cimac 2001

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C ON SEI L I NT ERN AT I O NA L INTERNATIONAL COUNCILDES MACHINES A COMBUSTION ON COMBUSTION ENGINES

CIMAC 921 Congress 2001, Hamburg

SWIRL INJECTION LUBRICATION

- A NEW TECHNOLOGY TO OBTAIN LOW CYLINDER OIL

CONSUMPTION WITHOUT SACRIFICING WEAR RATES

Sven Lauritsen, Advisor, Senior Project Manager, Hans Jensen Lubricators A/S

Smedevnget 3, DK-9560 Hadsund, Denmark

Tel.: +45-9857 1911 Fax: +45-9857 1387 Email: [email protected] Dragsted, Senior Superintendent Engineer, A. P. Mller

Esplanaden 50, DK-1098 Kbenhavn K, Denmark

Tel.: +45-3363 3363 Fax: +45-3363 4543 Email: [email protected]

Bert Buchholz, Project Manager, MET Motoren- und Energietechnik GmbH

Erich-Schlesinger-Str. 50, D-18059 Rostock, Germany

Tel.: +49-381 440 3220 Fax: +49-381 440 3212 Email: [email protected]

ABSTRACT

Cylinder wear and cylinder lubrication oilconsumption are essential parameters for theoperating economy of large 2-stroke dieselengines. Therefore it is of great economicimportance to minimise the cylinder wear and/or theconsumption of cylinder oil.

The new HJL SIP lubrication concept meets these

demands. Tests in service have shown a potentialof up to 50% cylinder oil savings without trade-off inliner wear.

The paper describes the system and the methodsused in designing and testing the system, beforefinal testing in service. The results obtained duringapproximately 10,000 hours of service aredescribed and discussed.

INTRODUCTION

During the last 25 years a continuous increase inspecific cylinder oil dosage has taken place for 2-stroke engines. This development has partly beentriggered by the introduction of super long strokeengines where actual service results have provedthis increase particularly necessary. Similarly,results from the other engine types with shorterstroke have led to engine designers

recommendations of increased lubrication.

One may wonder why this development never hasbeen thoroughly questioned, as

a) the sulphur content has by and largeremained unchanged and consequently theTBN of the lubrication oil has remainedmore or less firmly at the level of 70irrespective of lubrication brand.

b) the pure lubrication demand has never

reached the dosages actually applied.

The cost split of keeping a cylinder unit in service,as shown in fig. 1, may indicate why.

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The distribution shown is based on a rather lowspecific cylinder lubrication consumption of 0.65g/BHPh (MCR) when liner and rings are run-in.

It is obvious that the apparent cure for poorcylinder condition, the increased lubrication oilconsumption, has been accomplished on the usersaccount, being the only party to pay for the cylinderoil.

The high lubrication oil dosage is not just a matterof increased costs it is just as well a matter ofenvironmental considerations as engine smokeformation and oil sludge production naturally isincreased with the cylinder oil consumption.

In the light of the above considerations it is quiteobvious that the cylinder lubrication system as suchneeded thorough scrutinisation. The finaldevelopment and testing of a new system is basedon a cooperation and development between theauthors companies.

The result of the endeavours is now available in the

shape of a patent pending system, HJ SIPlubrication (Swirl Injection Principle) [1]. Here a newtechnology has been brought about to enable apossible lower lubrication dosage. The gas swirl inthe cylinders of 2-stroke uniflow engines is used tocarry fine atomised oil droplets to the entireperiphery of the cylinder liner prior to the passing ofthe piston rings. The basic idea of bringing thelubricant close to the top of the cylinder liner wherethe highest wear takes place is not new and hasbeen known in several 2-stroke engines with 2-levellubrication, however, the oil distribution on the liner

sliding surface has previously not been dealt with.The results obtained show a potential for most

users to reduce the lubrication dosage by about50%.

