Marine Engine Turbocharging: Principles and Practice

Turbocharging is the single largest contributor to the power density of a modern marine diesel engine. By capturing energy from the engine’s exhaust gases and using it to drive a compressor, the system forces more air into each cylinder than atmospheric pressure alone could supply. More air means more fuel burned per cycle and more work extracted from the same cylinder volume. On a large slow-speed two-stroke engine, the turbocharger raises scavenge air pressure to between 3 and 4 bar absolute, roughly tripling the air mass compared to a naturally aspirated engine of the same bore. That compression gain is also what makes NOx control via Miller timing possible, because the turbocharger must compensate for the volumetric efficiency loss that early valve closing creates.

The companion surge-margin calculator quantifies how close the operating point sits to the surge line at any engine load. The charge-air cooler effectiveness calculator covers the thermal side of the scavenge air system, and the turbocharger axial vs radial selection tool maps turbine type to engine power range. For the scavenge air pressure side of the two-stroke gas exchange, see the engine scavenge pressure calculator.

Historical development

Alfred Buchi patented the principle of exhaust-gas turbocharging in 1905 and demonstrated a working installation on a ship in 1925. Through the 1930s and 1940s, turbochargers were fitted selectively on large four-stroke marine diesels to recover some of the exhaust enthalpy that would otherwise be wasted. The first generation of slow-speed two-stroke installations followed in the 1950s, when the manufacturing tolerances needed for high-speed rotating parts at temperatures above 600 degrees Celsius became achievable in production.

The decisive shift came with the adoption of constant-pressure exhaust manifolds on two-stroke engines in the 1960s and 1970s. By smoothing the exhaust pulses into a steady pressure, designers could match axial turbines to large engines more accurately and raise boost pressure to levels that were not possible with pulse systems. Pressure ratios climbed from under 2:1 in the 1950s to over 4:1 by the 1980s. Today, single-stage units from Accelleron (the company that acquired ABB Turbocharging in 2022) and MAN Energy Solutions reach pressure ratios of 4.5 to 5 on large slow-speed engines; two-stage arrangements for Tier III NOx compliance push that to 8 or above.

The regulatory driver that reshaped turbocharger design most sharply was MARPOL Annex VI Regulation 13 on nitrogen oxides. Tier I limits applied to engines built after 1 January 2000; Tier II tightened the limit for engines built after 1 January 2011; Tier III imposes a roughly 75% reduction from Tier I for engines operating in designated NOx emission control areas (ECAs), currently the North American area and the United States Caribbean Sea area for ships built on or after 1 January 2016. The only in-cylinder route to Tier III without exhaust aftertreatment is extreme Miller timing, which requires boost pressure at the compressor outlet of 6 bar or above, which in turn requires two-stage turbocharging.

The turbocharger rotor: turbine and compressor

A turbocharger contains two aerodynamic machines on a common shaft: a turbine driven by exhaust gas and a compressor driven by the turbine. There is no mechanical connection to the engine crankshaft. The rotor spins on plain journal bearings lubricated by engine lubricating oil, reaching between 6,000 rpm on the largest slow-speed frames and 30,000 rpm on small high-speed units. Tip velocities on the largest compressor wheels approach Mach 1.

Axial turbines

An axial turbine has the working gas flowing parallel to the shaft axis through a series of stator vanes and rotor blades. Stator rows convert the thermal and pressure energy of the hot exhaust into a high-velocity jet directed onto the rotor blades; each rotor stage extracts work and reduces the gas pressure. Axial designs handle large volume flows efficiently, and all large marine turbochargers use them. The Accelleron A100-L and A200-L families for slow-speed two-stroke engines cover engine outputs from 3,000 kW to 28,000 kW per turbocharger in multiple frame sizes. The MAN TCA series covers a similar power range, from 3 MW up to 30 MW per unit. Both use single-stage axial turbines.

The turbine inlet temperature on a slow-speed diesel running on heavy fuel oil is typically 500 to 650 degrees Celsius before the turbine. Nickel-based superalloy blades are standard. The nozzle ring, which forms the turbine inlet, is a consumable item that accumulates deposits from combustion products and is inspected at every major overhaul.

