Four-Stroke Marine Diesel Engine Fundamentals

Four-stroke marine diesel engines complete a full thermodynamic cycle across four piston strokes and two crankshaft revolutions: intake, compression, power, and exhaust. Each cylinder fires once every two revolutions, in contrast to two-stroke marine diesel engines, which fire every revolution. The four-stroke architecture uses trunk-piston construction: the connecting rod attaches directly to the piston via a gudgeon pin, eliminating the crosshead assembly of slow-speed two-strokes and producing a shorter, lighter engine that runs at medium speed (300 to 1,000 rpm) or high speed (above 1,000 rpm). Medium-speed four-strokes from Wartsila, MAN Energy Solutions, Caterpillar (MaK), Bergen, and Hyundai HiMSEN power cruise ships, ferries, offshore vessels, and tugs through reduction gearboxes or diesel-electric arrangements. High-speed four-strokes from MTU and Cummins serve fast ferries and naval vessels. On large ships with slow-speed two-stroke main engines, four-stroke engines are the universal choice for auxiliary generating sets. Related tools on this site include the engine BMEP calculator, the compression ratio calculator, and the mean piston speed calculator.

Background and context

Why four-stroke engines occupy the medium-speed niche

The propulsion physics of large merchant ships favor slow rotational speed: a propeller sized for a 300,000-DWT tanker operates most efficiently at roughly 80 to 95 rpm. A two-stroke marine diesel at that speed can deliver 60 to 87 MW directly to the shaft. A four-stroke at the same 80 rpm would produce only half the power strokes per revolution and would be structurally impractical at that cylinder size. So the slow-speed two-stroke wins for large direct-drive propulsion.

Medium-speed four-strokes win in the range below about 15 to 20 MW per engine where gearboxes are affordable and where the ship’s layout demands multiple engines in parallel. A cruise ship with 80 MW total power installed across six or eight medium-speed engines can shut down two engines in harbor and still have full hotel power; a single massive two-stroke cannot offer that operational flexibility. Offshore support vessels, which need to shift rapidly between transit and dynamic-positioning modes, exploit the same multi-engine architecture. Ferries, tugs, and OSVs in the 5 to 30 MW range find medium-speed four-strokes more cost-effective than slow-speed two-strokes both in capital cost and in the shorter maintenance intervals that match their port-intensive duty cycles.

Brief historical placing

Rudolf Diesel demonstrated the compression-ignition cycle in 1897, and four-stroke diesel engines entered marine service in cargo and passenger ships by the 1910s. The slow-speed two-stroke emerged in the 1920s and progressively claimed the large direct-drive propulsion market after the 1950s. The four-stroke held the auxiliary generating set and medium-speed propulsion roles throughout that period. The 1970s and 1980s saw significant bore and BMEP increases: the Wartsila 46 (introduced 1986, 460 mm bore) set a benchmark for high-output medium-speed engines at around 1,050 kW per cylinder. Turbocharger efficiency improvements from the 1990s onward allowed higher boost pressures, higher BMEP, and consequently better specific fuel oil consumption (SFOC), narrowing the efficiency gap with slow-speed two-strokes.

The four-stroke cycle in detail

A four-stroke engine converts chemical energy in fuel to rotational work across four distinct piston strokes. The total crank rotation for one complete cycle is 720 degrees of crankshaft angle (°CA), two full revolutions.

Stroke 1: intake (induction)

The piston descends from top dead centre (TDC) to bottom dead centre (BDC). The intake valve opens before TDC on the previous exhaust stroke, typically 10 to 30°CA before TDC; this early opening is the start of the valve overlap period. As the piston descends, the cylinder volume increases, drawing in fresh charge air. On all marine four-strokes in propulsion or generating service, this air is delivered at elevated pressure by a turbocharger, typically 2.5 to 4.5 bar absolute at full load. The intake valve closes after BDC, typically 20 to 50°CA after BDC, trapping more air mass than the geometric cylinder volume would contain at atmospheric pressure; this is positive intake-valve closing and is called late inlet valve closing (LIVC) in some variable-timing applications.