DESCRIPTION OF THE SYSTEM

The main criterion for the project has, asmentioned, been the need for a better initial

distribution of the cylinder oil on the liner surface.Furthermore, it was desirable to improve the timingof the lubrication and, if possible, make it lesscrucial, based on the following conditions:

In most traditional systems the pressure level in thecylinder oil supply tubes is comparatively low,(typically a few bar). The cylinder oil can only enterthe inner liner surface, if the cylinder pressure islower than the pressure in the oil supply tubes.Consequently, the cylinder oil is not necessarilydosed at the optimal position of the piston, in

relation to the position of the oil quill. The oilcompliance due to dissolved air in the oil may evenimply that more lubricator piston strokes areneeded before oil is supplied to the cylindersurface.

In the HJ SIP lubrication system, the dosing takesplace before the piston top reaches the oil quills.Due to their special design the valves in thecylinder liner do not open until a well definedrelatively high pressure in the oil supply tubes hasbeen reached. The lubricator pump stroke is

effected shortly after the scavenging air ports close,during the upward movement of the piston. Thepressure in the oil supply tube will always be somuch higher than the pressure in the cylinder at thetime of dosing that dosing is taking place at everylubricator pump stroke.

Because of the early timing, the timing point is notas crucial as it is in systems where the oil has to besupplied during the very short period when thepiston rings are off the oil quill. The delay due to theultimate velocity of the pressure wave is in thisconnection of less importance, but it can of coursebe taken into consideration when the timing point isdetermined/optimised.

During the upward movement of the piston thescavenging air in 2-stroke uniflow diesel engines issubjected to a powerful rotation at the same time asthe gas is dislocated upwards in the cylinder. Thegas in the cylinder thus follows a helical path or aswirl on its way from the scavenging air ports to theexhaust valve. An oil drop in this swirl will due tothe centrifugal force be forced against the cylinderwall and eventually settle there. This effect is

utilised by introducing the oil portions in the cylinderthrough nozzles as a cone of oil droplets of suitablesize. By adjusting the nozzle dimensions, the oil

Manpower

3%

Piston

rings

2%

Pistons

6%

Cylinder

oil

73%

Cylinder

liners16%

Fig. 1, Typical Cylinder Cost Distribution,K90MC-engines

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outflow speed and the pressure before the nozzle, itis possible to control the system parameters in sucha way, that practically all the oil settles on thecylinder wall before it is caught up by the piston top.

The directions of the nozzles as compared to theflow in the cylinder are arranged in such a way that

the interaction between the individual oil dropletsand the gas flow in the cylinder ensures that the oildroplets hit the cylinder wall periphery over an areacorresponding to 1-2 times the distance betweentwo neighbouring oil quills. Within this area thethickness of the oil layer varies, however with agood coverage over at least an area correspondingto the distance between two neighbouring quills. Inthis way the oil is distributed nearly evenly over thecylinder periphery already before the piston ringpassage. Furthermore, the upward direction of thenozzle will cause the oil to hit the cylinder wallhigher up than the nozzles. If the nozzles areplaced at the same height in the cylinder as theexisting oil quills the oil will then, already whenintroduced in the cylinder, not alone be betterdistributed over the cylinder surface but also bedelivered to the cylinder surface closer to thecylinder top where the need for lubrication is higher.Both of these conditions cause a better utilisation ofthe oil with an expected improved ratio betweencylinder life and lube oil consumption.

The supply of oil to the cylinder surface is effectedtimed and in volumetric measured portions bymeans of a modified HJ timed lubricator.

The principle design of the system is shown in fig.2.

At suitable intervals a number of valves (3) areplaced in the cylinder liner (5), which arecharacterised by being adjusted to open at a certainpressure in the supply tube (2) leading from thelubricator (1) to the individual valves (3). At the endof the valves (3), a little recessed from the innerliner surface, a nozzle (4) is mounted, throughwhich the oil is transformed to droplets when thepressure in the supply tube (2) reaches a certainpre-set value. The valves (3) are designedaccording to the same principles as traditional fuelvalves. The oil leak occurring due to the design islead through a return tube (6) back to the lubricator(1) or the supply tank (7).

