Cylinder Bore and Stroke Selection for Marine Engines

Cylinder bore and stroke are the two dimensions that fix every other performance variable in a reciprocating engine. Set the bore and the stroke and you determine swept volume, rated speed, mean piston speed, propeller compatibility, engine height, structural mass, and the fuel-consumption envelope the vessel will operate in for its 25-year life. For slow-speed two-stroke marine diesels, bore and stroke selection is a propulsion-system-level decision, not an engine-internal one.

The stroke-to-bore ratio (S/B) in main propulsion slow-speed two-strokes has moved from approximately 2.0 in early crosshead diesels to more than 4.6 in the current MAN B&W G-series. That progression happened because lower rpm demands from propeller designers consistently outpaced the ability to reduce rpm by any other means. This article traces the physics driving that evolution, the trade-offs at each step, the manufacturer design families built around specific S/B bands, and the regulatory context that MARPOL Annex VI provides through EEDI and EEXI.

For the load-diagram context in which a selected bore and stroke are placed as rated MCR, see engine load diagram and operating envelope. For the cylinder liner design consequences of bore choice, see cylinder liner design for two-stroke engines. For the full thermodynamic underpinning, see two-stroke marine diesel engine fundamentals.

Swept volume and the power equation

The swept volume per cylinder, $V_s$, is:

where $d$ is bore in metres and $s$ is stroke in metres. This is the volume the piston displaces on each power stroke. For a two-stroke engine with $z$ cylinders running at $n$ revolutions per second, the brake power is:

For a four-stroke engine the right-hand side is divided by 2, because only one in every two revolutions produces a power stroke per cylinder. Slow-speed marine diesels are two-stroke uniflow-scavenged; the factor is 1.

Two things flow immediately from this equation. First, $P_b$ scales with the product of BMEP and swept volume and rotational speed. To maintain the same $P_b$ while reducing $n$ (to run a larger, slower propeller), you must increase BMEP, or swept volume, or both. Second, swept volume scales as $d^2 s$: doubling the stroke doubles swept volume; doubling the bore quadruples it. So increasing stroke gives more volume per unit of dimensional change than increasing bore, which is one reason engine designers have stretched stroke aggressively.

Mean piston speed: the tribological ceiling

Mean piston speed $c_m$ relates stroke, speed, and wear rate:

where $s$ is in metres and $n$ is in revolutions per second, giving $c_m$ in m/s. For a G70ME-C10.5 running at 76 rpm ($n = 1.267 \text{ rev/s}$) with a 3.256 m stroke:

This is consistent with MAN Energy Solutions’ published data for that engine at maximum continuous rating (MCR).

The limit matters because piston ring and cylinder liner wear rate rises steeply above approximately 8.5 m/s in slow-speed engines. Below about 6 m/s, scavenging quality suffers and combustion becomes irregular. The 6.0 to 8.5 m/s band defines the practical operating corridor for modern two-stroke designs. Historically, 1960s engines ran at 6.0 to 7.0 m/s; current ultra-long-stroke designs push 8.0 to 8.5 m/s at MCR. The tribological gain from modern plateau-honing of cylinder liners, combined with sulphur-adaptive cylinder lubrication, enabled that rise without proportional wear-rate penalties.

The key insight is that $c_m = 2sn$ connects stroke and speed in one equation. If the propulsion system demands $n = 70 \text{ rpm} = 1.167 \text{ rev/s}$ and the tribological ceiling is $c_m = 8.5 \text{ m/s}$, then stroke is constrained to:

That calculation tells the engine designer what the maximum stroke is before selecting any other parameter.

Brake mean effective pressure and its limits

BMEP characterizes the cycle-average gas pressure that, acting through the full stroke, delivers the observed work output. It allows comparison across different cylinder geometries without reference to physical size. Modern slow-speed two-stroke engines operate at 19 to 21 bar BMEP at MCR, up from 12 to 14 bar in 1980s designs. The jump came from higher peak firing pressures (now 190 to 220 bar in current crosshead engines), better fuel injection timing control under electronic management (the “ME-C” and “ME-GI” designation in MAN B&W nomenclature), and improved scavenging through longer stroke.