The mass of air trapped per cylinder per cycle is the parameter that determines the maximum fuel that can be burned and therefore the maximum power per cylinder. If $m_a$ is the trapped air mass and $\alpha$ is the air-to-fuel ratio (typically 20 to 30 by mass for four-stroke marine diesels at full load), then the maximum fuel mass per cycle is $m_f = m_a / \alpha$. The indicated mean effective pressure (IMEP) is bounded by this fuel mass and the lower heating value of the fuel.

Stroke 2: compression

With all valves closed, the piston ascends from BDC to TDC, compressing the trapped air. The compression ratio $r_c$ is defined as:

where $V_{BDC}$ is the cylinder volume at BDC (total volume: swept plus clearance) and $V_{TDC}$ is the clearance volume at TDC. Medium-speed four-stroke marine diesels typically run geometric compression ratios of 13:1 to 16:1. The compression ratio calculator computes this from bore, stroke, and clearance volume measurements.

The compression pressure at TDC (the compression pressure $P_{comp}$) is related to the inlet manifold pressure $P_1$ by the isentropic relationship:

where $\gamma$ is the ratio of specific heats for air (approximately 1.35 for the elevated temperatures in a diesel cylinder). At $P_1$ = 3.0 bar absolute and $r_c$ = 14, this gives roughly 115 to 120 bar, consistent with measured compression pressures on modern medium-speed engines. Temperature at TDC under isentropic compression reaches 700 to 900 K, well above the auto-ignition temperature of diesel fuel (approximately 520 to 580 K), so the injected fuel ignites without a spark. Fuel injection begins 5 to 15°CA before TDC, with ignition delay typically 0.5 to 2.0 ms.

Stroke 3: power (combustion and expansion)

This is the only stroke that produces net work. Fuel injected near TDC ignites and burns, raising pressure rapidly to a peak firing pressure $P_{max}$ that for modern medium-speed engines is 180 to 230 bar. The piston is forced downward by the combustion gas pressure, transmitting force through the connecting rod to the crankshaft. The exhaust valve opens before BDC, typically 40 to 60°CA before BDC; this early opening terminates the expansion and begins the blowdown phase where the high-pressure exhaust gas discharges into the exhaust manifold. The timing is chosen to balance the work lost from early expansion termination against the pumping work saved by not pushing hard against exhaust backpressure during the early exhaust stroke.

The ratio $P_{max}/P_{comp}$ (also called the pressure ratio across TDC) is a key design metric: a high ratio indicates efficient combustion timing but increases mechanical loading on the connecting rod, bearings, and cylinder head. Modern medium-speed engines typically target $P_{max}/P_{comp}$ in the range 1.6 to 1.9.

The brake mean effective pressure (BMEP) relates the net cycle work to the cylinder swept volume $V_s$:

where $T_{brake}$ is the brake torque and $n_{cyl}$ is the number of cylinders. For medium-speed four-strokes the factor of $2\pi$ (rather than $\pi$ for two-strokes) reflects the two-revolution cycle. Modern medium-speed engines achieve BMEP of 25 to 34 bar. The BMEP calculator computes this from power, speed, and displacement. At the high end, the Wartsila 31 achieved 30.4 bar BMEP at its 2016 introduction, at the time the highest BMEP of any medium-speed four-stroke engine in production.

Stroke 4: exhaust

The piston ascends from BDC to TDC with the exhaust valve open. Combustion products are displaced upward and out through the exhaust valve into the exhaust manifold, then through the turbine of the turbocharger. The exhaust valve closes near or slightly after TDC on the intake stroke, creating the valve overlap period.

Valve overlap and its function

Valve overlap is the period, measured in degrees of crankshaft angle, during which both the intake valve and the exhaust valve are simultaneously open. Overlap runs from the opening of the intake valve (before TDC on the exhaust stroke) to the closing of the exhaust valve (after TDC on the intake stroke). On a turbocharged medium-speed four-stroke, overlap typically spans 30 to 80°CA.

During overlap, the pressure difference between the charged intake manifold (at turbocharger boost pressure) and the exhaust manifold drives fresh air through the cylinder. This cross-flow performs two functions simultaneously: it scavenges residual exhaust gas from the clearance volume, reducing dilution of the next charge; and it cools the piston crown, cylinder head, and exhaust valve. The cooling benefit is not trivial. Exhaust valve temperatures on medium-speed four-strokes can reach 700 to 800°C at the stem and seat; the overlap air flow holds the valve seat at a temperature that sodium-cooled and Nimonic alloy valves can sustain across their operating life.