The valve design is in principle the same as isknown from traditional fuel valves. They arecharacterised by not opening until the pressure infront of the valve has reached a pre-set pressure,which is the pressure necessary for the formation ofoil droplets. Furthermore, there is a very smallvolume between the valve seat and the nozzle,which minimizes after-drops and ensures thatdosing starts practically at the same time as thevalve opens.

Fig. 2, HJ SIP Lubrication System

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The valve design appears from fig. 3: The pressure

tube is connected to the inlet (2) and the oil is leadthrough a strainer unit (3) to the space (5) betweenthe valve housing (4) and the inner part (7). Fromthere further on to the inclined holes (9) in the innerpart which also contains the seat for the valveneedle (8). The valve is kept closed by the spring(6). When the axial force on the valve needle fromthe oil pressure exceeds the spring force the valveopens for the oil to enter the space (12) in front ofthe nozzle opening. The spring force and therebythe opening pressure of the valve can be adjustedby means of the adjusting screw (1). The leak oil is

carried off through the leak oil connection (10). Thevalve housing is mounted in a bore in the cylinderliner (11) and fastened with two screws as shown infig. 4.

Fig. 2 shows the valves mounted radially in thecylinder liner. This mounting is clearly the mostsimple, however, outer conditions may necessitatea non-radial mounting. Such possibilities are shownin fig. 4.

The system may be mounted on new engines or

retrofitted on engines already in operation.

PLANNING

By the end of year 2000, the system has beenoperating on two cylinders on a MAN B&W12K90MC in a sailing vessel for approximately

10,000 operation hours. As mistakes in connectionwith such a test can be extremely costly, anextensive 3-step design and verification programmepreceded the installation:

1. Before designing the system a series ofcomputer simulations were carried out withthe purpose of making the expected effectof the system probable. The simulationsare based on a dynamic model of the gasflow in an engine cylinder in which the oil isintroduced through nozzles, as described

above.

2. Design of valves etc. for the system.

3. With the purpose of demonstrating the oilsettling on the cylinder wall and the functionof the system on the whole, a test rig in theform of a helical wind tunnel, fig. 5, wasdesigned and constructed. In full scale itsimulates the scavenging air flow in thecylinder at the expected injection time ofthe cylinder oil. The results from the tests

were in good conformity with the computersimulations and the expectations to thesystem. Computer simulations of the airflow in the wind tunnel formed the basis forthe actual wind tunnel dimensions.

COMPUTER SIMULATIONS

Three major areas for the utilisation of simulationtools within the project were defined. The first partwas to determine the gas flow conditions inside the

cylinder in order to obtain knowledge about theenvironment into which the oil spray was to be

Fig. 4, HJ SIP Lubrication Valves, MountingPossibilities

Fig. 3, HJ SIP Lubrication Valve

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injected. The second part was to determine a spraytype which is able to produce an appropriate oil filmat the cylinder liner surface. Finally, the geometry ofa wind tunnel for testing different valves andprototypes was developed.

Determination of flow conditions inside the cylinder

The main targets for the oil film were easily found:distribution around the full liner circumference, evenwall film thickness, horizontal position andemerging time according to the engine dynamics,maximised efficiency (oil mass injected/oil mass inwall film).

What could not be answered readily was thequestion of which spray parameters would benecessary to produce such an oil film. Whereasspray lubrication is standard technology in severalindustrial applications, the utilisation of spray forlubrication inside a cylinder of a 2-stroke diesel

engine is a completely new application. The mainchallenges to be faced were the strong andconstantly changing gas flow conditions inside thecylinder.

Consequently, the determination of the flowconditions inside the cylinder of a large 2-strokecross-head diesel engine was the task. It wasdecided to carry out a CFD-Analysis(Computational Fluid Dynamics) of the flowprocesses inside the cylinder of an MAN B&WK90MC engine during scavenging and

compression. A detailed virtual model of thecylinder was generated. It consisted of the cylinderliner with its inlet ports and a part of the air box, thecylinder cover with exhaust valve, exhaust valvehousing with bottom piece, and the piston. Themovements of piston and exhaust valve weremodelled to allow for scavenging flow andcompression. Using this model, a transientsimulation of the turbulent flow from begin ofscavenging till the end of compression (piston inTDC) was carried out. The fluid properties of thescavenging air and the exhaust gas were

considered as well as the component temperaturesand pressure values. Fig. 5 shows the velocity fieldinside the cylinder for a position of 220 crankshaftusing a vertical cutting plane through the computermodel, and a horizontal cutting plane A-A asshown. The latter is enlarged for clarity.