Peak firing pressure is the binding structural constraint on BMEP. For a given bore, the gas force on the cylinder cover bolts scales as:

Doubling the bore quadruples this force. Larger bore engines therefore require proportionally heavier cover bolts, larger tie rods, and stiffer bedplates. This is why bore is extended cautiously and in discrete steps, while stroke has been stretched more freely: stroke does not change the peak gas force on the cover.

Beyond about 22 bar BMEP the thermal loading on the piston crown, the exhaust valve, and the cylinder cover reaches the limit of current Inconel and austenitic steel alloys used in the flame-impinged regions. Class societies, specifically IACS Unified Requirement M66, set limits on maximum cylinder pressure as a function of cylinder bore and engine certification class.

The stroke-to-bore ratio: historical progression

The S/B ratio is a single number that summarizes the cylinder geometry:

It also equals the ratio of stroke to bore with no unit dependency. Its historical progression for slow-speed main propulsion two-strokes is documented in MAN Energy Solutions’ own engineering literature:

The G70ME-C10.5, with a 700 mm bore and 3,256 mm stroke, has S/B = 4.65, the highest ratio in a production two-stroke engine as of 2024. The G95ME-C9.5, the largest cylinder bore in the G-series at 950 mm, runs a 3,460 mm stroke for S/B = 3.64, illustrating that the largest bore engines cannot physically carry the same S/B ratio as mid-bore engines because absolute stroke length becomes prohibitive for engine height.

For the per-cylinder performance of specific G-series and X-series engines, the site’s calculator pages give rated data: see for example the MAN ES G70ME-C10.5 calculator or the WinGD X82 calculator.

Why long stroke lowers optimal rpm

Propeller open-water efficiency is governed by the ideal actuator-disk momentum exchange, which for a given thrust and ship speed favours a large, slow propeller over a small, fast one. The Froude efficiency expression shows:

where $C_T$ is the thrust loading coefficient $T / (0.5 \rho A_0 V_A^2)$, with $A_0$ the disk area and $V_A$ the advance velocity. For a fixed thrust and advance speed, a larger disk area $A_0$ reduces $C_T$ and raises $\eta_0$. A larger propeller diameter demands a lower rotational speed to keep blade tip Mach number and cavitation risk acceptable, typically targeting tip speeds of 35 to 48 m/s depending on cavitation-number budget.

For a 10 m diameter propeller at 45 m/s tip speed, the required rpm is:

A VLCC’s optimal propeller, constrained by a 10 m aperture, might demand only 65 to 72 rpm. Matching that without a reduction gearbox requires an engine that naturally produces maximum torque at those speeds. Long stroke achieves this by moving swept volume into the stroke dimension, allowing BMEP to reach 20+ bar at the lower rotational speed with mean piston speed still inside the tribological limit.

Carlton’s “Marine Propellers and Propulsion” (4th ed., Butterworth-Heinemann, 2018) documents the efficiency gain from shifting from 100 rpm to 70 rpm propellers on post-Panamax bulk carriers as 4 to 6 percentage points of open-water efficiency, corresponding to roughly 8 to 10 percent reduction in installed power for the same service speed. That is the primary economic driver for the ultra-long-stroke transition.

Bore selection: the power-per-cylinder calculation

Once the propulsion power, propeller speed, and number of cylinders are fixed, bore follows from BMEP and swept volume. Working through the calculation explicitly:

Suppose a container ship needs 50,000 kW at 80 rpm from an 8-cylinder two-stroke engine. Target BMEP is 20 bar. The required swept volume per cylinder is:

This is a large cylinder. If the engine designer targets a stroke of 3.2 m (S/B around 4.0 for a 800 mm bore), bore follows:

That puts bore near 950 mm, consistent with the G95ME-C class. If instead only 6 cylinders are available (perhaps an engine-room-length constraint), the swept volume per cylinder rises to 3.12 m³, pushing bore above 1 m, which exceeds current liner manufacturing limits. The designer must then accept 7 cylinders or raise BMEP to 21 bar to bring bore back into range.

