Scavenge Port Geometry and Timing in Two-Stroke Engines

Scavenge port geometry is the fixed-in-metal counterpart to the hydraulically adjustable exhaust valve: once the liner is machined, port height, width, angle, and edge profile are set for the life of the liner. Port height alone determines when scavenging starts and ends, making every port dimension simultaneously a flow parameter and a timing parameter. This article covers the geometry of the port window cut into the liner bore, the crank-angle relationships that translate height into timing, the time-area integral that governs charge delivery, edge treatment and ring-snagging limits, and the long-service issues of wear, deposit fouling, and recutting.

For the fluid-mechanics of the scavenging process, see Uniflow Scavenging in Two-Stroke Marine Engines. For the comparison between uniflow and cross-flow approaches, see Loop Scavenging Versus Uniflow Scavenging. For the liner material and manufacturing context in which ports are machined, see Cylinder Liner Design for Two-Stroke Engines.

Port geometry fundamentals

The port window

A scavenge port is a rectangular or slightly trapezoidal window cut through the liner wall. Each port has four defining dimensions: axial height $h_p$, circumferential arc width $w_p$, the radial wall thickness it penetrates (the liner wall thickness, typically 50 to 90 mm), and the tangential inclination angle $\alpha$ at which the port axis deviates from the cylinder radius. The product $h_p \times w_p$ is the geometric face area of one port. Multiplied by port count $N$ and the discharge coefficient $C_d$, it gives the maximum effective open area.

Modern slow-speed engines distribute the total port area across 14 to 24 individual ports per cylinder. MAN B&W G70ME-C10.5 engines, with a 700 mm bore, carry 20 ports. WinGD X72 engines, also 720 mm bore, use 18 ports. The difference reflects both casting practice and slightly different swirl targets. Distributing area across more ports narrows each port, which reduces the circumferential gap each ring must bridge during the port-belt stroke and raises the bridge-to-port width ratio, improving liner stiffness.

Port height and its structural constraint

Port height is the most consequential single dimension. It sets port-open duration in crank degrees (see section on timing below), it determines how much of the liner bore circumference is cut away in the axial direction, and it fixes the running length the piston crown spends passing over the open ports on each stroke.

Typical values are 120 to 220 mm for contemporary slow-speed engines. A 160 mm port on a 2,000 mm stroke (MAN B&W G50ME-C class) spans 8.0 percent of the stroke. The same 160 mm port on a 3,800 mm stroke (G95ME-C class) spans only 4.2 percent, so the same absolute height produces a materially shorter angular window on a longer-stroke engine. This is why larger-bore engines often carry taller ports in absolute terms: the G95ME-C uses ports of approximately 200 to 220 mm to keep the angular window comparable to smaller bore engines.

The structural constraint on port height comes from the bridge between adjacent ports. Each bridge must carry the hoop stress from combustion pressure, the thermal gradient stress from the hot gas side to the cooled outside, and the bending load imposed as the piston ring rides over the port opening. The Lame equation for hoop stress in the bridge region gives a minimum bridge thickness, which the designer then verifies with finite-element analysis. MAN Energy Solutions’ design guidelines specify a minimum bridge width of 25 mm for 700 mm bore liners, which limits total circumferential cutaway to roughly 65 percent of the liner bore circumference.

Tangential port angle and swirl generation

Each port is machined at an angle to the radial direction so that the incoming air jet has a tangential velocity component. This generates organized rotation (swirl) inside the cylinder, directed from BDC toward TDC. Swirl serves two functions: it improves fuel-air mixing during injection and combustion, and it carries residual exhaust gas toward the exhaust valve at TDC where it can be pushed out.

The tangential angle $\alpha$ is measured between the port centerline and the local cylinder radius. At $\alpha = 0$, the port fires radially inward and produces no swirl. At $\alpha = 20$ degrees, the tangential velocity component is $V_{jet} \times \sin(20°) \approx 0.342 \, V_{jet}$, and the axial component is $V_{jet} \times \cos(20°) \approx 0.940 \, V_{jet}$. Most design points sit between 17 and 23 degrees.