The results showed a fairly strong and stable,reproducible swirl flow inside the cylinder for asignificant period of time between end ofscavenging and end of compression. This flow wasanalysed regarding its velocity componentsrotational and upwards), near wall profile,temperature and pressure.

These results were assessed by all partners using

their individual specific experience and the timewindow most suitable for lubricating oil injectionwas fixed.

Fig 5, Calculated Axial and Rotational GasVelocity Field in Cylinder

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Determination of spray parameters and nozzlegeometry

In the second part of the simulation-based pre-development phase the interaction between thedroplets of a spray and the general flow inside thecylinder of an MAN B&W K90MC engine wasinvestigated. The target was to determine theinfluence of different spray parameters on thesprays capability to produce a wall film. Thesolution approach chosen for this investigation wasbased on a CFD-simulation strategy.

A sub-model of the cylinder was used and thesimulated flow field from the first step of theinvestigation was transferred on this sub-model.The formation of a spray cloud from a nozzle, theinfluence of the air flow and of the cylinder linersurface (wall) was predicted using additional, tailor-made numerical sub-routines. Thus, the majorproperties of the injected oil (density, viscosity,

temperature, surface tension, mass etc.), theinjection nozzle parameters (direction, spray coneangle, injection timing, injection velocity etc), the airflow (velocity components, pressure, temperature)and the liner surface parameters (shape andposition) were all taken into account for thesimulation of the wall film (position, size, thickness).

The influence of different spray parameters couldthus be investigated very efficiently. A considerable

number of different cases was analysed to obtain agood overview on spray parameters and theirinfluence on the wall film generation. Spray patterns

for two different injection angles are compared inFig. 6. The results of all cases were assessedregarding wall film parameters such as mass,covered area, shape etc.

The optimum spray parameters were establishedand the main nozzle parameters necessary togenerate such a spray were derived. These nozzleparameters were: the hole diameter, ratio of holediameter/hole length, spray direction, spray coneangle, operation pressure and pump stroke length.Unfortunately, some of these parameters cannot bedirectly determined by means of calculations orsimulations. Consequently, the fine tuning had to bedone using experiments.

Development of wind tunnel design

The subject investigated in the final part of thesimulations was the layout of an appropriate test

channel. This channel was to be used forexperimental investigations of different nozzles withrespect to sprays and wall films produced by thesenozzles. Therefore, the test channel not only had torepresent the general geometrical dimensions ofthe cylinder but also the typical flow field inside thecylinder as it was determined during the first part ofthe investigations.

Thus, the target was to fix the layout of the testchannel in such a way, that the shape of thechannel produces an internal flow field similar to theflow conditions inside the cylinder. A second targetwas to minimise the dimensions necessary for thistest channel in order to simplify the handling of thechannel and to minimise the size of the laboratoryneeded.

As a result of the simulations the geometry of thechannel was fixed, the segment to be used for theexperiments was described and advise was givenhow to improve the inflow conditions into the windtunnel. See fig 7.

FULL SCALE LABORATORY TESTS

The purpose of the laboratory tests was todemonstrate the results from the computersimulations under conditions as close to reality aspossible before mounting the equipment incylinders in operation. In this way the largestpossible safety for eliminating mistakes beforemounting in a vessel was obtained. A test rig of aprinciple design as shown in fig. 7 was designedand constructed for the laboratory tests.

The central part of the test rig is designed as atwisted wind tunnel made from transparent plastic(acrylic). The diameter of the inner side of the

Fig. 6, Calculated Spray Pattern for TwoDifferent Spray Angles

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tunnels outer wall is 900 mm, corresponding to thecylinder diameter of the MAN B&W K90MC engineon which the equipment was later to be mounted.

