated power, rpm, and material tensile strength, with the diameter sized to limit nominal torsional shear stress. The UR’s allowable stresses assume the shaft operates in pure torsion with limited bending. When alignment error introduces bending stress in addition to torsional stress, the effective safety factor against fatigue falls below the UR’s assumption. The shaft diameter calculator (IACS UR M68) computes scantling diameters; the bending contribution from misalignment is an additional check outside the basic scantling formula.

Gap-and-sag coupling check

The gap-and-sag method is the standard field check for coupling alignment. The two flanges are brought close together without bolting. A clock gauge is mounted on a bridge bar spanning across one flange and reading against the other. Four readings are taken: top, bottom, port, starboard. At each position, two quantities are recorded:

• Gap: the axial separation between the flange faces at that position.
• Sag: the dial reading on the bar bridge (which sags under its own weight when horizontal and must be corrected for bar deflection).

The differences give the angular offset (the two flanges are not parallel) and the parallel offset (the flange centrelines are offset but parallel). Both must fall within the coupling manufacturer’s and engine builder’s tolerances, typically:

These values are indicative; the engine builder and class surveyor use the approved alignment calculation document, which specifies the exact tolerances for the installation.

Sighting methods: optical and laser

Optical alignment uses a precision telescope or jig transit to establish a sight line along the shaft centreline. Bearing positions are measured against this reference. The method is accurate to approximately 0.05 mm over a 10 m span. Optical methods are standard in new-build yards for initial bearing positioning before the shaft is lowered.

Laser alignment tools use a laser emitter and target, giving readings accurate to 0.01 mm over 10 m, with the readings logged electronically. The laser system’s advantage over optical is the ability to measure bearing offset directly without human parallax error, and the automatic calculation of angular and parallel offsets from the raw position data. DNV and Lloyd’s Register both accept laser alignment records in the approval documentation package.

The primary limitation of optical and laser methods is that they give a static geometry measurement. They tell you where the shaft centreline is at that moment; they do not directly give you the forces at the bearings. The forces require an additional calculation from the geometry and the shaft/bearing stiffness data.

Jacking and reaction measurement

Jacking tests (also called bearing reaction tests or jack-up tests) measure the actual bearing loads directly rather than inferring them from geometry. A hydraulic jack is placed under the shaft near each bearing location, and the shaft is lifted incrementally by 0.1 mm steps. The jack load at which the shaft lifts off the bearing (the bearing reaction changes from loaded to unloaded as the jack passes through zero) is recorded. This load equals the actual bearing reaction at that location.

The procedure is time-consuming (the jack must be repositioned at each bearing location, with the shaft at operating temperature if a hot-check is required) but it is the most direct confirmation of the bearing load distribution. DNV’s class rules for shaft alignment require jacking tests to be carried out on the final installation as part of the survey documentation for vessels where the calculated alignment shows any bearing carrying less than 50% of nominal load, as a safety check on the calculation.

The bearing hydrodynamic film calculator uses bearing geometry and oil viscosity inputs to estimate the minimum film thickness at a specified bearing load, giving the surveyor a cross-check against the jacking result.

Alignment methods: comparison

The most reliable alignment verification combines at least two of these methods. A common sequence on a large new-build is: optical positioning during installation, crankshaft deflection checks during engine-over-turning before launch, gap-and-sag at coupling-up after launch, and a final crankshaft deflection afloat at full displacement before sea trial.

Cold vs hot alignment: what changes

Thermal growth at the engine

When the slow-speed diesel warms from cold ambient (approximately 20°C) to operating steady-state (mean crankcase temperature 60 to 80°C, main bearing housing approximately 50°C above ambient), several dimensional changes occur:

The bedplate rises at mid-length relative to its ends, because the crankcase sides run hotter than the tank-top interface. For a 7-cylinder MAN B&W G80ME-C the bedplate’s mid-length thermal rise is approximately 0.12 mm above the cold-dock condition, reducing the apparent hogging deflection at mid-engine.

The main bearing housings expand both axially and vertically. Axial growth affects coupling gap checks. Vertical growth shifts the crankshaft centreline upward relative to the top tank plates.

The jacket water system, when circulating at 85°C, heats the cylinder jackets and the top of the bedplate differentially. WinGD’s alignment guidance for X-series engines gives specific cold-to-hot offset values for each engine model and cylinder count, which are incorporated into the alignment target at installation.