WinGD’s X-series engineering documentation specifies 20 degrees as the nominal port angle for the X52, X62, X72, and X82 bore families. MAN Energy Solutions uses the same 20-degree nominal for the G and S series, with the note that some engines in the earlier B&W MC series used 17 degrees. CIMAC Working Group 17’s 2016 guideline on two-stroke scavenging cites 15 to 25 degrees as the accepted range across all major OEMs.

Steeper angles (higher $\alpha$) produce a higher swirl number at TDC but also reduce the axial component of the inflowing jet, which can impair blowdown of the combustion gases in the early scavenge phase. At 25 degrees the axial component drops to 90.6 percent of the jet velocity, a reduction that is measurable in CFD studies but modest in absolute terms. The practical limit on angle comes from port-to-port interaction: at angles above about 28 degrees the tangential jets from adjacent ports begin to collide in the cylinder center, producing turbulence rather than organized swirl.

Port width and circumferential coverage

Port width is the arc length of each port window. Total circumferential coverage, the sum of all port widths, is typically 55 to 68 percent of the liner bore circumference on modern engines. The remainder is bridges. Wide individual ports (say, 30 mm each on a 700 mm bore liner) require fewer bridges and simplify casting but force each piston ring to bridge a larger unsupported gap on each stroke, increasing the risk of ring fracture.

The ring-bridging criterion sets a hard upper limit. Classification society rules (for example Lloyd’s Register Rules for the Manufacture, Testing and Certification of Materials, Table 4.3.2) specify that the maximum unsupported port width a piston ring may bridge is a function of ring thickness and ring material. For typical slow-speed top rings of 15 to 20 mm thickness in SG iron or steel, the limit is approximately 50 to 60 mm individual port width. This pushes designs toward narrower ports and more of them.

Port bridge geometry and stress

The bridge between two adjacent ports is a thin, axially oriented column of cast iron. In service, each bridge experiences:

1. Peak combustion pressure of 180 to 220 bar acting on the bore face.
2. A thermal gradient of approximately 150 to 200 K across the 50 to 90 mm wall thickness, creating tensile stress on the cooled outer surface.
3. Piston-ring contact stress as each ring traverses the port belt twice per cycle.

MAN B&W service experience data (published in MAN Energy Solutions’ periodical engine bulletin series) indicates that bridge cracking initiates most often at the axial ends of the port opening, where the stress concentration from the rectangular corner is highest. This is why port corners are radiused rather than left sharp: a 5 mm fillet radius at the corner of a 160 mm × 25 mm port reduces the theoretical stress concentration factor from approximately 3.0 to 1.6, based on standard Kt curves for rectangular notches.

Bridges also wear on the running surface because piston rings pass over them. Bridge running-surface wear rates are typically 0.01 to 0.05 mm per 1,000 hours in normal service. After 20,000 to 30,000 hours, the bridge surface has worn below the surrounding liner bore by 0.3 to 1.0 mm, creating a minor step that affects ring seating.

Port timing and crank-angle relationships

Symmetry about BDC

Port timing is controlled solely by the piston position. The piston’s upper edge uncovers the top of the port window on the downstroke and re-covers it on the upstroke. Because the slider-crank geometry of the piston mechanism is, to first order, symmetric about BDC, the crank angle at which the port opens (PO) before BDC equals the crank angle at which it closes (PC) after BDC. The ports are inherently symmetrically timed.

This is qualitatively different from the exhaust valve, which is actuated by a hydraulic servo and can open and close at independently chosen crank angles. The port designer cannot shift the opening event without equally shifting the closing event.

From port height to crank-angle duration

For a slider-crank with stroke $S$ and connecting-rod length $L$, the piston displacement from TDC as a function of crank angle $\theta$ (measured from TDC) is:

The port opens when the piston top edge reaches the top of the port window; it closes when the piston top edge re-covers the top of the port window on the upstroke. The piston position at which the port is exactly fully open is:

where $h_{BL}$ is the distance from the piston top ring to the piston crown (the “top land” height, typically 80 to 150 mm on slow-speed engines). Setting $x(\theta_{PO}) = x_{PO}$ and solving for $\theta_{PO}$ gives the port-open crank angle. For a connecting-rod ratio $L/S = 2.2$ (typical for modern long-stroke engines), and port height of 160 mm on a 3,000 mm stroke engine, the port-open angle works out to approximately 47 degrees BBDC. Port-close is then 47 degrees ABDC.