Radial turbines

A radial turbine has the gas entering at the periphery of a wheel and exiting axially. The flow is turned through 90 degrees inside the impeller. Radial designs are simpler to manufacture and adequate in efficiency for smaller flow rates. They dominate on medium-speed engines below approximately 5 MW per turbocharger. The MAN TCR series and MET-ER series from MHI Marine Machinery & Equipment (MHI-MME) are radial-turbine units for four-stroke medium-speed applications.

The centrifugal compressor

The compressor stage is centrifugal (radial) on virtually every marine turbocharger regardless of turbine type. Intake air enters axially at the eye of the impeller, is accelerated by the rotating blades to a high tangential velocity, and exits at the periphery at high kinetic energy. The vaneless or vaned diffuser downstream converts most of that kinetic energy into static pressure. The scroll collects the air and delivers it to the charge-air cooler.

Centrifugal compressors are simpler and more tolerant of inlet distortion than axial compressors, which is why they are used universally in this application. A well-matched centrifugal compressor achieves an isentropic efficiency of 78 to 84% at its design point.

Constant-pressure vs pulse turbocharging

The route the exhaust takes from the engine cylinders to the turbine determines whether the turbocharger system is constant-pressure or pulse.

Constant-pressure systems

In a constant-pressure system, all cylinders exhaust into a single large manifold. The manifold volume is sufficient to damp the individual pressure pulses, so the turbine sees nearly steady flow. The turbine can then be optimised for a single operating point and can use the efficient axial design. Slow-speed two-stroke engines invariably use constant-pressure systems because the near-continuous exhaust flow of a multi-cylinder engine, combined with the large flow rates involved, makes the large manifold practical and the efficiency gains from axial turbines decisive.

The main disadvantage is the loss of the kinetic energy in each exhaust pulse. That energy is wasted in the manifold rather than recovered by the turbine. On smaller engines with few cylinders, this loss can be 5 to 10% of turbine efficiency relative to a pulse system.

Pulse systems

In a pulse system, the exhaust from each cylinder enters the turbine through a short, small-bore pipe sized to preserve the pressure wave from the exhaust blowdown. The turbine is driven by a series of high-energy pulses rather than steady flow. Pulse systems are favoured on four-stroke medium-speed engines in the power range below about 10 MW, and on two-stroke engines when only a small number of cylinders is available. The partial-admission turbine used in pulse systems achieves high instantaneous efficiency on each pulse, even though the time-averaged efficiency may be lower than a constant-pressure axial turbine on a larger engine.

Comparison

Compressor map, surge, and the surge margin

Every centrifugal compressor has a characteristic map: a graph of pressure ratio on the vertical axis against corrected mass flow on the horizontal axis. The map shows a family of constant-speed lines. Two boundaries define the operating range.

The surge line is the left-hand boundary. At any given rotational speed, there is a minimum mass flow below which the compressor cannot sustain stable operation. When flow falls below that minimum, the adverse pressure gradient through the impeller blades becomes too steep, the flow separates aerodynamically, and the pressure downstream temporarily exceeds the pressure the compressor can produce. Gas reverses through the compressor, pressure collapses, flow re-establishes, pressure rises again, and the cycle repeats. This is surge. The frequency is typically 5 to 50 Hz and the mechanical loads on impeller blades, shaft, and thrust bearing are destructive if the condition persists.

The choke line is the right-hand boundary, where the flow passages reach sonic velocity and additional mass flow is impossible regardless of how far the outlet pressure is reduced.

The engine’s air consumption traces a line across the map from idle to maximum continuous rating (MCR). A well-matched turbocharger keeps the engine load line well to the right of the surge line at all loads. The surge margin is typically expressed as the percentage increase in pressure ratio at constant mass flow from the operating point to the surge line:

where $\pi_{surge}$ is the pressure ratio on the surge line at the operating mass flow and $\pi_{op}$ is the actual operating pressure ratio. A surge margin of 10% or above is the practical minimum for a reliable installation; many engine-tuning specifications require 15% or more at part load.

Surge is most likely at part load when the engine flow demand is low but the turbocharger is still spinning at a speed set for a higher load, giving a high pressure ratio at low mass flow. Sequential turbocharger cut-out addresses this directly: by shutting down one turbocharger on a multi-unit installation, the remaining unit receives the full exhaust energy of all cylinders and operates on a more favourable part of its map.

Use the turbocharger surge margin calculator to evaluate the surge margin at different load points when reviewing turbocharger matching for a specific installation.