The p-V diagram of the four-stroke cycle shows the intake stroke as a nearly horizontal expansion at slightly below intake manifold pressure, the compression stroke as a steep pressure rise, the power stroke as a near-isentropic expansion with a combustion pressure spike near TDC, and the exhaust stroke as a return to near-atmospheric pressure. The enclosed area above the BDC line represents net indicated work per cycle; the area below (the pumping loop) is the work exchanged with the gas exchange system and is slightly positive on a turbocharged engine because boost pressure exceeds exhaust backpressure at full load.

Trunk-piston construction

The piston, connecting rod, and gudgeon pin

The defining feature of four-stroke marine diesel engines is the trunk-piston arrangement. The piston crown, which directly faces the combustion, connects through a piston skirt to a gudgeon pin (wrist pin). The connecting rod big end attaches to the crankshaft journal; the small end is fitted with a bush through which the gudgeon pin passes. This layout means the connecting rod’s angular motion applies a lateral force component to the piston: the resultant pushes the piston skirt against the cylinder wall on what is called the thrust side (the side opposite the crankshaft throw at TDC).

The piston skirt is therefore a structural load-bearing element that distinguishes the trunk-piston from the crosshead. In a crosshead engine, an independent crosshead block and sliding guide absorb this lateral force, keeping the piston rod vertical and isolating the piston from side loads. The trunk-piston simplification makes the engine shorter (no crosshead space needed), lighter, and less expensive to build, but it transfers the side-load friction to the cylinder wall, slightly increasing ring-pack wear and requiring the crankcase lubricant to serve as cylinder lubricant simultaneously.

Oil system: one circuit for all duties

The single-oil-system consequence of trunk-piston construction has practical implications. The crankcase oil (typically SAE 40 or SAE 30, with a total base number of 10 to 30 mg KOH/g depending on fuel sulphur content) must simultaneously lubricate the main bearings, bottom-end bearings, gudgeon pins, valve gear, and cylinder walls. The cylinder walls contaminate the oil with combustion blow-by products (sulphur acids, soot, partial combustion products) and with any unburned fuel during cold starting. Lube oil consumption through ring-pack leakage into the combustion space is typically 0.2 to 0.4 g/kWh on a well-maintained medium-speed four-stroke, an order of magnitude lower than the cylinder oil consumption of a slow-speed two-stroke.

Cylinder block, liner, and head

The cylinder block of a medium-speed four-stroke is a single cast iron or ductile iron structure housing all the cylinder bores, the main bearing housings, the camshaft tunnel, and the coolant passages. Cylinder liners are cast iron sleeves pressed or shrunk into the block bores; they carry the honed running surface on which the piston rings seal. The cylinder head bolts to the top of the block over each cylinder and houses the intake and exhaust valve seats, the fuel injector (or injectors), and the charge air and jacket water passages.

V-configuration engines (8V, 12V, 16V, 18V, and 20V cylinder arrangements) mount two rows of cylinders at a 45° or 60° included angle on a shared crankshaft. The crankshaft in a V-engine has adjacent throws offset to fire the paired cylinders at the correct crank intervals; the firing order alternates between the left and right banks. V-configuration engines deliver more power in a shorter overall length, which is important in engine rooms with transverse length constraints.

Crankshaft and torsional vibration

The crankshaft of a medium-speed four-stroke engine carries crank throws separated by main bearing journals. For a six-cylinder in-line engine the firing order produces power strokes at 120°CA intervals, giving smooth torque delivery. For engines with fewer cylinders, or for V-configurations with paired cylinder banks, torsional vibration resonance in the shafting must be calculated and addressed. Classification societies (Lloyd’s Register, DNV, Bureau Veritas) require a torsional vibration analysis under their Rules for classification; barred speed ranges may be imposed on the operating speed envelope to avoid resonances that would exceed allowable stress. The engine mean piston speed calculator is useful for confirming that mean piston speed at the barred range limits stays within the manufacturer’s mechanical rating.