Why cold-check is preferred

Cold deflection checks are preferred for routine service because the engine can be barred over immediately after a port stop or maintenance period, without waiting for cooling or running for warmup. The consistency of the cold condition makes baseline comparison meaningful over years of service data.

Hot checks are required when the manufacturer specifies a hot tolerance (as WinGD does for some X-DF installations) or when the cold check shows a marginal reading that may pass in the hot condition. Some class societies require a hot check as part of the five-year special survey alignment verification.

Chocking, holding-down bolts, and long-term stability

Epoxy chocks

Epoxy resin chocks (Chockfast Orange and equivalent two-component systems) are standard for new-build installations. Liquid epoxy is poured between the bedplate sole plate and the hull tank top, with retaining dams around the perimeter. The cured chock, typically 20 to 60 mm thick, has a compressive strength of 80 to 120 MPa and a modulus of elasticity of 3 to 5 GPa, giving it enough compliance to distribute load evenly across the bedplate contact surface without point-loading the tank-top plating.

Epoxy chock life is 20 years or more if undisturbed. Cracking occurs from impact loads (if a piston crown failure drops a heavy mass on the bedplate), thermal shock (coolant leak that floods a hot chock), or hydraulic lock (if sea water enters the chock interface and freezes in cold-climate service). A cracked chock under a bearing support point shifts the local bearing height and produces a detectable deflection anomaly within 500 to 1,000 running hours.

Inspection at major overhauls involves visual inspection for cracks plus a hammer-tap test around the perimeter. A hollow sound indicates delamination. Load cells placed temporarily under the bedplate during a jacking test can confirm whether a suspect chock is still transmitting load.

Holding-down bolts

Holding-down bolts clamp the bedplate to the hull structure through the chocks. They carry tensile load during normal running (preventing the engine from lifting under firing pulses transmitted through the tie rods) and shear load during manoeuvring. MAN Energy Solutions specifies hydraulic bolt tensioning to defined torque values per bolt, with load cells confirming achieved preload. Incorrect bolt preload is a common post-refit alignment problem: bolts under-tensioned after a chock repair allow the bedplate to rock, bolts over-tensioned distort the bedplate locally.

The engine-bedplate-construction article covers the structural design of the bedplate box girder, the tie-rod pre-compression system, and the load paths in detail.

Cast iron chocks

Older installations and some specialist applications use machined cast iron chocks. These are precision-machined to the measured clearance between bedplate and tank top and are shimmed to the final height. Cast iron chocks are easier to replace individually (no curing time, no dam-and-pour work) but require more preparation time at installation and are sensitive to corrosion at the mating surfaces.

Class society approval of alignment calculations

What the alignment calculation package contains

Before a new-build receives class approval for initial sea trial, the yard must submit an alignment calculation document to the class society for review. This document typically includes:

• The engine builder’s crankshaft deflection tolerances for the specific engine model.
• The hull deflection diagram (from the naval architect’s finite element model of the hull girder) showing predicted mid-ship deflection at full-load, ballast, and dock conditions.
• The proposed installation target deflection values, showing the dock-condition intentional bias.
• The propulsion shafting alignment calculation, including the gap-and-sag targets at each coupling and the bearing reaction calculation for at least three loading conditions.
• The expected cold-to-hot correction values.
• The bearing load distribution showing each bearing within the minimum/maximum load ratio specified by the class society (DNV: minimum 40% of nominal for stern tube bearing; Lloyd’s Register: minimum 30% for intermediate shaft bearings).

Approval basis

DNV’s rules for classification of ships require that the alignment calculation be approved by the class before the initial survey of the shafting. Lloyd’s Register’s rules (Part 5 Chapter 5) require the calculation and the sea trial deflection records to be submitted together for final class confirmation. ABS and ClassNK have equivalent requirements in their respective machinery rules.

The class surveyor does not typically perform independent alignment calculations. The surveyor reviews the yard’s submission, verifies that the calculation method is consistent with the class rules, checks that the tolerances used match the engine builder’s approved data, and witnesses the crankshaft deflection measurements and coupling checks at installation and sea trial.

Where the as-found deflection measurements show exceedances, the yard must document the corrective action taken and provide revised measurements before the class surveyor will issue the survey record.

