mc engines
DESCRIPTION
MC engines developmentTRANSCRIPT
Contents
The MC EnginesService Experience
Page
Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3
Cylinder Condition . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4
Design and Experience Interaction . . . . . . . . . . . . . . . . . . . . . . . . . 6
Piston. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6
Exhaust valve . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9
Valve spindle . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9
Bottom piece . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9
Valve housing. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9
Air spring . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9
Actuator . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10
Fuel oil system. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10
Fuel valves. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11
Bearings . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12
Main engine structure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12
Top bracing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13
Conclusion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13
Introduction
As the first L35MC entered service in1982 and the first 6L60MC in December1983, the MC generation is backed upby extensive experience accumulatedfrom the considerable number of run-ning hours logged by the very largenumber of plants in service.
Throughout the years, the engines havebeen updated and uprated on a regularbasis, when this has been justified byservice experience and made possibleby the use of new materials and moreadvanced calculation methods.
This situation also means that, from timeto time, it has been difficult to distinguishbetween various generations of the MCengines. To alleviate this, we introducedthe Mark designation several years agofor easy reference.
Fig. 1 shows the reference list as at98.10.01. It can be estimated that theaccumulated service experience gainedfrom the more than 4,000 engines inservice exceeds 75 million running hours.
In the following, we will describe recentdesign features introduced in order tocater for higher ratings. These features
will enhance reliability and availabilitydespite the higher ratings becausethey increase the design margins.
The MC engine programme of 1998-1999is larger than ever before, comprising25 engine types, each available in anumber of cylinder configurations, seeFig. 2.
The MC EnginesService Experience
115 25
98
912
196
1382
657
427
Number of engines
Engines in servicetotal 4,164
Engine type
Engines orderedtotal 5,043
0
100
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80 70 60 50 46 35 26
11031,100
1,000
167
42
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159
90
128
367
539
1171
872
158
776
152
Fig. 1: MC engines in service and ordered as at 1998.10.01
50 120Speed r/min
250200160140100807060
L42
L35
S26
L80
S35
K90
S80
S90-C
L60S80-C
S60-C
706050
40
30
20
15
6
4
1
10
8
100
PowerBHP kW
x 1000
60
80
10
8
2
5
3
40
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L50
S42
L70
K98
K80-CK90-C
S50
S60
L90-C
S70
S46-CS50-C
K98-C
2
S70-C
Fig. 2: The two-stroke MC engine programme 1998-1999
3
Cylinder Condition
Reliability and economy are the importantfactors for the ship operators, alsowhen talking about cylinder condition.
• Reliability is ensured by safe andsufficiently long overhaul intervals
• Economy depends on wear rates ofcylinder liners and piston rings and,even more, on the cylinder lube oilconsumption.
We therefore carefully monitor the rela-tionship between wear, lube oil dosagesand Time Between Overhauls (TBO) toobtain the optimum relationship oneach engine type.
Wear rates and overhaul intervals areclosely related to the cylinder lube oildosages actually used. The MAN B&Wguideline feed rates are chosen so asto be on the safe side, and are notnecessarily the optimal oil dosage, seenfrom a purely economical point of view.This dosage varies widely, dependingon the parts and manpower costs ofmaintenance, overhauling opportunitiesin relation to the ship’s schedule, thecurrent cylinder lube oil price, etc.
While purely economical considerationsof the cylinder lube oil cost versus thecost of cylinder liners may favour acertain degree of oil starvation, mostowners choose to locate their cylinder oilfeed rate at a comfortable 0.9 g/bhph.
Wear rates of piston rings and cylinderliner, together with cylinder lube oil con-sumption, are shown in Figs. 3, 4 and 5.
From our database, we have extracteddata regarding cylinder liner wear for 26,35 and 42-cm bore engines. The statis-tics include 455 liners, and the averagespecific wear is 0.037 mm/1000h, withan average lube oil feed rate of 0.90g/bhph. With our average wear limit of0.6% of the liner diameter, this corre-sponds to a liner lifetime of 42-60,000hours.