This calculation is the fundamental bore-selection workflow: propulsion power and propeller rpm constrain swept volume per cylinder, the BMEP target further constrains it, and stroke choice then determines bore. The calculation can be run in reverse via the engine BMEP calculator or the engine mean piston speed calculator.

Cylinder count decisions

Cylinder count is typically 5 to 12 for slow-speed two-strokes; 6 and 7 dominate container-ship main engines while 5 appears in smaller vessels and 8 to 9 in very large container ships. Adding a cylinder at constant bore and stroke raises rated power by 1/z (one-cylinder fraction) and adds roughly 1.5 to 2 m of engine length. Removing a cylinder to shorten the engine forces either a bore increase or a BMEP increase to compensate. Neither is free: bore increase raises structural mass and cover loads; BMEP increase raises peak firing pressure and thermal load.

The WinGD X82 (820 mm bore, 3,375 mm stroke, S/B = 4.12) is commonly supplied as 5-cylinder to 8-cylinder sets for container ships in the 25,000 to 40,000 kW range. The WinGD X82 calculator gives rated data at 84 rpm and 7,150 kW per cylinder at MCR. An 8-cylinder set delivers 57,200 kW.

Stroke selection: the four governing constraints

Stroke is not chosen freely. Four constraints bound it simultaneously.

Mean piston speed limit. As shown, $s \leq c_{m,max} / (2n)$. At 70 rpm and 8.5 m/s, stroke cannot exceed 3.64 m. At 80 rpm, the ceiling falls to 3.19 m.

Engine height. Engine height grows roughly 1.1 to 1.3 mm per mm of stroke extension, accounting for the connecting rod length increase and the longer A-frame. The largest current engines (G95ME-C at 3,460 mm stroke) stand approximately 18 m tall. Engine rooms must accommodate this including overhead lifting clearance for crankshaft and piston withdrawals. For tankers and bulkers with constrained engine-room height, stroke may be limited by deckhead clearance before the tribological ceiling is reached.

Crankshaft forging limits. Crank throws (half the stroke) require open-die forging. At 1.6 to 1.7 m throw (3.2 to 3.4 m stroke), the forging is at the limit of available press capacity in Japan, South Korea, and Germany. Larger throws are made as semi-built crankshafts: individual crank webs and journals forged separately, then shrink-fitted. The G95ME-C uses semi-built crankshafts at its 1,730 mm throw. Beyond 2.0 m throw there is no established production capability.

Scavenging quality. Uniflow scavenging in two-stroke engines needs adequate port-open duration and piston velocity at the scavenge port opening to sweep the exhaust gases effectively. Longer stroke improves this: the piston spends more time near BDC, the scavenge ports open wider in crank-angle terms, and the scavenging process is less rushed. Short-stroke engines at high rpm have historically suffered from poor scavenging quality, which wastes fuel and raises exhaust temperatures. The shift to long stroke partly addressed this. For more on how liner port geometry interacts with scavenging, see cylinder liner design for two-stroke engines.

SFOC, EEDI, and EEXI: the regulatory incentive structure

The connection between bore-stroke geometry and fuel regulation is not indirect. EEDI (MARPOL Annex VI Regulation 21, for new ships) and EEXI (Regulation 28, for existing ships) both reward lower SFOC and higher propulsive efficiency, and both reference the main engine’s certified MCR.

SFOC falls as stroke rises because:

1. Longer stroke raises expansion ratio (the piston travels further past the injection event), extracting more work from each combustion.
2. Lower rated rpm matches larger-diameter propellers, delivering more thrust per kJ of fuel to the propeller.
3. Improved scavenging at long stroke reduces residual exhaust gas fraction in the charge, improving combustion quality.

MAN Energy Solutions’ published SFOC values for the S90ME-C10.5 (S/B around 4.0) at ISO conditions are approximately 162 to 165 g/kWh at MCR, dropping to 153 to 157 g/kWh at the tuning load (typically 75 to 80 percent MCR). For the WinGD X92 at similar rating, published SFOC is 159 g/kWh at MCR. Both figures are verifiable in the respective maker’s project guides.