Higher $L/S$ ratios reduce the second-order term and make the timing more nearly sinusoidal. Lower ratios, found in earlier short-stroke designs, shift the piston more quickly through BDC and slightly widen the angular window for a given port height.

Typical port-open and port-close angles

Published MAN B&W performance data for the G-series ME-C engines lists port timing in the range of 40 to 55 degrees BBDC / ABDC depending on specific engine variant and rating. The G70ME-C10.5, for example, carries a port-open angle of 46 degrees BBDC at its nominal rating. WinGD X82 engines operate with port opening at approximately 50 degrees BBDC in standard configuration.

Earlier MC-series engines from MAN B&W (e.g., S70MC-C7) used narrower ports and correspondingly earlier opening, around 40 to 43 degrees BBDC. The shift to deeper ports on modern ME-series engines reflects the need to maintain adequate scavenge windows at the higher maximum continuous rating (MCR) power densities of current designs, where the time available per cycle (at higher rpm) is shorter.

The time-area integral

The port-open duration in crank degrees tells only half the story. What matters for charge delivery is the cumulative product of open area and time, the port time-area integral $I_{TA}$:

where $A_{eff}(t)$ is the instantaneous effective open area (geometric area × $C_d$) and the integral runs from port-open time to port-close time. In crank-angle form, dividing by engine angular velocity $\omega = 2\pi n / 60$:

The unit of $I_{TA}$ is m² · s, and it is proportional to the total charge mass that can enter the cylinder at a given pressure differential. CIMAC WG17’s guideline tables for two-stroke engines list target time-area values of $8 \times 10^{-4}$ to $1.4 \times 10^{-3}$ m² · s per litre of cylinder displacement for engines in the 40 to 100 cm bore range. Engines with time-area below this range tend to show high scavenge pressure drops and elevated exhaust temperatures at full load.

The time-area approach was the principal design tool before CFD became practical. Blair’s “Design of Two-Stroke Engines” (SAE International, 1996) gives the complete derivation; Woodyard’s “Pounder’s Marine Diesel Engines” (9th edition, 2009) applies it to large slow-speed engines specifically. For the current wave of me-series engines, CFD has supplemented rather than replaced the time-area calculation because CFD is costly and the time-area still serves as a quick screening tool for alternative port configurations.

Port effective area at crank angle $\theta$

At any crank angle during the scavenge event, the portion of the port window exposed below the piston top edge is:

and zero outside the port-open window. Here $x_{port,top}$ is the piston displacement at which the top of the port is just uncovered. The effective area across all ports is:

Peak effective area occurs at BDC when the full port height is exposed and $C_d$ is at its maximum (the flow velocity through the port is lowest at BDC, so the contraction coefficient is highest). A representative peak effective area for a G70ME-C with 20 ports of 25 mm width and 160 mm height, $C_d = 0.72$, is approximately 0.058 m².

Discharge coefficient and port edge profile

Why $C_d$ is less than unity

The geometric face area of the port overstates the actual flow capacity because the flow contracts as it enters the sharp-edged port opening. The vena contracta is narrower than the opening; beyond it the flow re-expands, dissipating kinetic energy. The ratio of actual mass flow to the theoretical mass flow through the geometric area at the measured pressure differential is the discharge coefficient $C_d$.

For marine scavenge ports at design conditions (scavenge pressure 2.5 to 3.5 bar absolute, pressure ratio across the port 1.2 to 1.5), $C_d$ is typically 0.65 to 0.78. The lower end applies to sharp-edged ports; the upper end to well-profiled edges. MAN Energy Solutions’ liner engineering documentation quotes 0.72 as the standard design value for profiled ports on current ME-C series engines.

Edge chamfering and the ring-snag criterion

Port edges on the bore side (the side the piston ring contacts) are chamfered to a radius of 1 to 3 mm. This serves two simultaneous purposes. First, it eliminates the sharp step that a piston ring would otherwise have to snap over, which can fracture the ring or score the liner. Second, it reduces the entry loss coefficient.

On the scavenge-air side (outer face), chamfers of 3 to 6 mm are typical. Larger chamfers on the outer face produce a smoother entry geometry and improve $C_d$ by reducing the vena contracta effect, but they also reduce the effective wall thickness at the port edge, which must be checked against the structural criterion.