The normal range of top piston ringwear centres around 0.26 mm/1000h
which, with a wear limit for the ring of2.5 to 3 mm, permits intervals for pis-ton overhauls of above 10,000 hours.
It will be noted that wear rates of cylin-der liners and piston rings for 50MC,60MC and 70MC, at an average of0.06 mm/1000h and 0.48 mm/1000h,respectively, are higher than for oursmall bore engines, and that the cylin-
der lube oil dosage at 0.89 g/bhph ismarginally lower. The time betweenoverhauls is, on the average, slightlylonger for the medium bore engines.
The superlong stroke ‘S’ engines, typi-cally used in Cape-size and VLCC ves-sels, have an average liner wear rateof 0.075 mm/1000h, and a ring wearof 0.46 mm/1000h, at an average
0.001
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0.1000.010 1.000 0.001 0.1000.010 1.000
Number of pistonrings measured
0.1000.010 1.000Wear rate in mm/1000 hours
0.001
On 50, 60 and 70MCAverage of 403 cylinders:0.48 mm/1000 h
On S70 and S80MCAverage of 313 cylinders:0.46 mm/1000 h
On 26, 35 and 42MCAverage of 394 cylinders:0.257 mm/1000 h
Fig. 3: Piston ring wear
0.001
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0.1000.010 1.000 0.001 0.1000.010 1.000
Number of cylinders measured
0.1000.010 1.000
Wear rate in mm/1000 hours
0.001
On 50, 60 and 70MCAverage of 547 cylinders:0.06 mm/1000 h
On S70 and S80MCAverage of 407 cylinders:0.075 mm/1000 h
On 26, 35 and 42MCAverage of 455 cylinders:0.037 mm/1000 h
Fig. 4: Cylinder liner wear
4
cylinder lube oil consumption of 1.00g/bhph.
The K-MC/MC-C engines of the Mark3 version are well known for their goodcylinder condition, with low wear ratesand long times between overhauls.
Cylinder liner wear rates down to 0.02mm/1000h and piston ring wear ratesdown to 0.1 mm/1000h are normal, oftenpermitting more than 20,000 hours be-tween piston overhauls.
However, recently some K-engines, es-pecially the K80 and K90 in the Mk 5-6versions, have suffered from scuffing ofthe piston rings, leading to high linerwear rates. Investigations into the originof this problem have revealed that onemajor contributory factor is carbon de-posits on the piston top provoked byhigh lube oil dosage in combinationwith high air humidity. Together, thesefactors may lead to ‘bore polish’, oilstarvation and even scuffing.
Countermeasures to rectify this situationhave been introduced, see below.
New cylinder lube oil system. Testswith a new cylinder lubricating systemhave been initiated. This system, whichis shown in Fig. 6, is based on twogroups of injectors per cylinder whichinject a specific volume of oil into thecylinder for every fourth revolution. Theoil supplied to the injectors is pressurisedby a pump. The system is controlledby a computer in such a way that theoil can be introduced to the individualcylinder at any piston position but, pref-erably, when the piston rings are ad-jacent to the lubricating quills. Thecomputer synchronises itself at eachrevolution, when the piston for cylinderNo. 1 is at top dead centre.
The basic cylinder oil feed rate can beset by means of a screw which limitsthe stroke of the main lubricator piston.
The amount of oil injected can be ad-justed, as required, by increasing orreducing the number of injections, e.g.depending on the actual engine loadchanges.
In the event of malfunctioning of one ofthe injector groups, solenoid valve ortransducer, the oil dosage for the othergroup will automatically be doubled. Ifthe oil pressure fails, the computer willstart a stand-by pump, close the faultypump and set off the alarm.