Engine Power Limitation (EPL) under EEXI constrains the maximum continuous power a ship may use. EPL is enforced either by a physical shaft power limiter or by a revised fuel pump rack setting. For a ship that cannot meet EEXI at its certified MCR, the operator limits power to a fraction of MCR, which the EPL required MCR calculator can resolve. Because EPL reduces the engine to a fraction of its design swept volume, it also reduces thermal efficiency by pushing the engine down its SFOC-versus-load curve to a less efficient operating point. This is a structural penalty of EEXI compliance for over-powered ships that were built before the regulation existed.

The EEXI attained calculator accepts limited MCR, reference speed, capacity, and fuel factors. For any vessel close to the EEXI limit, a bore-and-stroke analysis to identify whether a lower rpm (and therefore different propeller) could have passed EEXI without EPL is instructive, even if only for informing a future retrofit.

Manufacturer bore families

MAN Energy Solutions and WinGD (the two surviving makers of production slow-speed two-stroke marine diesel engines) have each built their product lines around discrete bore families, each with specific S/B bands.

MAN Energy Solutions: S-series and G-series

MAN’s current catalogue for new-builds centres on two series:

S-series (super-long-stroke): bore sizes 35, 40, 46, 50, 60, 65, 70, 80, 90 cm. These are the prior-generation super-long-stroke engines. The S90ME-C10.5, the largest S-series unit, has bore 900 mm and stroke 3,188 mm for S/B = 3.54, rated at 7,320 kW per cylinder at 84 rpm. The S-series is no longer offered for most new-builds but remains in production for replacement and specialist markets.

G-series (ultra-long-stroke): bore sizes 50, 60, 70, 80, 90, 95 cm. The “G” prefix was introduced to distinguish the extended stroke geometry. Key data from MAN’s published programme:

The data above is taken from MAN Energy Solutions’ published Marine Engine Programme. The G70ME-C10.5 has the highest S/B in the current production range. The G95ME-C9.5 has the largest bore and the highest per-cylinder output.

ME-C denotes electronically controlled fuel injection and exhaust valve actuation. ME-GI and ME-LGIM designations add gaseous fuel capability. See the per-engine calculator pages, for example MAN ES G95ME-C9.5, for rated point details.

WinGD: X-series

WinGD (Winterthur Gas and Diesel, the successor to the former Wartsila two-stroke division and the RT-flex platform) offers the X-series in bores from 350 mm to 920 mm. Each model is identified by bore in decimetres: X35 is 350 mm bore, X92 is 920 mm bore.

Data from WinGD’s published X-Engine Project Guides. The X-DF variants of each bore add dual-fuel (LNG primary) capability with the same bore-stroke geometry. The WinGD X92 calculator and WinGD X82 calculator give MCR-versus-cylinder-count data for preliminary sizing.

MCR derating and the layout diagram connection

The bore and stroke define the absolute corners of the layout diagram: maximum bore speed (set by mean piston speed limit) and maximum BMEP (set by structural and thermal limits). Inside those corners the engine designer places the L1 to L4 layout field, and the shipyard then selects the specified MCR (SMCR) within that field.

Derating is the deliberate reduction of MCR below the engine’s maximum capability. For bore and stroke purposes, derating effectively places the operating point at a lower BMEP and mean piston speed than the maximum certified values. This is done when:

• The propeller efficiency optimum falls at a lower rpm than the engine’s MCR speed.
• The ship’s installed power needs to comply with EEXI without EPL.
• The owner anticipates operating primarily at partial load (for CII compliance) and wants the SFOC curve optimized at a lower power band.

Derating has a fuel-consumption cost if the engine’s thermal efficiency peak moves away from the operating point. MAN’s ME-C platform allows the tuning point to be set independently of the rated MCR, so it’s possible to rate the engine at 85 percent of maximum and still tune combustion for peak efficiency at 70 percent, the typical slow-steaming operating point. WinGD’s Intelligent Tuning feature provides the same capability.

For the MCR derating calculation, see the engine MCR derating calculator. For the load diagram context of a chosen SMCR position, see engine load diagram and operating envelope.