The cross-sectional dimensions of the tunnel are200 x 400 mm. A powerful blower supplies thetunnel with air flow in the twisted part of the tunnelcorresponding to the ultimate velocity of thescavenging air in the engine at the time of dosing.Between the blower and the twisted part of thetunnel nearest to the blower a grid consisting of apile of 140 mm long tubes, with their centre linesparallel to the flow, have been inserted togetherwith a straight piece of wind tunnel of 2.25 meters.The purpose is to stabilise the flow before it entersthe twisted part of the tunnel. For the same reason

there is a non-transparent piece of straight tunnelafter the twisted part of the tunnel. The transparenttwisted part of the tunnel allows visual observationof the dosing process and the resulting distributionof the oil on the tunnel surface.

The valve injecting the oil is mounted in a cylindricalcover in the curved outer surface of the tunnel. Theinner side of the cover is flush with the cylindersurface in order to avoid turbulence. The cover isattached by a special holder, facilitating mountingand dismounting of the valve and allowing accessto the inner side of the tunnel for cleaning betweenthe tests. The valve is placed somewhat after thechange between the straight and the twisted part ofthe tunnel, but in the lower half of the tunnel. At thisplace the effect of the flow shifting between the

straight and the twisted part has decreased somuch that it has no practical significance. Fig. 8shows a photo of the actual test rig where the valve

position can be seen.

Air velocities in the tunnel are measured by meansof a Pitot tube. Fig. 9 shows the variation of the airvelocity over the cross-section of the wind tunnel,measured in the straight part of the wind tunnel justbefore the entrance to the twisted part. A modifiedtraditional timed Hans Jensen lubricator wasapplied for dosing the oil to the valves.

Cross-sectional dimensions, climbing angle andpositioning of the dosing valve in the tunnel were

determined based on computer simulationsdescribed elsewhere in this paper.

The static pressure in the wind tunnel was theatmospheric pressure and the air temperature wasapproximately 20 degrees Celsius. Bothparameters deviate from the values prevailing in anengine in operation. It would, however, be verydifficult to consider these in the test. In order tocompensate best possible for the temperatureinfluence an oil was applied, which viscosity atroom temperature is the same as the cylinder oilviscosity at liner temperature which is presumed tobe approximately 85 degrees Celsius. It wasestimated that the risk in connection with thedeviations was minimal due to the high pressure inthe oil supply tube and based on considerations

Fig. 8, Wind Tunnel

Pressure-Transducer

Wind tunnel

SIP Valve

Oil supply

Blower

Fig. 7, Wind Tunnel, Principle

Cylinder liner(900mm)

SIP Valve

Wind Tunnel

Blower

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regarding among other things evaporation of oilwhich was estimated to be minimal at the prevailingtemperatures.

Two dynamic pressure gauges are inserted in thepressure tube. One immediately after the lubricatorand one immediately before the dosing valve. Inconnection with the carrying out of the full scaletests it was verified that the oil compressibility andthe tube compliances with the applied valves haveno negative effect on the timing even in supply

tubes up to more than 6 meters long. The testsshowed that oil is dosed at each lubricator pumpstroke at the pump frequencies and the dosingrates valid for the actual engine.

Fig. 10 shows a representatively simultaneouspressure recording measured on the two pressuregauges. From this appears that the time delay forthe pressure wave is approximately 5 millisecondsfor a 6.2 m tube length, corresponding to a velocityof approximately 1,200 m/s, i.e. approximately thesonic speed in the oil. This time delay can simplybe included when the timing is determined. Duringthe same period the engine piston has movedapproximately 60 mm. Even a considerableuncertainty as far as the delay is concerned istolerable, as the demand for the timing is far lessstrict than for systems, where the oil is dosedduring the passage of the piston rings.

The results shown in fig. 10 are obtained with theoil used for the wind tunnel tests. Similar tests withthe correct cylinder oil have confirmed theseresults.

To save time, the tests have been concentrated on

an MAN B&W K90MC engine. It is reasonable toassume that as test and computer simulationsconform so well with each other, as is the case,computer simulations should be sufficient basis forimplementation in other engine types and sizes.