The benefit of the new system, com-pared with the traditional one, is that itensures more accurate timing, makingit possible to obtain a lower cylinderlube oil consumption. Tests on e.g. a12K90MC look very promising in thatthe same good cylinder condition canbe obtained with a lower cylinder lubeoil consumption. Owing to the fact thatcylinder lube oil represents a significantportion of the total operating costs,such a reduction is very attractive.
Depending on the outcome of prolongedtests on different engines, the new cyl-inder lube oil system may become ourstandard.
0
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1.75g/bhph (MCR)
1.050
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160
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0.451.751.050.45 1.751.050.45
Number of cylinders measured
On 50, 60 and 70MCAverage: 0.89 g/bhph
On S70 and S80MCAverage: 1.00 g/bhph
On 26, 35 and 42MCAverage: 0.90 g/bhph
Number of cylinders measured
Number of cylinders measured
Fig. 5: Cylinder lubricating oil consumptionC
ontro
l val
ve
Pre
ssur
e sw
itch
Pressurized
Pumpstation
Crankshaft position
Pressure control unit
Load changes
Return
Accu-mulator
Accumulator
Control box
Tank
Engine load
Cylinder
Fig. 6: New cylinder lubricating oil consumption
5
Design and ExperienceInteraction
Piston
Piston crown. We have introduced pis-tons with high topland as standard,Fig. 7, with the aim of protecting thepiston rings against the thermal loadfrom the combustion gases. The per-formance of the piston ring pack isthereby improved significantly. Thegradual loss of tension in the pistonrings is reduced, resulting in higherTBO for the pistons.
A further benefit is that the high top-land has made it possible to lower themating surfaces between cylinder linerand cylinder cover, thus reducing thethermal load on the cylinder liner andthus the lube oil film. The high heat in-
put is absorbed by the steel cylindercover, which has a much higher thermalstability than the grey cast iron used forthe cylinder liners. As a consequence,the reliability of the cylinder liner is furtherenhanced by the introduction of thehigh topland.
Tests with ‘high topland’ pistons werestarted about four years ago on anS80MC engine and showed a significantimprovement in the wear condition andan increase in the time between over-hauls from some 4,000 hours to some14,000 hours for the particular engine.
A number of engines with high toplandpistons on all cylinders are in servicewith very good results, an example isshown in Fig. 8.
Piston ring. The reliability of the com-bustion chamber components and the
cylinder condition strongly depends onthe performance of the piston ring packwhich, accordingly, is continuouslybeing optimised.
The two uppermost piston rings arehigher, giving these rings greaterstrength and thermal stability, and thepiston ring material has been optimisedin accordance with the increased ther-mal load that results from the higherrating of the engines.
Some years ago we tested and subse-quently introduced a special patentedCPR (Controlled Pressure Relief) pistonring as the top piston ring, see Fig. 9.This ring has a double lap joint, and anoptimal pressure drop across the toppiston ring is ensured by relief grooves.Furthermore, a ceramic coating hasbeen applied to the top piston ring toincrease wear resistance.
Cylinder 5Cylinder 4
Fig. 8: High topland, 7S70MC Mk 6 after 3000 hours service
Previous standard New standardHigh topland
Topl
and
Fig. 7: Piston with high topland
6
With the increasing mean indicatedpressure, the traditional angle-cut ringgap may result in increased thermalload on the cylinder liner. With the new
CPR piston ring, the thermal load onthe cylinder liner is significantly re-duced as no gas will pass through thedouble lap joint. The relief grooves en-
sure an almost even distribution of thethermal load from the combustiongases over the circumference of the linerand, as a consequence, the thermal loadon the cylinder liner as well as the sec-ond piston ring is reduced. This hasbeen confirmed by temperature meas-urements.
Furthermore, the pressure drop acrossthe top piston ring has been optimisedwith respect to wear on the cylinderliner, piston rings and ring grooves.
Thanks to the double lap joint, thepressure drop will be almost constantirrespective of the wear on the cylinderliner and piston rings. This is in contrastto the traditional angle-cut ring, withwhich the cylinder condition slowly de-teriorates as the liner wears. With theCPR piston rings, a constant good cyl-inder condition and low wear rate canbe expected over the whole lifetime ofthe liner.