Charge-air cooler and water mist catcher

Compressing air raises its temperature. An ideal diatomic gas compressed from 1 bar and 35 degrees Celsius to 4 bar reaches approximately 228 degrees Celsius after adiabatic compression alone. Real compressors are less than perfectly efficient and the actual delivery temperature is higher. Air at 200 degrees Celsius and elevated pressure contains far less mass per unit volume than cooled air, which defeats the purpose of turbocharging. The charge-air cooler is the heat exchanger between the compressor outlet and the air receiver that cools the compressed air before it enters the cylinders.

On marine two-stroke engines, the cooler is typically a tube-bundle heat exchanger. Compressed air flows on the shell side over finned tubes; engine freshwater (in a closed loop) flows on the tube side. The freshwater circuit is cooled separately by seawater in a secondary cooler. See the marine sea water cooling systems article for the seawater side of this arrangement.

A well-designed cooler on a modern slow-speed engine reduces the air temperature from approximately 180 to 220 degrees Celsius at compressor outlet to 35 to 50 degrees Celsius at the air receiver. The charge-air cooler effectiveness calculator quantifies the temperature reduction as a function of water and air flow rates and fouling resistance. The pressure drop through the cooler is typically 50 to 150 millibars; values above 200 millibars indicate fouling that is affecting engine air supply.

Water mist catcher

When warm, moist ambient air is compressed and then cooled to near the freshwater temperature, condensation occurs. Liquid water carryover into the air receiver and cylinders causes hydrogen embrittlement of cylinder liner cast iron, accelerated ring and bore wear, and in severe cases hydraulic lock. The water mist catcher (moisture separator) is fitted at the cooler outlet. It uses a combination of baffles, impingement surfaces, and a sump with a drain to remove droplets before they reach the cylinders. Automatic float-controlled or solenoid drain valves discharge accumulated water; their correct operation is verified during every port watch inspection on a well-run engine room.

Air-side fouling of the cooler is accelerated by oil mist from leaking turbocharger seals, particulates, and salt spray in the intake air. The contamination reduces the heat transfer coefficient and raises pressure drop. Cooler cleaning is scheduled at major maintenance intervals, using mechanical brushing or alkaline chemical cleaning of the tube bundle.

Two-stroke vs four-stroke turbocharging arrangements

The turbocharger installation differs between engine types in ways that go beyond the manifold design.

Slow-speed two-stroke engines

Modern slow-speed crosshead engines (MAN ME series, WinGD X series) carry one to three turbochargers depending on cylinder count. A six-cylinder MAN ME-GI running at MCR uses two TCA turbochargers; a twelve-cylinder unit uses three. The turbochargers sit at the exhaust end of the engine, above the exhaust manifold, and discharge compressed air into the scavenge air receiver that runs the full length of the engine block.

Gas exchange in a two-stroke engine depends entirely on the turbocharger. There is no separate intake stroke and no poppet intake valves. The piston uncovers circumferential scavenge ports in the cylinder liner on the downstroke; exhaust valves in the cylinder head open simultaneously. Compressed air from the scavenge air receiver flows in through the ports and pushes exhaust gas out through the head valves. This process requires a scavenge air pressure consistently above 2.5 bar absolute across the full load range to ensure complete gas exchange. See the related articles on uniflow scavenging in two-stroke marine engines and scavenge port geometry and timing for the fluid-dynamics detail.

Auxiliary blowers at low load

Below approximately 35 to 40% of MCR, the turbocharger cannot produce adequate scavenge air pressure. The exhaust energy entering the turbine at low load is too low to drive the compressor to a pressure ratio sufficient for clean scavenging. Most slow-speed two-stroke engines are fitted with two electrically driven auxiliary blowers that operate in parallel with the turbocharger at low engine loads and during manoeuvring. They cut in automatically when scavenge air pressure falls below a set threshold, typically around 1.2 to 1.4 bar absolute, and cut out when pressure recovers on acceleration. The system is essential during starting, harbour manoeuvres, and slow steaming below the auxiliary blower cut-out point. Failure of both blowers at low speed results in scavenge pressure below the gas-exchange requirement and visible black smoke or misfiring.