The period from the lubricator pump stroke and untilthe oil settles on the cylinder wall is so short, that itis difficult to follow it with the naked eye, andtraditional video recordings are too slow. Therefore,video has been recorded with high speed samplingequipment with a picture frequency of 1,000

pictures per second. The video sequences haveconfirmed the assumption that the oil acts asexpected as compared to the presentation and thecomputer simulations.

Fig. 9, Air Velocities across the Wind TunnelCross Section

Fig. 11, Oil Distribution on Wind Tunnel Wall typical Example

Quills

Fig. 10, Pressure Recordings, immediately afterthe Lubricator and immediately before the Valve

6,7

6,9

7,1

7,3

7,5

7,7

7,9

8,1

8,3

8,5

2,17 2,175 2,18 2,185 2,19 2,195

Time [msec]

Pressure[MPa]

Pipe length 6.2 [m]

pressure at valve

pressure at lubricator

5[msec]

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By running a few pump strokes with the lubricatorand manually drawing the distribution of the oil on atransparent piece of paper on the transparenttwisted part of the wind tunnel, a series of drawingsof the distribution contours have been produced.These show good conformity with the predictions ofthe computer simulations. Fig. 11 shows areproduction of such a distribution graph. Thedarkest areas represent the thickest layer of oil.

The cone angle of the oil droplets has in the testsshown to be somewhat smaller than assumed inthe simulations but the oil distribution on thecylinder wall showed to be in reasonably goodconformity with the simulations.

There has been no attempt to collect the oil settlingon the wind tunnel wall in order to verify, in thatway, that the cylinder wall benefits from all the oil.However, it turned out that it was not possible todetermine oil settling in the straight part of the wind

tunnel after the twisted part, even after a longperiod of running the equipment. This is interpretedas proof that the cylinder surface actually benefitsfrom all the oil.

It has furthermore been demonstrated that amodified traditional HJ lubricator is capable ofdelivering the pressure necessary for the formationof oil droplets. Tests have been carried out at 80bar without problems, however, based on the testresults, the pressure level for the operation testswere set at 35-40 bar.

TESTS IN SERVICE

The practical test has been carried out onboardM/V Sovereign Maersk, a container vessel tradingbetween Europe, the Far East and the west coastof the United States. The vessel is equipped with a12-cylinder MAN B&W K90MC, mark VI enginehaving the following particulars of special interestfor this test, see table 1.

This engine type has, since the very first went intoservice, seen quite a development in fuelequipment in order to combat heavy coke depositsin the combustion chamber and the exhaustsystem. The coke formation on the piston top landreached such dimensions that it wiped off the oilfrom the cylinder sliding surface. As a consequenceof the above the standard fuel valves wereexchanged with so-called slide valves as described

by Pedersen [2], and furthermore a piston cleaningring, (PCR) positioned in the top of the liner wasintroduced to prevent coke from

touching the liner, as described by Mikkelsen &

Bryndum [3].

The PCR was installed in all cylinders prior tostarting the test and the slide valves wereintroduced during the tests as explained below.

The average wear rate of the liners until the slidevalves were introduced were 0.150 mm/1000h(scuffed liners excluded) at a cylinder oil dosagebetween 0.65 and 0.85 g/BHPh (MCR) afterrunning-in of the liners and rings.

It was decided to install the new lubrication systemon two cylinders, with cylinder 7 as the first andcylinder 12 next - after thorough test of systemcomponents in cylinder 7. This decision was provedright, as some components actually neededmodification before functioning as expected.

At the time of writing the system has been in

operation for about 10,000 hours and apart frommeasuring the liners in the traditional manner andmonitoring the cylinder oil dosage, a special test of15 days duration with an activated top piston ringhas been performed.

Traditional wear measurements

In fig. 12 the cylinder lubrication oil dosage and thecylinder liner wear is shown covering the entire testperiod (that is to be continued for some time).

It is clearly seen that some difficulties wereencountered on cylinder 7 during the first nearly4,000 running hours, hence the dotted line in thatperiod for the liner wear and similarly for cylinder 12for the first 1,000 hours. What went wrong wasunforeseen lubricator adjustment problems leadingto an uncontrolled reduction of cylinder oil dosage.Furthermore the timing needed adjustment. It is,however, worth to mention that no scuffing occurredon liner and rings during that period.