The ring-snag constraint is the tighter of the two. A piston ring passing over a port edge can catch if the edge protrudes inward (proud of the bore) by more than approximately 0.05 mm, or if the chamfer radius is smaller than the ring’s own edge radius. In service, port edges wear toward the bore as liner wear progresses. At some point, a previously acceptable edge profile becomes undercutting, and the risk inverts: the edge is now recessed below the bore and no longer a snag risk, but the recess traps oil, affecting lube distribution.

Deposit fouling and its effect on $C_d$

Hard carbon and sulfated ash deposits accumulate on port edges over service intervals. Heavy fuel oil (HFO) combustion produces more deposit-forming compounds than distillate fuels, but LSFO and VLSFO fuels, while sulfur-reduced, can still deposit wax and lighter ash at lower exhaust temperatures in slow steaming. MAN Energy Solutions’ maintenance guidelines for ME-C engines specify port cleaning at every piston overhaul, typically every 12,000 to 18,000 running hours, or earlier if scavenging space inspection shows thick deposits.

A 2 mm deposit layer on port edges reduces the free area approximately in proportion to the ratio of the deposited layer to the port height. On a 160 mm port, 2 mm per side reduces free height by 2.5 percent. The $C_d$ reduction from edge blunting (the deposit rounds the entry, actually improving $C_d$ slightly) partially offsets the area reduction, but in practice a fouled port delivers 2 to 4 percent less charge mass than a clean port, which shifts exhaust temperature upward by a similar proportion.

Interaction with exhaust valve timing

The asymmetric overall gas-exchange sequence

Although the scavenge ports open and close symmetrically about BDC, the overall gas-exchange event is not symmetric. The exhaust valve opens long before the ports, typically 80 to 110 degrees BBDC, to allow blowdown of the cylinder. The sequence on a MAN B&W ME-C engine at full load is:

1. Exhaust valve opens: approximately 95 degrees BBDC.
2. Blowdown: 95 BBDC to approximately 40 BBDC, cylinder pressure falls from peak combustion pressure to scavenge receiver pressure.
3. Scavenge ports open: 46 degrees BBDC (for G70ME-C10.5).
4. Scavenging: 46 BBDC through BDC to 46 ABDC, fresh charge enters, residual gas exits via the exhaust valve.
5. Scavenge ports close: 46 degrees ABDC.
6. Exhaust valve closes: approximately 15 to 35 degrees ABDC (load-dependent on ME-C).
7. Compression begins.

Steps 5 and 6 overlap: the ports close at 46 ABDC but the exhaust valve stays open until 15 to 35 ABDC (these are on the same positive side of BDC, so the exhaust valve closes well before port-close in this configuration). Wait: re-examining the sequence. On ME-C engines the exhaust valve closure angle (EVC) is typically 15 to 35 degrees ABDC, meaning EVC occurs before port-close (PC at 46 ABDC). So the effective scavenge window runs from port-open until exhaust valve closure, after which the cylinder is sealed on both sides and compression begins with fresh charge that can no longer escape. Port-close at 46 ABDC is later than EVC, so there is a period (EVC to PC) when the cylinder is sealed on the exhaust side but the ports are still open on the scavenge side, allowing additional fresh charge to enter at scavenge pressure and compress slightly against the closed exhaust valve. This late-intake effect slightly increases trapped charge mass.

Load-dependent EVC and its effect on trapping efficiency

ME-C and ME-GI engine series use electronically controlled hydraulic actuation (FIVA valves) to vary EVC with load and speed. At low load, advancing EVC (closing earlier, shorter after BDC) traps more compressed charge at lower in-cylinder pressure, improving part-load combustion stability. At high load, retarding EVC allows more exhaust gas to leave before trapping begins.

The port timing, being fixed in the liner, cannot be adjusted. This means the relative relationship between the fixed PC angle and the variable EVC shifts with every operating point. At 25 percent MCR, EVC may be as early as 5 degrees ABDC; at 100 percent MCR, it may be 30 degrees ABDC. In both cases, PC remains at 46 ABDC. The scavenge efficiency and trapping efficiency both vary as a consequence, and the engine management system accounts for this in its air-fuel ratio control loop.