Service monitoring and periodic re-verification

Deflection schedule in service

Most engine builders and class societies recommend crankshaft deflection checks at the following intervals:

• At every cylinder overhaul (typically every 5,000 to 8,000 hours on crosshead engines).
• At every major overhaul (typically every 30,000 to 40,000 hours).
• After any event that could have shifted alignment: grounding, dry-docking, structural repair to the engine room bottom, replacement of a main bearing, or replacement of a chock.

The results are compared against the new-build baseline record. A trend of increasing vertical deflection at mid-engine over successive checks points to progressive hull-girder creep or chock degradation. Trend data is more valuable than any single reading: a single reading outside tolerance is an alarm; a consistent trend toward tolerance is an early warning while corrective action is still straightforward.

Bearing condition indicators

Three monitoring channels cross-check the alignment condition in service:

Bearing temperature: Main bearing drain-oil temperature sensors (standard on modern slow-speed engines) respond to increased bearing load within one to two watch periods. A bearing gaining 5°C above its normal running temperature while neighbouring bearings remain constant is an early sign of load redistribution.

Oil analysis: Tin (Sn) and lead (Pb) in main-bearing drain oil samples indicate white-metal wear. An ICPMS (inductively coupled plasma mass spectrometry) analysis of monthly drain oil samples shows whether the rate of tin pickup is increasing, which precedes visible bearing damage by hundreds to thousands of hours on a slow-speed engine.

Crankshaft vibration: The engine torsional vibration analysis article covers torsional resonance, but lateral vibration at bearing frequency also changes with bearing load distribution. A bearing underloaded to near-zero carrying capacity loses its oil film and develops a characteristic orbit instability. Shaft vibration sensors (available on research vessels and a growing number of commercial tonnage with continuous monitoring) can detect this onset.

Alignment after dry-docking

Dry-docking always shifts alignment. The ship’s displacement drops to zero (keel blocks support the hull from below, replacing buoyancy), the hull assumes a configuration close to hogging (centre supported, ends free), and the double-bottom temperature changes as the ship is ventilated. Crankshaft deflection checks taken immediately after flooding the dock before undocking will show this shifted pattern. The practical rule, documented in Lloyd’s Register’s guidance notes on shaft alignment surveys, is that deflection measurements taken in dry dock are valid for confirming the dock-condition baseline but are not directly comparable to the afloat service limits; the afloat sea trial check is required for service acceptance.

For a vessel going through a special survey at a five-year interval, the standard practice is to take crankshaft deflection readings in dock, carry out bearing inspections if indicated, and then take readings again afloat before the class surveyor closes the survey.

Limitations

The deflection-meter technique measures the deformed geometry of the crankshaft. The back-calculation from geometry to bearing loads involves the crankshaft stiffness data, which the engine builder provides as part of the alignment calculation package. That stiffness data is derived from finite element analysis and is specific to the engine model and cylinder count. Applying a deflection reading from one engine type to a different engine’s stiffness data produces incorrect load estimates.

The gap-and-sag coupling check captures misalignment at the coupling location but does not directly reveal the bearing loads. A coupling can be within tolerance and still produce a bearing overload if the intermediate shaft is not correctly supported between couplings.

Jacking reaction tests are accurate to about 5 to 10% of the bearing load because the jack placement is rarely exactly at the bearing centreline, the test requires the shaft to be accessible and the bearing housing to be removable or have jacking provisions, and the measured jack load includes some shaft bending contribution from the span between the jack and neighbouring supports.

Hull-girder deformation predictions from finite element models carry uncertainties of 10 to 20% compared to measured values in service. The alignment target set in dock assumes the FE prediction is correct; if the actual service deflection differs, the bearing loads will differ correspondingly. This is why periodic in-service deflection checks are not optional: they are the calibration loop that catches divergence between the predicted and actual hull behaviour.

All alignment calculations assume the chocks are intact and uniformly transmitting load. A partially delaminated epoxy chock can transmit load through a smaller contact area, increasing the local contact pressure even though the visible external dimensions appear correct.

The influence-coefficient method for bearing reactions

The term “influence coefficient” appears in MAN Energy Solutions’ and WinGD’s alignment documentation without always being explained. The underlying mechanics are worth understanding because the method sets the theoretical ceiling on what crankshaft deflection data can tell you.