Camshaft and valve train

A four-stroke engine requires a dedicated valve train because the intake and exhaust valves must open and close once per cycle at precisely timed intervals. The camshaft rotates at exactly half the crankshaft speed (one camshaft revolution per two crankshaft revolutions per cycle). Cams with specific lift profiles drive the valves through rocker arms, push rods, or, on modern overhead-camshaft designs, directly through followers.

Most medium-speed four-stroke marine engines use an overhead camshaft arrangement (camshaft in or above the cylinder head) rather than a cam-in-block with push rods. The overhead arrangement reduces valve train mass and improves valve control accuracy at the higher speeds where push-rod deflection and follower bounce become significant. Each cylinder typically carries four valves (two intake, two exhaust), with the larger combined seat area improving charge flow compared with a two-valve design.

The camshaft also drives the fuel injection timing. On conventional cam-actuated systems, a fuel pump element sits on a cam lobe timed to inject before TDC; the injection timing is mechanically fixed relative to the crankshaft. Modern common-rail systems (Wartsila common-rail on the W31, W32, W46F; MAN’s common-rail CR on the 32/44CR and 48/60CR) replace the fuel cam with a high-pressure accumulator rail and electronic injectors, decoupling injection timing from the camshaft. This allows independent optimization of injection start, duration, and pressure across the full load range, improving SFOC at part load and supporting alternative fuel injection sequences.

Turbocharging and charge-air cooling

All medium-speed marine diesel engines running at propulsion power ratings above approximately 500 kW are turbocharged. Turbocharging is not optional at these power densities; without boost, the trapped air mass per cycle could not support the fuel burn rates that produce 25 to 34 bar BMEP.

The turbocharger on a four-stroke

The turbocharger sits in the exhaust-gas stream. Exhaust leaving the cylinder at 400 to 550°C and 1.1 to 2.0 bar absolute drives the turbine wheel, which is shaft-coupled to a centrifugal compressor. The compressor draws in ambient air and delivers it to the intake manifold at 2.5 to 4.5 bar absolute and 150 to 200°C after compression. This heated air passes through a charge-air cooler (also called an intercooler or aftercooler) before entering the cylinders. See the marine engine turbocharging article for the full turbocharger design context.

Charge-air cooling

Cooling the charge air after compression serves two purposes. First, denser (cooler) air contains more oxygen per unit volume, increasing the trapped oxygen mass and allowing more fuel burn per cycle. Second, lower charge temperature at the start of compression reduces the peak firing temperature, which is the primary driver of thermal NOx formation. A typical charge-air cooler reduces air temperature from 160 to 200°C at the compressor outlet to 35 to 50°C at the cylinder intake valve, which is close to seawater coolant temperature. Some advanced engines use a two-stage cooling arrangement to approach seawater temperature more closely.

At part load, the turbocharger delivers less boost pressure because there is less exhaust energy to drive the turbine. This creates a problem for the four-stroke cycle: at low load, the BMEP drops steeply and SFOC rises because the cycle becomes less efficient. Two approaches address this. Variable turbine geometry (VTG) adjusts the turbine nozzle area to maintain higher boost pressure at part load. Two-stage turbocharging uses a high-pressure and a low-pressure turbocharger in series to achieve very high overall pressure ratios at full load while maintaining acceptable boost at part load.

Miller cycle and late intake valve closing

High-efficiency medium-speed engines apply the Miller cycle by closing the intake valve well after BDC, typically 40 to 80°CA after BDC (late intake valve closing). The piston ascends against the trapped charge, doing work to push air back out through the still-open intake valve before the valve eventually closes; the effective compression stroke is shorter than the geometric stroke. This reduces the compression temperature and therefore the NOx formation temperature without requiring exhaust gas recirculation. The turbocharger must deliver correspondingly higher boost pressure to compensate for the reduced trapping efficiency. The Wartsila 31 uses an aggressive Miller cycle with early closing that delivers SFOC of 165 to 170 g/kWh at optimal load, among the lowest for a four-stroke engine in production.