Four-stroke medium-speed engines

On trunk-piston medium-speed engines (MaK, Wartsila, Bergen), the intake valve provides the charge directly. The turbocharger charges a common intake manifold; each cylinder draws from that manifold through its intake valve at the start of the intake stroke. The timing of the inlet valve closing determines how much of the charged air is trapped. This is the lever that Miller timing exploits. Four-stroke installations frequently use pulse manifolds grouped in pairs or triplets of cylinders firing sequentially to preserve exhaust pulse energy.

Sequential turbocharging and cut-out for slow steaming

An engine designed for MCR at sea speed carries too much turbocharger capacity for the lower loads of slow steaming, which became common after 2008 as shipowners responded to high bunker prices and later to CII rating pressure under MARPOL Annex VI Regulation 28. At 50% engine load on an engine with two turbochargers both operating, each turbocharger receives approximately half the exhaust energy it was designed to handle. Both run at low speed, producing low pressure ratios and low efficiencies. The operating points on both compressor maps migrate toward the surge boundary.

Sequential turbocharging addresses this by shutting down one turbocharger at low engine loads. With one turbocharger cut out, the remaining unit receives all the exhaust energy from every cylinder. It spins faster, operates at a higher pressure ratio, and sits further from its surge line. Fuel consumption at the same power output decreases because the better-matched turbocharger delivers denser charge air more efficiently.

WinGD offers this as Automated Sequential Turbocharging (aSTC) on X82-2.0 and X92-1.1 engines with multi-turbocharger configurations. MAN’s ME series includes a similar function in its engine control system. The cut-out and cut-in events are automated, with hysteresis to prevent rapid cycling. During the transition, the engine control system adjusts fuel injection timing and quantity to avoid momentary air deficiency.

The cut-out turbocharger’s shaft continues to spin freely. If the oil supply to its bearings were interrupted, the bearings would fail. Lubrication is maintained even when the unit is not producing power.

Miller timing, two-stage turbocharging, and NOx Tier III

NOx forms primarily through the Zeldovich mechanism at the high temperatures and oxygen concentrations present in the combustion chamber during and immediately after ignition. Reducing peak combustion temperature is the most direct in-cylinder NOx control method.

Miller timing closes the intake valve earlier than the conventional bottom dead centre (BDC) timing. On a four-stroke engine, this means closing at 40 to 60 degrees before BDC; on a two-stroke engine with an intake valve, an equivalent retardation of the valve-close event achieves the same effect. After the valve closes, the air trapped in the cylinder expands slightly before the piston reverses into the compression stroke. This expansion lowers the charge temperature by 20 to 40 degrees Celsius at the start of compression, which reduces peak combustion temperature by a broadly proportional amount.

Research published in ASME Journal of Engineering for Gas Turbines and Power confirmed that combining Miller timing with two-stage turbocharging and intercooling achieves pressure ratios above 8 and NOx reductions of 50% or more compared to a conventional single-stage Tier I installation. MAN Diesel reported NOx reductions of over 30% and a simultaneous fuel consumption reduction of up to 8% in trials using intensive Miller timing with two-stage turbocharging at pressure ratios of 6.5 to 7.

In a two-stage arrangement, a low-pressure (LP) compressor stage takes ambient air and raises it to an intermediate pressure. An intercooler removes the compression heat. A high-pressure (HP) compressor stage then raises the already-compressed air to the final boost pressure. Each stage is matched to a corresponding turbine section. The intercooler between stages is the thermal key: without it, the two-stage compression would produce delivery temperatures too high for the charge-air cooler to handle within the available cooling water temperature.

The pressure ratio split between stages is typically 2.5 to 3 per stage to achieve an overall ratio of 7 to 9. At those pressure ratios, even with good inter- and after-cooling, the air receiver temperature is maintained below 50 degrees Celsius on a well-maintained installation.

Variable turbine geometry and waste-gate control

A fixed-geometry turbocharger is matched for one design point, usually the engine’s MCR on a specific fuel and ambient condition. Off-design operation changes both the exhaust energy available and the air demand, and the turbocharger responds by moving along its operating map.

Variable turbine geometry (VTG, also called variable turbine angle or VTA in some manufacturers’ literature) adjusts the nozzle ring vane angle while the engine is running. Closing the vanes reduces the effective turbine inlet area. The exhaust gas accelerates through the smaller passage and strikes the turbine blades at higher velocity, extracting more work from a given mass flow rate. This allows the turbocharger to spin faster at low engine loads than it would with fixed geometry, delivering higher boost pressure and better gas exchange. Opening the vanes at high load increases the inlet area to match the larger exhaust flow without over-speeding the rotor.