Thermal load and combustion-chamber geometry

Bore size directly governs the combustion chamber’s surface-to-volume ratio at top dead centre (TDC). A larger bore gives a shallower, wider combustion space for the same compression ratio and clearance volume, which changes:

Flame travel distance. In a central-injector configuration, the flame front must travel from the injector tip to the cylinder wall. For a 950 mm bore at 22:1 compression ratio, the TDC bowl diameter is roughly 500 to 600 mm; the flame travels 250 to 300 mm from injector tip to the farthest wall region. Modern slow-speed engines use 2 to 4 injectors symmetrically arranged around the piston crown to limit flame travel to acceptable distances. A 700 mm bore engine with 3 injectors places the farthest combustion zone about 160 mm from any injector.

Wall quench zone. A larger bore has proportionally less combustion-chamber wall area per unit of volume, which can slightly reduce quench-zone hydrocarbon formation, but the absolute wall temperature matters more than the ratio. Bore-cooled liners, which run the liner at 75 to 85 degrees C coolant inlet instead of the lower 50 to 65 degrees C of older water-jacket designs, keep the wall surface above the sulphuric acid dew point and reduce cold corrosion. This design is standard in all current MAN and WinGD two-stroke liners above 500 mm bore.

Piston crown thermal gradient. A wider bore spreads the piston crown heat flux over a larger area. Piston crowns in the 800 to 950 mm range use internal water cooling with a coiled tube circuit, or telescopic water injection, to keep the crown surface below 400 degrees C. The design of this circuit is a bore-dependent engineering task: it scales with piston cross-sectional area, not with stroke.

Scavenging quality: why stroke helps

Uniflow scavenging drives combustion products out through a central exhaust valve in the cylinder cover while fresh charge enters through ports in the lower liner. The quality of the scavenge depends on the port-open duration in crank-angle degrees, the scavenge air pressure delivered by the turbocharger system, and the piston velocity profile near BDC.

Longer stroke improves scavenging through two mechanisms. First, more crank-angle degrees are spent near BDC for a given port height, giving more time for the fresh charge to sweep the exhaust. Second, the scavenge air receiver pressure, which is linked to the turbocharger’s exhaust-gas energy, is set by the engine’s pumping work, which scales with displacement. An engine with longer stroke at the same bore has more displacement per cylinder and thus more turbocharger driving energy per unit of power output.

Scavenging efficiency (the fraction of the cleared volume actually filled with fresh charge) in modern uniflow-scavenged two-strokes runs at 0.92 to 0.96 at MCR. At low load (below 30 percent MCR), scavenging pressure from the turbocharger falls and scavenging efficiency drops, which is why auxiliary blowers are fitted on all production slow-speed two-strokes: they maintain scavenge pressure during port maneuvering and low-load operation. The interaction between bore, stroke, and scavenging efficiency is one of the arguments cited by MAN Energy Solutions for the shift to the G-series over the S-series for VLCCs and large bulk carriers that spend significant time at 20 to 30 percent MCR during slow-steaming legs.

Engine height and shipboard integration

Engine height is the primary installation penalty for long stroke. The main components that govern height are the cylinder block and liner assembly (a direct function of stroke), the A-frame that encloses the crosshead and guide system (also scales with stroke), and the sump and bedplate below. MAN’s published dimensions for the G-series show that the G70ME-C10.5 stands approximately 14.9 m from baseplate to exhaust manifold top, while the S70ME-C8.5 at a shorter stroke stands approximately 13.4 m. The 1.5 m difference is significant for engine-room arrangement.

For bulk carriers and VLCCs with deep double-bottom tanks and high engine rooms, 15 to 18 m engine heights are routinely accommodated. For tankers with lower freeboard or vessels with high shafting arrangements, engine height is a genuine constraint. The solution in constrained installations is either to accept a shorter stroke (and a gearbox or higher-rpm propeller) or to use a medium-speed engine with a reduction gearbox. For a quantitative comparison between slow-speed direct-drive and medium-speed geared arrangements, the engine power and BMEP relationships article covers the trade-off in detail.

Weight and structural consequences

Cylinder block and crankshaft mass scale strongly with bore and stroke. The crankshaft of a G95ME-C8-cylinder engine weighs approximately 650 tonnes. The cylinder block and A-frame for the same engine add another 1,200 to 1,400 tonnes. Total engine mass for a large slow-speed main engine is 2,000 to 3,500 tonnes.