Performance data and comparison with two-stroke engines

Four-stroke vs. two-stroke: a direct comparison

The choice between four-stroke and two-stroke is primarily decided by ship type, propulsion architecture, and power level rather than by any single technical superiority.

The SFOC advantage of two-strokes (typically 10 to 20 g/kWh lower at optimal load) is real and matters on a large ship burning 50 to 100 tonnes of fuel per day: a 15 g/kWh reduction at 20 MW is roughly 300 kg of fuel per day, which at VLSFO prices of $600 to$700/tonne is $180 to$210 per day in fuel savings. At that scale, the two-stroke wins economically for long-voyaging ships. But for a cruise ship where fuel cost per berth-day is less important than the flexibility of multi-engine operation, or for an OSV where the propulsion load varies by a factor of 10 between DP operation and transit, the four-stroke multi-engine arrangement wins on operability grounds.

The gearbox penalty and diesel-electric alternative

Coupling a medium-speed engine to a fixed-pitch propeller through a reduction gearbox incurs a loss of 1.5 to 3% of shaft power in the gearbox, plus the mechanical complexity and cost of the gearbox itself. Gearbox reduction ratios for medium-speed engines are typically 3:1 to 10:1; a 720 rpm engine driving a propeller at 120 rpm needs a 6:1 reduction ratio. The gearbox must also transmit the full rated torque continuously; a 10 MW engine at 720 rpm produces a crankshaft torque of about 133 kN·m, which the gearbox multiplies to roughly 800 kN·m at the propeller shaft at 6:1 reduction.

Diesel-electric arrangements avoid the gearbox entirely by converting mechanical power to electrical power through a shaft-coupled generator, then driving an electric motor at the propeller. The electrical conversion and motor losses are typically 6 to 10% of rated power, worse than a gearbox, but diesel-electric offers operational advantages that often outweigh the efficiency penalty: the engines run at constant speed regardless of propeller demand, the propulsion can be cut to electric-only at low loads, and the main engines can power both propulsion and hotel loads from a common bus. Cruise ships and icebreakers commonly use diesel-electric. Platform supply vessels use it for the load-sharing flexibility in DP mode.

Fuel flexibility and dual-fuel engines

Conventional HFO and VLSFO

Medium-speed four-strokes accept heavy fuel oil (HFO) in the same way as slow-speed two-strokes: the high jacket water temperature (80 to 90°C) and the low-sulphur lube oil with sufficient alkalinity reserve handle the acidic combustion products. Following IMO 2020 (MARPOL Annex VI Regulation 14), ships outside scrubber-equipped vessels must burn VLSFO (max 0.50% S) globally, or 0.10% S distillate in Sulphur Emission Control Areas (SECAs). The shorter combustion duration in a four-stroke at higher rpm places more demand on fuel atomization quality; fuel viscosity at the injector must be maintained in the range 10 to 20 cSt, achieved by heating heavier residual fuels.

Dual-fuel LNG medium-speed engines

The Wartsila DF (dual-fuel) series, introduced with the Wartsila 34DF in 1995 and extended to the 20DF, 50DF, and W31DF, uses Otto-cycle gas combustion: natural gas is admitted with the intake air at low pressure (typically 4 to 6 bar gauge), mixed homogeneously, and ignited by a small diesel pilot injection at TDC. The engine runs as a conventional diesel in diesel mode if gas supply is interrupted. In gas mode, the lean premixed combustion temperature is lower than diesel combustion, reducing thermal NOx formation to below 1.3 g/kWh, meeting IMO Tier III in most NECAs without additional aftertreatment.

The MAN 51/60DF uses a similar Otto-cycle approach at the high-output end (510 mm bore, 600 mm stroke, rated to 1,650 kW per cylinder). It entered service on cruise ships from 2019.

Methane slip (unburned methane passing through the exhaust) is the key environmental limitation of the Otto-cycle gas mode: methane is 80 times more potent than CO2 over a 20-year global warming potential horizon. Manufacturers report methane slip of 0.5 to 3 g/kWh depending on load and design. At low loads, slip rises because the lean mixture approaches the flammability limit and not all the gas burns. This makes gas-mode operation at low part load less environmentally beneficial than it appears from the CO2 perspective.