On a two-stroke engine, VTG is used to maintain adequate scavenge air pressure at loads below the normal cut-in point for auxiliary blowers, reducing reliance on electric blowers during slow steaming. When integrated with two-stage turbocharging and extreme Miller timing, VTG in the HP stage allows the NOx control to be modulated without changing valve timing, useful when the engine is outside an ECA.

A waste-gate is a simpler, coarser control: a bypass valve around the turbine that diverts exhaust gas when the turbine would otherwise over-speed the compressor. Waste-gates are common on four-stroke medium-speed engines and on smaller two-stroke installations where variable geometry would not be cost-effective. They are a form of energy disposal: the bypassed exhaust enthalpy is wasted. Variable geometry is preferred wherever the capital cost is justified.

Exhaust gas bypass and power turbines

On highly rated slow-speed engines, the turbocharger may produce more than enough compressed air at full load, leaving excess exhaust energy after the turbocharger turbine. Two strategies recover that energy.

An exhaust gas bypass valve opens at high load to route a fraction of the exhaust around the turbine. This prevents the turbocharger from over-speeding and reduces the compressor outlet pressure toward the target. The bypassed energy is wasted as heat in the exhaust gas economiser uptake. This is the conventional solution for managing excess turbine energy on current engine designs.

A power turbine (also called a turbo-compound turbine) is an additional turbine stage downstream of the main turbocharger. It extracts the remaining pressure energy from the exhaust after the main turbine has done its work. The power turbine output can be connected mechanically to the engine shaft via a gearbox and overrunning clutch, or it can drive a generator. Mitsubishi Heavy Industries published development work on a super waste-heat recovery system combining a power turbine with a steam turbine bottoming cycle; MHI-MME has fitted power turbines on selected large two-stroke installations to recover energy that would otherwise leave via the exhaust. The turbo-compound arrangement adds mechanical complexity, but a well-integrated system can reduce fuel consumption by 3 to 6% at full load, directly improving the attained CII rating under MARPOL Annex VI Regulation 28.

Turbocharger manufacturers and product families

Four principal manufacturers supply the marine market. Their product ranges are described below based on manufacturer-published technical material.

Accelleron (formerly ABB Turbocharging) separated from ABB in 2022 as a standalone company headquartered in Baden, Switzerland. The A100-L and A200-L series for slow-speed two-stroke engines cover engine outputs from 3,000 kW to 28,000 kW per turbocharger and had surpassed 10,000 orders as of Accelleron’s own announcement. The A100-H is the single-stage product for medium-speed four-stroke applications. The legacy TPL series remains in service on a large portion of the world fleet and receives continued support through Accelleron’s service network.

MAN Energy Solutions produces turbochargers in-house alongside its two-stroke and four-stroke engine lines. The TCA series (axial turbine, slow-speed) covers 3 MW to 30 MW per unit. The TCR (radial turbine, medium-speed) and TCT series address the four-stroke market. MAN states that its turbochargers are in service on more than 65,000 two-stroke propulsion and four-stroke auxiliary engines. The TCT was introduced in 2019 specifically for Tier III applications on MAN four-stroke engines.

MHI Marine Machinery & Equipment (MHI-MME) supplies the MET series. The MET-MA was the original axial-turbine line for large two-stroke engines; the MET-MB succeeded it, with the first MET71MB delivered in 2011 and the MET37MB frame size added in 2017. The MET-MBII generation, launched in 2018, is 16% greater in air flow capacity than the MET-MB at smaller overall dimensions. The MET-ER is the radial-turbine product for four-stroke medium-speed engines. MHI-MME reached 30,000 cumulative MET deliveries in 2017. In 2022, MHI-MME concluded a licensing agreement with Mitsui E&S Machinery to produce and sell MET turbochargers, expanding production capacity.

Kompressorenbau Bannewitz (KBB) is a German manufacturer producing turbochargers for medium-speed diesel and gas engines on ships and locomotives since 1953. Its product range covers smaller frame sizes than the three dominant marine suppliers, supplying primarily the medium-speed segment on smaller vessels and generator sets.