For comparison, the entire structural steel of a deckhouse on a medium-sized bulk carrier might be 800 tonnes. The engine is the single heaviest component aboard most cargo ships. Its mass and centre of gravity directly affect the lightship displacement, trim, and stability calculation. Bore and stroke decisions therefore feed into the ship’s stability and hydrostatic design long before the keel is laid.

Operations and maintenance: bore and stroke over the vessel’s life

The bore-stroke combination selected at build time determines maintenance costs for the life of the engine. Three cost drivers are bore-and-stroke-dependent.

Liner wear life. Mean piston speed ($c_m = 2sn$) correlates with the rate of sliding wear at the liner running surface. Long-stroke engines at the same rated speed as a short-stroke design have higher mean piston speed and thus potentially faster liner wear. In practice, modern plateau-honing and sulphur-adaptive cylinder lubrication have allowed ultra-long-stroke engines to achieve 24,000 to 30,000 hours between liner replacement overhauls, comparable to or better than 1990s short-stroke engines. The improvement comes from lubricator timing matched to the scavenge port opening event, which ensures fresh cylinder oil is deposited at exactly the right point in each stroke.

Piston and piston rod overhauls. Piston overhaul interval in modern slow-speed engines is 18,000 to 24,000 hours regardless of stroke, driven by ring pack groove wear and piston crown deposit accumulation. Longer stroke produces longer piston rod life because the piston rod sees lower bending stress (crosshead guide clearances scale with liner length, not absolute stroke in most designs).

Spare parts cost. Larger bore components (liners, pistons, piston crowns, cylinder covers, exhaust valves) cost more to manufacture and hold as spares. A single G95ME-C cylinder cover can cost USD 250,000 to 400,000 as a loose part; a G70ME-C cover is USD 120,000 to 200,000. Owners operating multiple vessels with the same engine family consolidate spares inventories to manage this cost.

For the liner-specific design consequences, including bore-cooling circuit design, plateau honing specifications, and anti-polishing ring geometry, see cylinder liner design for two-stroke engines.

Limitations

Several aspects of bore and stroke selection sit outside the scope of this article:

Four-stroke medium-speed engines follow broadly similar S/B logic, but the practical S/B range is narrower (1.0 to 2.0) because four-stroke valve timing demands more compact combustion chambers and shorter strokes are needed to fit all four valves in a reasonable cover diameter. The propulsion-matching argument for long stroke still holds but is addressed through reduction gearing rather than direct-drive.

Dual-fuel and gas injection geometry. Adding gas admission valves or pilot injection systems to the cylinder cover changes the combustion-chamber design around a given bore. WinGD’s X-DF series and MAN’s ME-GI series both accommodate this within existing bore families, but the gas admission valve placement is a bore-dependent design element this article does not cover in detail.

Torsional vibration. Cylinder count and bore (which sets crankshaft throw mass) both enter the torsional vibration analysis. Adding a cylinder to a 6-in-line raises the firing frequency and can excite the propeller-shaft natural frequency. This is a separate analysis outside engine geometry selection.

Propeller diameter and aperture clearance. The propulsion system argument made here assumes that a larger propeller can be fitted as rpm falls. In practice, hull aperture, classification society tip clearance requirements (typically 15 to 25 percent of propeller radius from the hull), and propeller-induced vibration limits constrain diameter. The bore and stroke selection is only valid if the propeller design is physically feasible in the given hull form.

Cold corrosion at low BMEP. Ultra-long-stroke engines operating at very low loads (below 15 to 20 percent MCR) face sulphuric acid condensation on the liner because exhaust gas temperatures fall below the acid dew point. This is a cylinder lubrication management issue, not a geometry issue per se, but it is aggravated by the low scavenging pressure inherent at those loads in a long-stroke engine.

Bore and stroke in new-build specification and SFOC guarantees

When a shipyard places an order with a main engine maker, the bore, stroke, cylinder count, and rated speed are fixed in the main engine purchase order. The maker then generates a project-specific engine programme (PSEP) that documents the exact layout diagram, the SMCR, the service rating, and the guaranteed SFOC at the specified tuning point.