Blowdown adequacy and the EVO-to-PO interval

The interval between exhaust valve opening (EVO) and scavenge port opening (PO) must be long enough for cylinder pressure to blow down from combustion residual pressure (typically 8 to 15 bar at EVO on a modern engine) to scavenge receiver pressure (2.5 to 3.5 bar) before the ports open. If the ports open before blowdown is complete, high-pressure combustion gases flow backward into the scavenge receiver, contaminating the fresh charge and potentially damaging the scavenge receiver and turbocharger.

The blow-down factor $B_{d}$, defined as the ratio of cylinder pressure at PO to scavenge receiver pressure, should be as close to 1.0 as possible. Values above 1.05 indicate late port-opening relative to EVO, which can cause contamination of the scavenge space. MAN B&W design practice (as documented in the ME-C Technical Overview) calls for a minimum EVO-to-PO interval of 45 to 55 crank degrees to ensure adequate blowdown across the load range.

This interval is set by two fixed geometries (the exhaust valve cam profile and the liner port height) and cannot be changed in service without machining the liner or reprogramming the exhaust valve actuator.

Port area optimisation and competing constraints

The four-way trade-off

Port geometry must satisfy four mutually constraining requirements simultaneously. More port area means better breathing but weaker liners. Earlier port opening (taller ports) means more scavenging time but earlier trapping of residual gas. Steeper port angles mean stronger swirl but weaker axial flow and higher risk of short-circuiting. More ports per circumference means smaller individual ring-bridging gaps but more bridge stress concentrations.

The design solution is not a single optimum but a feasible region within which the designer can trade off one performance metric against another. MAN Energy Solutions and WinGD have converged on similar solutions (16 to 20 ports, 120 to 220 mm height, 17 to 23 degree angle) because the same physics constrains both. Small differences in the details reflect different scavenge-box geometry, different exhaust valve actuation characteristics, and different target swirl numbers.

Port height versus cylinder bore

Absolute port height scales with bore but not linearly. The structural constraint (bridge stress) scales with bore because combustion pressure force scales as bore squared while bridge cross-section scales as bore. The result is that port height as a fraction of stroke increases with bore. A 500 mm bore ME-C engine carries ports of approximately 120 to 140 mm; a 900 mm bore ME-C carries ports of 200 to 230 mm. The ratio of port height to stroke stays roughly in the range 6 to 9 percent across the family.

Port count versus port width

Holding total port area constant, the designer can use fewer wide ports or more narrow ports. The ring-bridging limit caps individual port width at around 50 to 60 mm for standard ring materials. At a 700 mm bore, with 66 percent circumferential coverage and 25 mm bridge widths, the maximum port count is approximately 20 for 25 mm wide ports. The minimum port count consistent with adequate total area and a 50 mm width limit is approximately 12. The 16-to-20 range seen in practice occupies the central part of this design space.

Discharge coefficient and port edge design

The effective port area is the geometric area scaled by $C_d$. Improving $C_d$ from 0.68 to 0.74 is equivalent to adding 8.8 percent more geometric area without any structural change to the liner. Edge geometry is therefore a significant design lever. MAN Energy Solutions’ liner development work (referenced in ME-C technology descriptions) credits improved edge profiling as a contributor to the increase in specific output across MC to ME-C generations.

The limits on edge improvement are:

• Minimum bore-side edge thickness consistent with the ring-snag criterion (chamfer must not leave a step that snags).
• Minimum air-side edge thickness consistent with bridge stress (large chamfers thin the bridge at the port corner).
• Manufacturing feasibility of consistent edge geometry across all ports on all liners.

In-service wear, recutting, and reconditioning

Liner bore wear and port geometry change

As the liner bore wears, the bore diameter increases uniformly. This changes port geometry because the absolute port dimensions are fixed in the liner wall while the running surface radius increases. The bore-side edge of each port effectively moves inward relative to the new bore surface as the bore increases. After 40,000 to 50,000 hours without liner replacement, typical bore wear of 1.0 to 1.5 mm per side on a 700 mm bore engine shifts each port edge inward by the same amount.

This bore-side shift is small relative to port width (a 1.5 mm edge recess on a 25 mm wide port is a 6 percent change) and its main practical effect is on $C_d$: the recessed edge slightly reduces the contraction effect, marginally improving $C_d$, but the worn surface also has higher roughness, which partially offsets the gain.