Methanol and ammonia capability

MAN Energy Solutions has introduced methanol-capable versions of the 32/44CR and other medium-speed engines through its GenSets ME-M program. Methanol requires about 2.3 times the volume flow of diesel for equivalent energy (lower heating value of methanol: 19.9 MJ/kg versus 42.7 MJ/kg for diesel); injection equipment must be sized accordingly. The advantage is methanol’s near-zero sulphur content and low particulate matter compared with HFO.

Ammonia as a four-stroke fuel is in development at multiple manufacturers. Ammonia’s narrow flammability range (15 to 28% by volume in air) and low flame speed require specific combustion chamber geometries and ignition energy; the pilot fuel fraction must be sufficient to reliably ignite the ammonia mixture across all operating conditions. The NOx challenge with ammonia is different: fuel-bound nitrogen in ammonia combustion produces fuel NOx through a separate reaction pathway from the thermal NOx that drives Tier III limits.

Marine applications

Cruise ships: diesel-electric with medium-speed engines

Modern cruise ships use the diesel-electric configuration with four to eight medium-speed engines driving generators that feed a common power bus. An Oasis-class ship (Royal Caribbean) has six 16-cylinder Wartsila 46F engines plus three 12-cylinder Wartsila 46F engines, totaling approximately 97 MW of installed generating capacity. The propulsion motors and bow thrusters draw from the same bus; the engines start and stop based on demand, keeping loaded engines in a fuel-efficient range. No reduction gearbox exists in the propulsion chain; the electric motor has inherent variable-speed capability.

Ferries and ro-pax ships

Ferries crossing fixed routes on tight schedules use medium-speed engines through reduction gearboxes to controllable-pitch propellers (CPPs), which change propeller blade pitch to adjust thrust without changing engine speed. This is efficient at high throughput; the engine runs near its optimal speed and SFOC point regardless of whether the ship is maneuvering into port or crossing at full speed. The MAN 32/44CR and MAN 48/60CR are common choices for ferry propulsion in the 10 to 30 MW range.

Offshore support vessels and DP systems

Offshore support vessels operate in dynamic positioning (DP) mode for much of their working life, holding station precisely while a cargo transfer or well intervention is in progress. DP requires responsive and redundant power; losing propulsion unexpectedly could allow the vessel to drift into a riser or an installation. Medium-speed diesel-electric with typically four to six engines and redundant switchboards provides the necessary redundancy. Vessels operating under DP Class 2 or DP Class 3 certification must design their power generation so that loss of any one compartment does not reduce DP capability below the design environmental condition.

Tugboats and workboats

Tugs rarely need sustained power over long voyages; they need high torque at low ship speed during bollard pull and then moderate power during transit. Two medium-speed engines through a reduction gearbox to twin shafts or a Z-drive arrangement is the standard. The Bergen B33:45 (315 mm bore) and the Caterpillar 3512 are common in the 1 to 4 MW range per engine for harbor and coastal tugs. High-speed four-stroke engines from MTU (Series 4000) at 1,100 to 3,500 rpm serve fast rescue boats and naval auxiliary craft.

Auxiliary generating sets on two-stroke-powered ships

The most numerically common application of the four-stroke marine diesel is the auxiliary generating set. A large container ship with a slow-speed two-stroke main engine carries three or four gensets, each a four-stroke diesel engine coupled to an alternator at fixed speed (typically 720 or 900 rpm for 60 Hz, or 750 rpm for 50 Hz generation). Genset outputs range from 500 kW to 3,000 kW per unit on large vessels. The genset engines are nearly always in-line configurations for compactness; the MAN L21/31, Wartsila 20, and Caterpillar C32 are among the most common. The system auxiliary engine medium-speed 4-stroke calculator covers genset selection and load analysis.

A “hotel load” of 2 to 8 MW is typical for a large container ship at sea: refrigeration of cargo, accommodation, lighting, navigation, and deck machinery. In port, the hotel load may double because cargo cranes, ventilation, and port-side services run simultaneously. Class societies, including Lloyd’s Register and DNV, require that at least one genset can start automatically and take load within 30 seconds of a blackout; the emergency genset high-speed diesel calculator covers the emergency generating set sizing requirement.