Each main bearing is treated as a spring support in the multi-span beam model of the crankshaft. An influence coefficient $a_{ij}$ is the vertical reaction force at bearing $i$ caused by a unit imposed displacement at bearing $j$, with all other bearings fixed in position. The engine builder calculates these coefficients from the crankshaft geometry and stiffness properties using finite element analysis; they are published as a square matrix in the alignment calculation package for each engine model and cylinder count.

Given a measured deflection pattern across all throws, the actual bedplate deviation at each bearing position can be estimated. The bearing reaction vector $\mathbf{R}$ is then:

where $\mathbf{A}$ is the influence-coefficient matrix and $\mathbf{d}$ is the vector of bearing-position deviations from the ideal straight line. Inverting $\mathbf{A}$ is straightforward when the system is well-conditioned, but ill-conditioning can occur on engines with many main bearings (9 or more) if the bearing spacings are non-uniform. MAN Energy Solutions’ ME Engine Alignment Program handles this numerically and flags cases where small deflection errors would produce large reaction uncertainties.

The key practical consequence: two adjacent bearings that are both displaced in the same direction by similar amounts will show only a small change in their individual reactions, because the crankshaft bridges across the common displacement. A single bearing displaced while its neighbours remain fixed produces a much larger reaction change. This explains why isolated deflection anomalies at one throw are diagnostically more serious than a slow background trend across all throws.

Stern-tube bearing slope and the seawater-buoyancy correction

Single-slope and double-slope stern-tube bushes

The stern-tube aft bearing carries the largest single bearing load in the propulsion shafting system because it supports the tail shaft close to the propeller, where the propeller weight and thrust eccentricity both act. A straight-bore stern-tube bush concentrates this load at the aft edge of the bearing under the propeller weight moment, producing a non-uniform pressure distribution that accelerates wear at the entry edge.

Single-slope (conical bore) stern-tube bushes address this by machining a slight taper into the bearing bore, typically 0.05 to 0.15 mm over the bearing length, so the shaft enters the bearing at a controlled angle. This spreads the load across the full bearing length under the expected operating shaft angle.

Double-slope bushes carry this further: the forward half of the bush has one bore angle and the aft half has a slightly different angle, matching the expected shaft curvature under combined propeller weight and hydrodynamic lift. DNV’s shaft alignment guidance notes that double-slope bushes are specified when the calculated aft-bearing slope angle under propeller weight exceeds 0.3 milliradians, which is common on single-screw vessels above 30,000 DWT with propeller diameters above 6 m.

Stern-tube wear-down matters for alignment because the aft bearing is the anchor point for the entire shaft system. If the aft stern-tube bearing wears down by 0.5 mm (by no means unusual after 60,000 running hours on a conventional white-metal bush), the tail shaft sag increases, which changes the bending moment at the stern tube forward seal, shifts the coupling angle at the intermediate shaft coupling, and redistributes load to the intermediate shaft bearings. Lloyd’s Register’s guidance on five-year special surveys requires wear-down measurement of the aft stern-tube bearing as part of the routine shaft alignment survey, with a re-alignment calculation required if wear-down exceeds 1.0 mm or 25% of the design clearance, whichever is smaller.

Cutlass rubber and composite polymer bushes (increasingly specified for environmental reasons on vessels where oil-lubricated stern tubes are being phased out under MARPOL Annex I requirements) have lower wear rates than white metal but different elastic behaviour. Their stiffness is lower (modulus around 0.5 to 1.0 GPa for UHMWPE versus 50 to 70 GPa for white metal in bending), so the bearing acts as a softer spring support, changing the influence coefficients and the load distribution through the rest of the system.

Afloat vs drydock measurements and the buoyancy correction

A tail shaft in water experiences buoyancy along its submerged length. For a 400 mm diameter solid forged steel tail shaft, 5 m submerged in seawater, the buoyancy force is:

This force acts upward along the submerged length, reducing the effective weight of the tail shaft on the aft stern-tube bearing. In drydock, that buoyancy is absent, and the aft bearing carries the full tail-shaft weight. The difference between the afloat and drydock bearing loads at the aft position can therefore be 6 to 15 kN for typical tail shaft dimensions, which is enough to shift the alignment condition detectably.