Port edge erosion from gas flow

Scavenge air enters the port at velocities of 60 to 120 m/s at design conditions. Gas velocities during blowdown (when the exhaust valve first opens) are higher, up to 200 to 300 m/s, but the flow direction at that point is outward through the exhaust valve, not through the ports. The inward scavenge flow through the ports at 60 to 120 m/s does not by itself erode the port edges significantly in normal service. Erosion becomes a concern when:

• The charge air contains abrasive particulate from a poorly maintained air filter.
• Back-pressure conditions cause combustion gases to flow through the ports in reverse during blowdown, exposing the bore-side edge to hot gas erosion.
• Localized temperature non-uniformity causes one side of the port belt to run hotter than average, softening the liner material at the affected port edges.

MAN Energy Solutions’ service bulletins document cases of port-edge erosion after air cooler fouling, where the elevated charge air temperature reduced the cooling effect on the liner port belt. The corrective action in those cases was cooler cleaning plus port-edge re-chamfering.

Port recutting

When port edges wear or erode beyond acceptable limits, the standard reconditioning procedure is re-chamfering. This is done in situ with a hand-held grinder and template, or with a port-chamfering machine that clamps to the liner bore and cuts a controlled radius. The target is to restore the edge profile to original drawing dimensions.

Re-chamfering removes material from the port edge, slightly widening each port by typically 0.5 to 1.5 mm per rework. After three to four re-chamfers, the cumulative port width increase can be 3 to 6 mm, which must be re-evaluated against the ring-bridging criterion. MAN Energy Solutions’ maintenance guidelines for ME-C engines specify that liner replacement should be considered when re-chamfering has increased individual port width beyond the limit of 60 mm (for standard ring configurations).

WinGD’s X-series maintenance manual takes a similar approach, with a port-width alert limit of 55 mm for the X52 bore and 60 mm for the X82 bore.

Deposit cleaning and chemical descaling

Port deposit cleaning at overhaul follows one of three approaches depending on deposit type and severity. Light deposits (thin ash films) are removed by wire brush and scraper. Moderate deposits are treated with chemical descalants, typically alkaline solutions that dissolve sulfate and carbonate deposits without attacking cast iron. Severe hard carbon deposits may require mechanical chipping followed by chemical treatment.

Chemical descaling requires care: if the descalant contacts the running surface above the port belt, it can affect the liner’s honed finish and the boundary layer for oil retention. Standard practice is to mask the bore above and below the port belt before applying the chemical.

Port wear monitoring

In-service port condition monitoring consists of two measurements taken at every piston overhaul:

1. Port width measurement (to detect progressive widening from ring-bridging wear or re-chamfering).
2. Bridge surface step measurement (to detect differential wear between the bridge running surface and the surrounding liner bore).

A step of more than 0.3 mm on the bridge running surface indicates that rings are dragging over the edge of the bridge and that ring wear is likely elevated. This measurement is taken with a depth micrometer bridged across two adjacent ports.

MAN B&W MC/ME maintenance records historically cite bridge step as one of the three leading indicators of liner distress, alongside bore wear and corrosion pitting. WinGD’s X-series maintenance documentation includes a liner condition worksheet that records port-width and bridge-step at each piston overhaul for trend analysis.

Computational fluid dynamics in port design

Role of CFD in current practice

CFD now plays the primary role in establishing port geometry for new engine variants. The simulation domain includes the scavenge receiver, all ports, the cylinder bore, and the exhaust valve. The mesh moves with the piston using an arbitrary Lagrangian-Eulerian formulation, and calculations are carried through a complete scavenge cycle (typically from EVO to EVC plus some compression period to check swirl persistence).

Key outputs from a port-design CFD run:

• Scavenging efficiency as a function of time (how much of the trapped mass at EVC is fresh charge).
• Swirl number at TDC, defined as the ratio of mean angular velocity to engine rotational speed.
• Trapping efficiency (fraction of delivered air that is retained at EVC).
• Discharge coefficient as a function of port opening fraction, extracted from the simulation for use in simpler models.
• Temperature distribution in the port belt region, to check for thermal hot spots near the bridges.

MAN Energy Solutions and WinGD both use proprietary CFD tools calibrated against engine test-cell measurements. WinGD published results of CFD-assisted port-angle optimization for the X-series in the 2018 edition of their engine technical description, reporting a 1.2 percent improvement in scavenging efficiency by shifting port angle from 18 to 20 degrees across the bore range.