On testbed, the pressure drop acrossthe piston ring pack has been measuredwith different configurations of pistonring packs. With the CPR piston ring, itis possible to choose the total area ofthe relief grooves and the optimumnumber of relief grooves, so as to pro-vide the optimum pressure drop acrossthe whole piston ring pack. The opti-mum pressure drop is determined onthe basis of our service experience re-garding running-in conditions and cylin-der liner and piston ring wear rates foreach engine type.
With the controlled pressure relief pistonring, the thermal and mechanical loadson cylinder liner and piston rings arereduced, which means that the CPRpiston ring provides higher reliabilityof both the cylinder liner and the pistonring pack. Consequently, the CPRpiston ring helps to increase the TBOs,as has been clearly confirmed by morethan 20,000 hrs. in service.
A running-in layer of Aluminium Bronzehas recently been introduced on thepiston rings to facilitate running-in. Byvirtue of this, no special running-in pro-
Cylinder cover
Piston
Piston cleaning ring Cylinder liner
Fig. 10: Piston cleaning (PC) ring
Upper piston ring with double-lapS-seal and Controlled PressureRelief (CPR) gaps
Even heat distribution on 2ndpiston ring
2nd, 3rd and 4th piston rings withoblique cut ring gaps
New piston ring material: RVK-C for 50-26and RVK with plasma coating on 98-60
Fig. 9: CPR piston rings for MC engines
7
cedure is necessary, thus saving timeand money in service.
Piston cleaning ring (PC ring). Thepurpose of this ring (Fig. 10) is toscrape off excessive ash and carbonformation on the piston topland andthus prevent contact between thecylinder liner and these deposits,which would remove part of the cylinderoil from the liner wall. Long-term tests onan S80MC engine since 1994 showvery positive results, verifying the hy-pothesis that ‘bore polish’ may be adecisive factor in the deterioration ofthe cylinder condition, especially forhigh-rated large engines. Therefore,PC rings are now standard on themost recent large bore MC engines.
Now more than 80 cylinder units withPC rings are in service.
Cylinder cut-out system. In the caseof low loads, the traditional problem isfouling of the engine due to irregularinjection and atomisation, leading toincomplete combustion.
The irregular injection may be causedby jiggling of the governor, and/or playin the connections in the fuel pumprack control system. The effect in eithercase is that the fuel pumps, when oper-ating so close to the minimum injectionamount, may sometimes just have enoughindex to inject fuel, at other times just notenough index to do so.
By the introduction of a system whereapproximately half of the cylinders arecut out at low speed, the injection intothe remaining working cylinders is im-proved considerably, giving more stablecombustion and, consequently, stablerunning and keeping particle emissionin the low speed range at a minimum.
To avoid that excessive amounts ofcylinder lubricating oil are collected incylinders that are temporarily deacti-vated, the cutting out is made by turnsbetween two groups of cylinders inorder to burn surplus lubricating oiland keep the same thermal load on allcylinders. Turns between the groupsare made on a time basis. The groupseparation is determined in order tohalve the number of active cylindersand to get the smoothest possiblefiring order.
In order to obtain a safe start, the cut-out system is disabled during the start-ing period and until the engine hasbeen stabilised.
The system has been in service for ayear on a series of 11K90MC-C engines,and stable operation down to 13 r/min(MCR is at 104 r/min) is achieved bymeans of this system, see Fig. 11.
Without cylinder cut out
r/min
Cyl. 11
Min. r/min = 13
Min. r/min = 16
With cylinder cut out
r/min
Index
Solenoid valveGroup 1
Solenoid valveGroup 2
Air supply7 bar
0:30 1:30 2:30 3:30 4:30 5:30 6:30
0:30 1:30 2:30 3:30 4:30 5:30 6:30 min:s
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min:s
r/min
Cyl. 1
Fig. 11: Cylinder cut-out system
8
Exhaust valve
Valve spindleThe seats with Alloy 50 on the smallMk 3 engines and Stellite 6 on the largeMk 3 engines have in general performedsatisfactorily.