The SFOC guarantee is key. For a VLCC or a large container ship, the difference between 160 g/kWh and 155 g/kWh at the service rating represents roughly 4,000 to 7,000 tonnes of fuel per year across a typical operating profile. At USD 600/tonne, that is USD 2.4 to 4.2 million annually. The bore-stroke geometry is baked in at this stage; it cannot be changed post-delivery. This is why the propulsion system analysis that selects the bore and stroke belongs in the concept design phase, not the detailed design phase.

MAN Energy Solutions documents the SFOC sensitivity to engine load in its ME Tuning Guide (part of the Instruction for Projection series). For the G-series at a tuning point of 75 percent MCR, SFOC is typically 5 to 8 g/kWh lower than at 100 percent MCR. This is exploited by setting the SMCR at a power slightly above the expected steady-state service power, so the continuous operating point falls in the SFOC trough. The choice of SMCR position on the layout diagram is therefore a fuel-economy decision as much as a power-margin decision.

For the SFOC-versus-load curve and correction factor methodology, see specific fuel oil consumption curves. The engine thermal efficiency calculator converts SFOC to brake thermal efficiency (BTE) for direct comparison across engine families.

Bore nomenclature in MAN and WinGD model designations

Understanding how bore is embedded in model names removes ambiguity when comparing engine specifications from different sources.

MAN Energy Solutions uses a prefix letter followed by bore in centimetres: “G70” means G-series, 700 mm bore. “S90” means S-series, 900 mm bore. The suffix encodes the generation: “ME-C” is electronically controlled, “ME-GI” adds gas injection, “ME-LGIM” is dual-fuel with methanol or LPG injection. The digit string after the hyphen is the generation number; “C10.5” denotes the 10th series, 0.5 iteration. So “G80ME-C10.5” is: G-series (ultra-long-stroke), 800 mm bore, electronically controlled, 10th generation, 0.5 iteration.

WinGD uses “X” plus bore in decimetres: “X82” is X-series, 820 mm bore. “X-DF” adds dual-fuel. The earlier RT-flex designation (from the Wartsila-owned era) used “RT-flex” plus bore in centimetres: “RT-flex82C” is the predecessor to the X82.

These naming conventions mean you can read the bore directly from the model name in most cases, which simplifies the swept-volume calculation: bore is the number in the model name (in mm for MAN, in mm from the decimetres prefix for WinGD). Stroke must be read from the project guide or the engine programme table; it is not in the model name.

For a worked sizing exercise: a ship requires 45,000 kW at 78 rpm from a 7-cylinder engine. Target BMEP is 20.5 bar. Mean piston speed at 78 rpm with the G-series stroke ceiling of 3.5 m is $c_m = 2 \times 3.5 \times 78/60 = 9.1 \text{ m/s}$, slightly above the 8.5 m/s limit. Reducing stroke to 3.2 m gives $c_m = 8.32 \text{ m/s}$, acceptable. Swept volume per cylinder is $V_s = 45000 \times 10^3 / (20.5 \times 10^5 \times 7 \times 78/60) = 2.52 \text{ m}^3$. With stroke 3.2 m: $d = \sqrt{4 \times 2.52 / (\pi \times 3.2)} = 1.00 \text{ m}$. That’s 1,000 mm bore, at the manufacturing limit; the designer would revisit cylinder count or accept higher BMEP. Increasing to 8 cylinders: $V_s = 2.20 \text{ m}^3$, $d = \sqrt{4 \times 2.20 / (\pi \times 3.2)} = 0.935 \text{ m}$, close to the G95ME bore family. This is the iterative process documented in MAN’s Basic Principles of Ship Propulsion.

See also

• Engine load diagram and operating envelope
• Two-stroke marine diesel engine fundamentals
• Cylinder liner design for two-stroke engines
• Engine power and BMEP relationships
• Propeller theory: open-water efficiency
• What is EEDI?
• What is EEXI?
• Specific fuel oil consumption
• Specific fuel oil consumption curves

Related calculators:

• Engine mean piston speed
• Engine BMEP calculator
• Engine MCR derating
• EEXI attained
• EPL required MCR
• MAN ES G95ME-C9.5
• WinGD X92