IMO emissions framework

NOx: Regulation 13 and the three-Tier structure

MARPOL Annex VI Regulation 13 establishes NOx emission limits for marine diesel engines based on the engine’s build date and the operational area. The limits are expressed in grams of NOx per kilowatt-hour of output, at the test cycle weighted over the speed and load points of the applicable MARPOL Annex VI test cycle (E2, E3, D2, C1, or similar, depending on engine type and application).

Tier I applies to engines built before 1 January 2011. The limit is a speed-dependent formula: at 130 rpm it is 17.0 g/kWh, decreasing to 12.0 g/kWh at n = 2,000 rpm and above, with a log-linear interpolation in between. Most medium-speed four-strokes at 720 rpm sit in the range 12 to 13 g/kWh for Tier I.

Tier II applies to engines built from 1 January 2011, with limits approximately 20% lower than Tier I.

Tier III applies in NOx Emission Control Areas (NECAs: the North American ECA including the Caribbean, and the North Sea/Baltic from 1 January 2021) to engines built from 1 January 2016. The limit is approximately 80% lower than Tier I: at 720 rpm, roughly 2.4 g/kWh. Meeting Tier III on a diesel-mode four-stroke requires either selective catalytic reduction (SCR) after the turbocharger turbine, or exhaust gas recirculation (EGR) to reduce combustion temperature. Dual-fuel gas-mode combustion on engines like the Wartsila 34DF achieves Tier III inherently through the lean, cool premixed combustion. See the NOx Tier I-II-III and selective catalytic reduction articles for the regulatory and technical details.

SOx: Regulation 14 and scrubbers

Regulation 14 limits the sulphur content of fuel oil. The 0.50% global cap since 1 January 2020 (IMO 2020) applies to all vessels. In SECAs, the limit is 0.10%. Four-stroke engines burning VLSFO or distillate comply directly; those with open-loop scrubbers may continue burning HFO if port restrictions do not apply. The four-stroke engine itself has no inherent SOx advantage or disadvantage over a two-stroke burning the same fuel.

Carbon intensity and CII

The Carbon Intensity Indicator (CII), mandated under MARPOL Annex VI regulations as amended by MEPC.338(76) and MEPC.339(76), rates ships annually on their CO2 emissions per cargo tonne-nautical mile. A ship’s CII rating (A through E) feeds into the Enhanced Ship Energy Efficiency Management Plan (SEEMP III) requirement. Four-stroke engines on cruise ships and ferries face CII challenges because these vessel types carry passengers rather than cargo, and the denominator (gross tonnage multiplied by nautical miles) does not directly reflect revenue productivity the way cargo-tonne-miles would for a container ship.

Operating and maintenance considerations

Overhaul intervals and cylinder condition monitoring

Medium-speed four-stroke engines are designed for shorter but more predictable overhaul intervals than slow-speed two-strokes. Top-end overhauls (cylinder heads, exhaust valves, fuel injectors, charge-air cooler cleaning) occur every 8,000 to 18,000 running hours depending on manufacturer, fuel type, and load profile. Bottom-end overhauls (main bearings, connecting rod bearings, gudgeon pins) are at 20,000 to 40,000 hours. The shorter overhaul interval than slow-speed two-strokes is partly a consequence of higher rotational speed (more cycles per hour) and partly the absence of separate cylinder lubrication.

Genset engines on ships with two-stroke main engines often accumulate 20,000 to 25,000 running hours between dry dockings over a five-year classification period; the maintenance strategy must account for this without access to shore-side workshops.

Load management and SFOC curves

A medium-speed four-stroke achieves its minimum SFOC at roughly 75 to 85% of its continuous service rating. Below 50% load, SFOC rises steeply, sometimes by 20 to 30 g/kWh compared with the optimum, because the turbocharger operates inefficiently at low exhaust energy, reducing boost pressure and worsening the volumetric efficiency of the cycle. This is a stronger part-load penalty than a large slow-speed two-stroke experiences because the two-stroke’s long stroke maintains better scavenging down to lower loads.

Diesel-electric configurations partially mitigate this by shutting down engines as load falls, keeping the remaining engines operating at a higher fraction of their individual ratings.