Nimonic spindles are standard on en-gines of the 60 to 90 bore sizes. Wehave received some queries regardingspindles having many seemingly deepdent marks after a few thousand hoursin service. However, Nimonic has provedvery resistant to blow-by despite dentmarks. The guiding overhaul interval of16,000 hours for Nimonic spindles cannormally be met for the large engines(80-90). For the engine sizes 50-70,our experience indicates overhaul inter-vals of 10,000 to 12,000 hours.
However, we have received severalreports on short overhaul intervalsbecause the crew grind the valve seatsunnecessarily when they see dentmarks. We have therefore issued aservice letter, No. SL 98-356/HH withphotos for comparison purposes, sothat the crew will be better able toassess the condition of the valves.
With regard to grinding, we recommendthat no more than 0.3 mm should beremoved by grinding each time a Ni-monic valve is overhauled – except ifthere is a burn mark. In the latter case,grinding should proceed until the burnmark has been removed. To ensureproper grinding of the Nimonic spindles,a new, special grindstone has beenintroduced.
Disc underside. Service experiencehas confirmed that valve spindles withInconel 625 coated underside, evenunder severe running conditions, canlimit the hot corrosion rate to about0.3-0.6 mm/1000h, depending onengine size. Nimonic spindles have asimilar resistance.
Stem. Due to scattered cases of worn-down chrome layer at the air-springsealing rings, the thickness of thechrome layer was increased from 0.11to 0.3 mm.
Unfortunately, a fissuring problem hasbeen observed in a few cases in thelower part of the chromed area. Lately,we have also experienced some casesof sticking valve spindles which seemrelated to the peeling-off of the chrome,again related to a lack of bonding be-tween the valve stem and the chromelayer.
To improve the wear resistance of thestems and to avoid the use of chromefor environmental reasons, a numberof new surface treatments have beentested. A Cermet coating applied bymeans of the so-called HVOF-proce-dure (High Velocity Oxygen Fuel) hasshown excellent results. The Cermetcoating has now been introduced, to-gether with a matching new type ofseal. However, since this process isnot yet available everywhere, chrome-plating will remain an alternative stand-ard for the next few years.
Bottom pieceSeat. The great majority of Alloy 50and Stellite 6 seats show satisfactoryresults on Mk 3 engines. In new en-gines of the 60-90MC Mk 5 types, asemi-cooled steel bottom piece with ahardened seat has shown excellenttest results, and has subsequentlybeen introduced in combination withthe Nimonic 80A spindle.
O-ring damage. On some 50MC en-gines the lowermost O-ring betweenbottom piece and cylinder cover hasbeen damaged, apparently becauseof the formation of steam from watertrapped in the O-ring groove. Reduc-tion of the outer diameter of the bot-tom piece just above the O-ring grooveeliminates the problem.
Valve housingGas duct. Cold corrosion in the upperpart of the gas duct has caused dura-bility problems on the large Mk 1-3engine types (>50MC), while thesmall-bore engines generally do notsuffer from this phenomenon.
To increase the inside surface tempe-rature and to avoid cold corrosion, aheavy duct housing and reducedcooling have been introduced on the50-90MC engines.
By this means, it has been verified thatcorrosion at the usual locations hasbeen reduced to a negligible level.However, in some cases, especially onS80MC engines, we have observedcorrosion in a ‘new’ location at thebottom of the gas duct near the bot-tom piece. Additional investigationswere carried out, and we decided tochange the water flow. The modifiedwater flow has been tested on two cyl-inders of each of two S80MC engines.Inspections revealed a positive effectand, consequently, it was decided tointroduce the arrangement.