MAN Energy Solutions’ installation documents address this by providing separate alignment target windows for the afloat and in-dock conditions. The dock-condition target accounts for the absent buoyancy by allowing the aft bearing to be slightly higher (the shaft sets lower relative to the bearing, compensating for the added weight), so that when the ship is floated and buoyancy returns, the shaft settles to the intended afloat position.

Alignment calculations that ignore the buoyancy correction will show the bearing loads as correct in dock but overloaded at the aft position afloat. DNV’s rules for shaft alignment state explicitly that the bearing reaction calculation must be provided for both the afloat operating condition and the dry-dock condition, with the buoyancy correction applied to the tail shaft and any other submerged shaft sections.

Class survey triggers for re-alignment

IACS UR M68 defines minimum shaft diameters but does not prescribe a full re-alignment procedure; that requirement comes from the individual class society rules and the IACS unified interpretation framework. DNV’s rules, Lloyd’s Register Part 5 Chapter 5, ABS Machinery Rules Part 4, and ClassNK Machinery Rules all identify events that require a formal re-alignment survey before the vessel resumes service:

Grounding: any grounding that results in contact between the hull and the sea bed is a class notification requirement. The double-bottom deformation that follows even a moderate grounding can shift bearing positions by 1 to 3 mm locally, far exceeding deflection tolerances. Post-grounding crankshaft deflection checks are mandatory before restarting the main engine.

Main bearing failure: a spalled or scored main bearing indicates the bearing was operating outside the hydrodynamic film regime, which is consistent with misalignment. Replacing the bearing without checking the alignment removes the symptom but not the cause. All four class societies require a deflection survey and alignment recalculation after any main bearing replacement.

Structural repair to the engine room bottom: any repair involving hot work, cutting, or re-welding within 3 m of the engine seatings can relieve or introduce residual stress in the double-bottom plating, shifting the chock bearing surface. The class rule threshold (DNV Section 11, Lloyd’s Register Part 5) is typically any repair that involves replacing a floor, stringer, or tank top plate under or adjacent to the engine foundation.

Major port-state detention: when a vessel is detained for main-engine crankshaft deflection exceedances found during a port-state inspection, re-alignment with class verification is required before the PSC inspector will lift the detention order. This is documented in Paris MOU and Tokyo MOU port-state control procedures for propulsion-related deficiencies.

Common alignment defects found in service

Two failure patterns dominate the in-service alignment defect record across the major class societies’ survey databases.

Overloaded aft main bearing: the aftmost main bearing, adjacent to the engine output flange, tends to accumulate load because the intermediate shaft imposes a downward bending moment at the flange when the shafting system sags under the propeller weight. If the installation alignment did not correctly account for the shaft system’s static sag contribution to the engine flange moment, the aft main bearing is overloaded from day one. The first sign is elevated drain-oil temperature at that bearing (typically 5 to 8°C above the other bearings’ normal running temperature), followed within 3,000 to 8,000 running hours by tin pickup increasing in the drain-oil analysis.

Unloaded intermediate shaft bearing: an intermediate shaft bearing that is set too high during installation, or that rises due to chock creep, carries less than its design share of the shaft weight. The shaft spans across the bearing without contacting it consistently. This produces fretting wear at the bearing surface (the shaft oscillates against the bearing under torsional and lateral vibration without a stable oil film), and a characteristic increase in lateral vibration amplitude at the bearing’s natural frequency. The marine propulsion shafting and stern tube article covers the bearing support arrangements for intermediate shafting and the chock geometry that governs long-term height stability.

Both defects are detectable early from routine monitoring channels: temperature trending, oil analysis, and crankshaft deflection records. The consistent finding across surveys is that vessels with active trend-monitoring programmes catch and correct these defects before bearing failure; vessels relying solely on fixed-interval inspections frequently present for survey with bearings already at or past the fatigue limit.

See also

• Engine Bedplate Construction: Two-Stroke Marine Diesels
• Marine Engine Crankshaft and Main Bearings
• Marine Propulsion Shafting and Stern Tube Systems
• Engine Torsional Vibration Analysis
• Two-Stroke Marine Diesel Engine Fundamentals
• Shaft Alignment Cold-to-Hot Sag Calculator
• Shaft Diameter IACS UR M68 Calculator
• IACS UR M71 Crankshaft Scantling Check
• Bearing Hydrodynamic Film Calculator