Swirl number sensitivity to port angle

The swirl number at TDC, $Sw_{TDC}$, is related to the port angle $\alpha$ approximately as:

where $A_{bore} = \pi D^2 / 4$ is the bore area, and $k$ is a constant that depends on bore-to-stroke ratio, port height fraction, and cylinder cover geometry. For MAN B&W G-series engines, $k$ is in the range 0.15 to 0.25. At $\alpha = 20$ degrees and $A_{eff,peak}/A_{bore} = 0.15$, this gives $Sw_{TDC} \approx 0.022 \times k^{-1}$, which for $k = 0.20$ yields $Sw_{TDC} \approx 0.11$. Full CFD values reported by MAN and WinGD for production engines at design point are typically 0.8 to 1.4, indicating that the above simplified expression significantly underestimates the swirl because it ignores the rotational momentum accumulation from the distributed inflow across the full port-open interval. The simplified form is useful for relative sensitivity analysis but not absolute prediction.

The CFD result is that a 1-degree increase in port angle typically changes swirl number by 3 to 6 percent at constant port area, consistent with the $\sin\alpha$ dependence.

CFD validation against engine test data

Port-geometry CFD is validated against in-cylinder swirl measurement (by multi-port pitot traverses in deactivated cylinders), scavenging efficiency measurement (by tracer gas method, with argon or CO₂ injected at BDC and measured at the exhaust valve), and boundary condition validation from p-theta diagrams (cylinder pressure versus crank angle traces from indicator passages). The combination gives confidence in predicted scavenging efficiencies to within 1 to 2 percent and swirl numbers to within 5 to 10 percent.

Limitations

The scavenge port geometry analysis presented here applies to slow-speed crosshead two-stroke diesel engines in the 40 to 100 cm bore range, where ports are machined in the liner wall and controlled by piston motion. The following limitations apply:

Port timing formulas based on the slider-crank kinematics are accurate to within 0.5 crank degrees for engines with connecting-rod ratio $L/S > 2.0$. Shorter connecting-rod engines (L/S less than 1.8, unusual in modern practice) require more accurate numerical integration of the exact kinematic equations.

The time-area integral approach is a lumped-parameter model. It predicts charge delivery capacity but not spatial distribution, swirl number, short-circuit fraction, or temperature distribution. It remains useful as a screening tool but cannot replace CFD for detailed port design.

Discharge coefficient values quoted here (0.65 to 0.78) apply to ports in the fully open condition with pressure ratios of 1.2 to 1.5. At partial opening (port barely uncovered by the piston), $C_d$ is lower, in the range 0.4 to 0.6, because the approach flow is constrained by the narrow gap between the piston top edge and the port top edge. This effect is captured in numerical integration of the time-area integral but is lost in simplified calculations using a constant $C_d$.

Port wear rates quoted from MAN and WinGD documentation reflect engines operated within the fuel specification and load profile described in the original design basis. Engines running above design turbine-inlet temperature, using fuel outside specification, or with improperly maintained air coolers will show accelerated port-edge erosion and deposit rates.

Bridge cracking statistics cited from service bulletin data reflect the installed fleet composition at the time of publication. Engines with upgraded liner alloys (post-2015 production with higher Ni-Mo content, per MAN Energy Solutions’ circular on liner alloy revision) show lower bridge cracking incidence.

The interaction of port timing with exhaust valve timing described here reflects the ME-C series variable-timing capability. Older MC-series engines with mechanical exhaust-valve timing use a fixed EVC set by the cam profile, which means the EVC-to-PC relationship does not change with load. Analysis of those engines requires a different approach.

Related calculators

• Engine Scavenge Pressure Calculator for estimating scavenge receiver pressure as a function of load and turbocharger performance.
• Engine Mean Piston Speed Calculator for relating stroke, rpm, and piston velocity at the port belt.

See also

• Uniflow Scavenging in Two-Stroke Marine Engines
• Loop Scavenging Versus Uniflow Scavenging
• Cylinder Liner Design for Two-Stroke Engines
• Exhaust Valve Actuation in Two-Stroke Marine Engines
• Piston Ring Pack Design for Two-Stroke Engines
• Cylinder Lubrication Systems for Two-Stroke Engines