Spindle guide. It is sometimes seenthat the lower third part of the bronzebushing is worn and corroded, however,without disturbing the functioning ofthe valve. In a few cases we haveexperienced shrinking of the bushing,and to avoid this, the bronze materialhas been replaced by surface hardenedcast iron.
Air springIt is essential for the proper functioningof the air spring that the seal betweenthe air piston and the air cylinder, andthe seal between the spindle stem andthe air cylinder, are perfect.
Sealing air. The introduction of a sealingair arrangement supplied with controlair (instead of scavenge air) resulted ina reduction of carbon deposits, etc. inthe sealing air chamber, but the oil mistin the air spring was not able to preventwear of stem and seals in all cases.
We have therefore introduced a sealingair system which incorporates oil mistfrom the air spring chamber. This hasresulted in a cleaner sealing air cham-ber and less wear of the spindle stem,spindle guide and stuffing box seals inthe air spring.
9
The new sealing air system can beretrofitted on valves already in serviceas an ‘add-on’ block. On new valves,the sealing air block is mounted di-rectly on the air cylinder.
To ensure that only air mixed with oilmist is used, a bent pipe is fitted in theair cylinder with the intake well abovethe normal oil level. On large bore en-gines with many cylinders it is particu-larly important to ensure a proper airsupply to the air spring. An insufficientair supply might result in an excessivelyhigh oil level above the intake pipe forsealing air, resulting in camshaft oilconsumption and thus leading to cok-ing and spindle sticking in the spindleguide.
Seals. The sealing ring on the airspring piston usually shows no wear orvery little wear. The seals in the stuffingbox of the air-spring chamber haveoften proved to be worn on the earlytypes of sealing air system. When theoil mist system is used, the standardseals and the chromed stem generallyperform well.
ActuatorWear on oil cylinder, piston and pistonrings is usually extremely low, and thehydraulic actuators are generally caus-ing very few operational disturbances.
Fuel oil system
Fuel injection pump. The so-called‘Umbrella Type’ fuel injection pumpdesign has been used since 1991. Thisinvolves a new sealing arrangementwhich eliminates the risk of fuel oilpenetrating into the camshaft lube oil,and this means that the separate cam-shaft lube oil system can be dispensedwith. Consequently, we have introducedthe uni-lube oil system as standard,whereas the separate camshaft lube oilsystem is still available as an option. Asthere is thus no need for any tank, filters,pumps and piping for the separatecamshaft system, the uni-lube oil sys-tem allows reductions in installationcost, maintenance and space com-
pared with the traditional separatemain lubrication and camshaft lube oilsystems.
Fuel cam and roller. On the large boreengines, skatemarks on the cams androllers are occasionally observed dur-ing running-in. The skatemarks origi-nate from the reversing process wherethe roller shifts position. In most cases,these marks are rather of a cosmeticnature, but they do not look good.Phosphating of the rollers has elimi-nated the skating phenomenon.
Fuel pump reversing link. On anumber of 90MC engines, we have ex-perienced low-cycle fatigue cracks onthe reversing links, as shown in Fig. 12,position A.
Our service records contained no re-ports of earlier cracks or fractures atthat position, but a few incidents hadoccurred with a fractured link fromcracks at position B. However, inthose cases the cause was found tobe a deviation from the specified mate-rial properties.
With regard to the B position, the designwas reviewed with respect to geome-try, material, load and function, and theactual execution (geometry and mate-rial) was also investigated.
Comprehensive measurements ofstresses in the reversing link werecarried out, and the movement of theparts was tested to simulate all possi-ble operating conditions on differentengines in service as well as on testbed.A total of 15 engines were involved inthe investigation.
The conclusions from the investigationwere that the highest stresses comefrom the impact when the link shiftsposition. The link which shifts firstduring the reversing of the engine, andthus at the lowest speed of the cam-shaft, is exposed to the greatest impact.However, malfunctioning when a linkfails to change position, and when fuelinjection starts with the link in thewrong position, also gives rise to high
stresses close to the acceptable in thecritical areas of the link.