Turbocharger fouling and cleaning

Deposits accumulate on both the turbine and compressor sides of a turbocharger during service, more rapidly on residual fuel than on distillate. Fouling reduces efficiency and, if left unchecked, can unbalance the rotor or cause blade erosion when deposits spall.

Compressor washing

Compressor deposits are primarily oil and dust. Oil enters from leaking labyrinth seals on the turbine bearing housing and from crankcase mist carried in the scavenge air. Dust and salt enter with the intake air. Accelleron’s guidance, published in its Charge Magazine, specifies that compressor wet washing is applicable to both two-stroke and four-stroke engine turbochargers and is performed with the engine running at reduced load. A fine water spray is injected at the compressor inlet; the water evaporates partially and mechanically removes soft deposits from the impeller and diffuser. The engine load during washing is typically 30 to 50% MCR to keep the compressor outlet temperature at a level where the water injection doesn’t cause thermal shock.

Turbine cleaning

Turbine deposits are combustion products: vanadium pentoxide and sodium sulfate from residual fuel combustion, soot, and partially burned carbon. Accelleron distinguishes two methods by engine type. On four-stroke engines, turbine wet washing injects water into the turbine inlet when exhaust gas temperature before the turbine has been reduced to below 430 degrees Celsius by load reduction. On two-stroke engines, the turbine cleaning procedure uses dry granules (typically walnut shell or corn cob media) injected via a compressed air system into the turbine inlet. The granules physically abrade deposits from nozzle ring surfaces and turbine blades. Accelleron’s guidance specifies dry cleaning after every 25 to 50 operating hours depending on fuel quality and application.

Deferring cleaning leads to visible deterioration in turbocharger performance trends: rising exhaust gas temperature before turbine, falling scavenge air pressure at constant load, and increasing fuel consumption at the same power output. Marine engine performance monitoring covers the trending methods used to catch this degradation before it becomes irreversible.

Bearing arrangements and lubrication

The turbocharger rotor spins at high speed under the full radial load of the rotor weight plus unbalanced aerodynamic forces and the axial load from the pressure differential across the turbine and compressor. Two types of bearing are in use.

Plain journal bearings are the standard on all large marine turbochargers. They are hydrodynamic bearings: a continuous film of oil separates the journal from the bearing shell. The oil is supplied from the engine lubricating oil system at 3 to 5 bar, filtered through a fine strainer specific to the turbocharger. Drain return is to the engine sump. The oil supply must be maintained whenever the rotor is spinning; cutting the oil supply while the engine is decelerating causes bearing failure from the combined effect of residual heat soakback and inadequate lubrication. The marine lubricating oil systems article describes the system design requirements.

Rolling-element bearings (ball or angular-contact) are used on some smaller and medium-speed turbochargers. They carry higher radial and axial loads relative to their size and have lower friction, which improves transient response. Their service life is typically lower than plain bearings at equivalent operating hours, and they require a dedicated filtered oil supply at the correct pressure. Rolling-element bearings are gaining ground on medium-speed applications where fast transient response during power changes is valued.

Bearing temperature and lube oil inlet and outlet temperature are monitored through the engine room automation and monitoring system. Temperature trending is the primary indicator of bearing condition between overhauls.

Classification society requirements (IACS UR M73)

IACS Unified Requirement M73 governs the type testing, design approval, and certification of marine turbochargers. The requirement replaced UR M23 and addresses three areas:

Design approval requires the manufacturer to submit documentation of the rotor burst containment capability, the bearing design at maximum speed, and the predicted performance map. Class surveyors at the manufacturer’s works witness the prototype type test, which includes operation at 110% of the rated speed for a defined duration and a rotor burst test to confirm the casing contains blade fragments.

Matching to the engine requires the turbocharger manufacturer and engine builder to demonstrate that the compressor operating line at all engine loads stays within the certified performance map and above the minimum surge margin. This is verified at the engine test bed during type approval of the engine model.

Periodical survey in service follows the class society’s continuous machinery survey (CSM) programme. ClassNK’s Guidance on Continuous Machinery Survey (CMS), revised to Version 4 in June 2025, requires that the interval between consecutive examinations of each machinery item not exceed five years. Under the CSM programme, approximately one-fifth of the surveyed items are examined each year so that the entire inventory cycles through in the five-year period. For turbochargers, survey consists of internal inspection against manufacturer’s specifications, measurement of rotor clearances, bearing condition assessment, and performance comparison against the engine’s load cycle records. Class surveyors accept on-board trending data from the engine monitoring system as supporting evidence of condition.