Thermal management and jacket water

Jacket water temperature on medium-speed four-strokes is typically 70 to 90°C at the outlet. Higher jacket water temperature improves thermal efficiency (the engine approaches a higher top temperature in the Rankine sense of waste heat), reduces fuel condensation on liner walls, and allows jacket water heat recovery for fuel heating, accommodation heating, or fresh water generation. The cooling water system on a medium-speed engine balances these thermal goals against the temperature limits of the cylinder liner, piston crown, and cylinder head materials.

Limitations

Four-stroke medium-speed marine diesel engines carry several limitations that practitioners must recognize.

The two-revolution firing cycle means that, for a given power output and rotational speed, a four-stroke engine needs roughly twice the displacement of a two-stroke at the same speed, because it fires half as often. At propulsion power levels above 15 to 20 MW per engine, this results in very large and heavy engine packages that a gearbox must connect to the propeller at additional cost and loss.

The trunk-piston single-oil-system design constrains cylinder lubricant performance. The crankcase oil must serve bearings (needing a lower base number and stable oxidation resistance) and cylinder walls (needing sufficient alkalinity to neutralize sulphur acids from combustion). High-sulphur fuels cannot be accommodated simply by increasing the lubricant base number as with slow-speed two-stroke cylinder oils, because the same oil also lubricates sensitive bearing white-metal surfaces that are degraded by highly alkaline lubricants.

Valve gear complexity compared with a two-stroke is real: camshaft, followers, rocker arms, pushrods (on some designs), and valve springs all require regular inspection. Exhaust valve seat and spindle wear are the most common cause of cylinder head removal on medium-speed engines; valve seat face hardening (Stellite or similar) extends service intervals but adds cost.

The part-load SFOC penalty, as noted above, is steeper than for a slow-speed two-stroke. For applications with sustained low-load operation (ferry maneuvering, OSV standby DP), the fuel cost at part load must be included in life-cycle cost comparisons.

Methane slip in dual-fuel gas-mode operation is an emerging regulatory concern. If IMO introduces a life-cycle greenhouse gas metric that includes methane, dual-fuel medium-speed engines operating at low-load gas mode could face compliance challenges beyond current NOx Tier III requirements.

Relevant calculators

The following calculators on this site apply directly to four-stroke engine work:

• Engine BMEP calculator: compute BMEP from power, speed, and cylinder geometry.
• Compression ratio calculator: compute compression ratio from bore, stroke, and clearance volume.
• Mean piston speed calculator: verify mean piston speed at any engine speed.
• Engine BMEP from output data: cross-check from measured torque.
• System auxiliary engine medium-speed 4-stroke: genset sizing and load analysis.
• Emergency genset high-speed diesel: emergency generating set sizing.
• Engine combustion air flow: charge air mass flow at any load point.
• Engine cylinder balance: cylinder-to-cylinder pressure spread assessment.

See also

• Two-stroke marine diesel engine fundamentals: the parallel slow-speed architecture.
• Trunk-piston engine architecture: deep technical treatment of trunk-piston construction.
• Medium-speed four-stroke marine engines: the 300 to 1,000 rpm engine class in detail.
• High-speed four-stroke marine engines: engines above 1,000 rpm for fast craft and naval vessels.
• Marine diesel engine: the broader marine diesel context.
• Marine engine turbocharging: turbocharger design and matching.
• Crosshead diesel engine architecture overview: the two-stroke crosshead in detail.
• NOx Tier I-II-III: the MARPOL Annex VI NOx regulatory framework.
• Selective catalytic reduction: SCR for Tier III compliance.
• LNG as marine fuel: LNG properties and bunkering context.
• Methanol as marine fuel: methanol combustion and safety.
• Ammonia as marine fuel: ammonia combustion and NOx challenges.
• Specific fuel oil consumption: SFOC concept and calculation.
• Wartsila 31 medium-speed engine: the record-BMEP W31 in detail.
• MAN 32/44CR medium-speed engine: the CR common-rail medium-speed engine.
• Bergen B33:45 medium-speed engine: the Rolls-Royce Bergen medium-speed family.
• HiMSEN medium-speed engine: Hyundai medium-speed engines.