The combination of a modified geome-try, an improved air supply, and thepneumatic damping of the movementduring reversing, see Fig. 13 and, forthe larger engines, the introduction ofimproved material has increased themargins considerably. Furthermore,the changes in the pneumatic systemensure that all links shift position duringa reversing sequence, eliminating thesituation where the link on a unit wouldbe in the wrong position.
In parallel with the development work,vessels already in service as well asvessels entering service were equippedwith the latest improved design, to en-sure reliable operation. Follow-up on allthese vessels has shown that the prob-lem has been solved.
Fuel injection valve. Cylinder coverswith three fuel valves are now standardon the large bore 98, 90 and K80MC-C
A
B
Crack incamshaft
side
Fig. 12: Reversing link
10
type engines, leading to uniform heatdistribution and a general reduction ofthe temperature level. The introductionof the three-valve design has the addi-tional advantage that the earlier experi-enced cavitation inside the fuel valvespindle guides has been effectivelyeliminated, thanks to the reduced oilflow through each valve.
Fuel valves
During the last year, our so-called“mini-sac” fuel valves have been inservice on a number of container ves-sels with K90MC engines. The serviceresults show a cleaner combustionchamber and exhaust gas outlet ducts,compared to units with conventionalfuel valves.
The relatively large sac volume in thestandard design fuel nozzle has anegative influence on the formation ofsoot particles and HC. The so-called‘mini-sac’ fuel valve incorporates a con-ventional conical spindle seat as wellas a slide inside the fuel nozzle. The‘mini-sac’ leaves the flow conditions inthe vicinity of the nozzle holes similarto the flow conditions in the conven-tional fuel nozzle, but it has a stronglyreduced sac volume, only about 15%of that of the conventional fuel valve,which has proven to have a positiveinfluence on the cleanliness of thecombustion chamber.
Besides the mentioned benefit, themini-sac valves reduce the formationof NOx during the combustion andthey are standard on engines with IMONOx compliance.
We have recently conducted tests withregular slide type fuel valves on a12K90MC. Slide valves have the advan-tage over the mini-sac valve of havingno sac volume at all, thus the above-mentioned benefits of the mini-sacvalve are even more pronounced forthis type of valve. On the mentioned12K90MC, the application of the slidevalve compared to the mini-sac valvegave a 40% reduction of smoke(BSN10), while HC and CO werereduced by 33% and 42% respectively,though all from a low level. NOx wasreduced by 14%. The fuel consump-tion is virtually unchanged, however,with a slight reduction at part load, seeFig. 14.
The mini-sac fuel valves have nowbeen introduced as our standard forlarge-bore engines. Development work
External damping of reversing link
Air cylinder
Internal throttle valves
Internal throttlevalves
External throttle valve
Non-return valve Non-return valve
Safety valves
innerlips
Optimizedmachined groove
Slightly sloped forinner contact
Edges rounded
Less depth ofgroove aroundreversingshaft pin
A
B
A B
Increasedroundings
New optimized geometry of reversing links Internal damping of reversing link movement
Reversinglink
Reversinglink
Fig. 13: Countermeasures to prevent cracks in reversing link
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continues with a view to further reduc-tion of the sac volumes with a view toimprovement of the emission level.
Bearings
The MC engines were originally designedwith white metal bearings for the cross-head, crankpin and main bearings. Themain bearings were of the so-calledthick shell design, and the other two ofthe thin shell design.
The development towards high specificengine output has resulted in the grad-ual introduction of the thin shell designfor the main bearings too. All new en-gine types, small bore as well as largebore, introduced since the late eighties,have thus been provided with a mod-ern thin shell bearing design, offeringthe following advantages:
• The bore in the housing remains per-fectly circular when the engine is as-
sembled, as it is machined in thetightened condition
• Thin shell bearings normally have ahigher-quality running surface thanthick shell bearings, as the coolingis easier to control during casting
• With a thin shell bearing, it is possi-ble to use the stronger materialSn40Al (tin-aluminium) which, forthe smaller two-stroke engines, hasbeen applied already for the mainbearing with great success. Thebearing shell has no overlay and,consequently, no Ni-layer. Thesebearings have now been in serviceon several engines (S26MC - 42MC)for up to 20,000 hours with good re-sults.