Some class societies permit the chief engineer to conduct the survey of certain items in the machinery space under defined conditions, with a confirmatory survey by a class surveyor at the next port call. This is the “owner survey” provision that reduces the need for dry-dock surveys of running machinery.

IACS UR M72, revised to Revision 3 in April 2023, addresses the reciprocating internal combustion engine type test requirements that govern how the engine manufacturer must demonstrate the turbocharger-to-engine matching.

Interactions with MARPOL Annex VI emissions compliance

The turbocharger sits at the intersection of the NOx and energy-efficiency regulatory threads in MARPOL Annex VI.

On the NOx side, Regulation 13 requires engines built from 2016 operating in the North American and US Caribbean ECAs to comply with Tier III limits: for an engine with a rated speed of 130 rpm or below (typical slow-speed two-stroke), the Tier III limit is 3.4 g/kWh. The in-cylinder path to this limit requires the turbocharger to deliver charge-air pressure ratios that support extreme Miller timing. Engine builders MAN ES and WinGD both offer two-stage turbocharging options on new-build slow-speed engines for ECA operation.

On the energy-efficiency side, EEXI (Energy Efficiency Existing Ship Index) under Regulation 23 and CII (Carbon Intensity Indicator) under Regulation 28 both penalise fuel consumption. Turbocharger fouling, sequential cut-out calibration, and slow-steaming optimisation are among the practical measures ship operators and engine manufacturers address to improve the attained CII rating. A turbocharger running with 2% higher specific fuel oil consumption (SFOC) than the clean reference curve carries a quantifiable CII penalty that compounds across the year.

The engine performance monitoring workflow described in marine engine performance monitoring provides the data inputs needed to separate turbocharger-related SFOC drift from fuel quality variation and other engine condition factors.

See also

Related wiki articles

• Marine Diesel Engine
• Uniflow Scavenging in Two-Stroke Marine Engines
• Scavenge Port Geometry and Timing
• Marine Engine Performance Monitoring
• Marine Engine Combustion Analysis
• Marine Engine Fuel Injection Systems
• Marine Engine Common Rail Technology
• Marine Engine Camshaft and Valve Train
• Marine Engine Cylinder Liners and Pistons
• Marine Engine Crankshaft and Main Bearings
• Marine Sea Water Cooling Systems
• Marine Lubricating Oil Systems
• Marine Auxiliary Engines and Generators
• Marine Engine Room Automation and Monitoring
• Marine Engine Room Ventilation and Uptakes
• Continuous Survey of Hull and Machinery
• Cross-Scavenging in Legacy Two-Stroke Designs
• Loop Scavenging Versus Uniflow Scavenging

Related calculators

• Turbocharger Surge Margin
• Charge-Air Cooler Effectiveness
• Turbocharger: Axial vs Radial Exhaust Selection
• Engine Scavenge Pressure
• System: Main Engine Slow-Speed 2-Stroke
• System: Main Engine CAC Plate-Fin Cooler
• System: Auxiliary Engine Medium-Speed 4-Stroke
• Engine Pcomp vs Pmax Ratio
• Auxiliary Engine Load Factor

Limitations

This article describes turbocharging as applied to conventional marine diesel engines burning residual or distillate fuel. The turbocharging arrangements on dual-fuel engines (LNG, methanol, ammonia) differ in detail: fuel gas injection can occur upstream of the compressor in some low-pressure two-stroke designs, requiring different surge management and materials specifications. Ammonia combustion in particular raises questions about NOx formation chemistry that a standard Miller-timing analysis doesn’t fully address.

Two-stage turbocharging performance data cited here comes from published engine-builder and academic sources. Actual performance on a specific installation depends on ambient conditions, fuel specification, engine load profile, charge-air cooler condition, and exhaust back pressure from the gas economiser and exhaust ducting. The surge margin formula gives a single-point snapshot; real surge boundary location shifts with compressor fouling, wear, and ambient humidity.

IACS UR M73 requirements are correct as of 2025. Class societies issue supplements and national interpretations; the applicable version of the rule for a specific engine and flag depends on the engine’s build date and the class’s implementation schedule for UR revisions. Always consult the flag-state-accepted class rules for a specific vessel.