Although the stronger bearing metalcould withstand a higher load, the mainbearing load has not been increased,i.e. the stronger metal is used solely toincrease the safety margin.
The design of the thick bearing shellhas also been continuously improved,most recently by the introduction ofvertical guide pins between the upperand lower shells and modification of thebearing caps to avoid misalignment. Thegeometry of the bearing shells (both
thick and thin shells) is rarely ideal inthe free condition, i.e. when the bear-ings have just been unpacked fromthe supplier. These geometrical vari-ations sometimes make the assemblyof the upper and the lower shells diffi-cult, and misalignment between thetwo halves can create an unintendedoil scraping edge. This misalignment iseliminated by using the vertical guidepins, the design of which makes topclearance adjustment possible by us-ing shims and yet retaining the neces-sary adjustment properties.
Main engine structure
Some years ago, we received informa-tion about cracks in the supporting ribsof the foremost crossgirder in the thrustbearing section on some engines, andrepair methods have been worked outand used on the engines in questionwith good results.
To reduce sensitivity to variations inproduction and service load, we imme-diately modified the welding require-ments and introduced larger roundingsat the rib-ends for new engines, asshown in Fig. 15, and no problemshave been seen on these engines.
Mini sacStandard 90MC
Atomizer sacvolume 520 mm
Atomizer sacvolume 1690 mm3 3
Fig. 14: Fuel valves
Fillet weld
Improved designPrevious design
1/2-V butt weldwith concave finish
QualityEN601M,Q2
Full penetrationK-butt weld
Fillet weld withconcave finish
X - X
YYX
Y - Y
X
Fig. 15: Thrust bearing ribs
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However investigations in connectionwith increasing the engine power, andthus the propeller thrust, have promptedus to modify the thrust bearing housinginto the so-called calliper design. Bymaking the horizontal stiffener platethat connects the fore and aft thrust
bearing supports in one piece, andshaping it like a calliper, see Fig. 16, asignificantly wider design margin hasbeen obtained, even though the engineoutput has been increased at thesame time.
Top bracing
In 1989, we introduced the hydraulicallyadjustable top bracing with the aim ofhaving a system that is capable of cop-ing with the inevitable hull deflectionswhich, because of the use of high ten-sile strength steel in shipbuilding, giverise to excessive stresses in the tradi-tional mechanical type of top bracing.
After that a limited number of teethingproblems have been solved, this systemworks satisfactorily.
The drawback of the hydraulicallyadjustable top bracing, however, is itshigher first cost, compared with thetraditional mechanical top bracing. Forthis reason we have tested a reviseddesign of the mechanical top bracing.
In our efforts to design an improved me-chanical top bracing, we have recentlymade a design with two beams, asshown in Fig. 17. This design has beentested in service since early 1997.
The basic idea with this design is that itis rigid in the athwartship direction andsufficiently flexible in the longitudinaldirection to be able to adapt to move-ments between the engine and the hull.
As of now, 15 vessels are in service withthis design and the service experienceis excellent.
Conclusion
The well-proven MC engine designconcept introduced in 1983 has beenthe backbone of our engine programmeever since. However, continuous devel-opment and design work has beenadded all the time in order to increasethe reliability and the specific power ofthe engines so as to comply with therequirements of the highly competitiveand efficient marine market.
This has only been possible with closecooperation and many open and freediscussions among shipowners, enginemanufacturers and the engine designerover the years.
'Calliper' designB - B
Increased stiffness of thrust bearing support
A
B
C
A - A C - C
A
B
C
Indicated areato be UT-tested
Fig. 16 : Thrust bearing housing
A A
Fig. 17: Top bracing
A-A
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