Overall engine data:

Cylinder Bore, mm 900

Stroke, mm 2550

Revolutions, rpm 94

Mean Effective pressure, bar 18

Max. Cylinder Pressure, bar 140

Cylinder configuration:

Dual cast, split Cylinder Liners with PCR ring

CPR, RVK,PM14 plasma-coated, top piston ring

Texaco Taro Special Cylinder Oil, TBN 70

Table 1, Particulars of the Test Engine,MAN B&W 12K90MC

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After the rectification of the system thedevelopment in liner wear has been rather

moderate. In the latest period the wear rate has

been less than 0.02 mm/1000 h. At the same timethe lube oil dosage has gradually been reduced andthen kept at a fairly low level, in cylinder no.12down to 0.51 g/BHPh (MCR). It is of course difficultto measure such a low wear development with ahigh accuracy using the traditional cylindercalibration gauge. However, the wear pattern on theliner surface has since the start of the test beencarefully observed and the slow disappearance ofthe wave cut has thus confirmed that themeasurements were reasonably accurate.

As one of the features of this new lubricationtechnology is to distribute the oil evenly high up in

the cylinder liner and above the standard lube oil

quills of such engines one could fear that this wouldlead to oil starvation in the lower part of the liner.The wear measurements taken from top to bottomand in four diagonal directions show, however, thatthe specific wear all over the liner is very low andactually below some of the best results obtained ina similar, comparable engine with the traditionallubrication system. Fig. 13 shows a comparison ofwear distribution in the liner between the traditionaland the spray lubrication system. The lube oildosage for the traditional system is 0.65 to 0.7g/BHPh to be compared with the dosage shown for

cylinder no. 12 in fig 12.

0,0

0,1

0,2

0,3

0,4

0,5

0,6

0,7

0,8

0,9

1,0

1,1

1,2

1,3

1,4

1,5

1,6

1,7

1,8

1,9

2,0

10000 11000 12000 13000 14000 15000 16000 17000 18000 19000 20000

Engine running hours

Liner

wear,mm&lubeoildosage,g/BHPh(MCR

Dosage cyl 7

Liner wear cyl 7

Dosage cyl 12

Liner Wear cyl 12

Standard Fuel Valve Slide Valve

Fig. 12, Cylinder Oil Dosage and Liner Wear during the first 9,000 Hours Testing

0.00

0.02

0.04

0.06

0.08

0 500 1000 1500 2000

Distance from Liner Top, mm

SpecificWear,mm/1000 6334 h, tradit. lubrication

6037 h, spray lubrication

Fig. 13, Cylinder Liner Wear Pattern for aTraditional and a SIP Lubrication System

Fig. 14, Cylinder Wear and Sulphur Contentof Fuel Oil during the Test Period

0,0

0,5

1,0

1,5

2,0

2,5

3,0

3,5

4,0

4,5

10000 12000 14000 16000 18000 20000

Engine running Hours

Wearmm,FuelSulphur%

Sulphur Cyl. 7 Cyl. 12 Control cyl.6

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In order to compare the absolute wear resultobtained in the test cylinders and to the highestpossible degree eliminate any uncertainty,reference cylinders on the same engine have beenselected. Actually the first 6 cylinders of the engineis now equipped just as the two test cylinders inrespect of liner type, piston ring configuration andfuel valves. A comparison of the wear developmentof one of the reference cylinders (Cyl. No. 6) andthe two test cylinders is shown in fig. 14. The wearrate of cylinder 6 has during the test period been0.172 mm/1000h at a lubrication oil dosagebetween 0.65 and 0.7 g/BHPh (MCR).

In fig. 14 is also shown the sulphur content of thefuel bunkered during the entire test period. Thetrendline show that the sulphur content has beenlow between 15,000 and 19,000 hours. This is,however, not reflected in the wear rates, the wearof the control cylinder is actually increased duringthat period.

During most of the test cylinder drain oil analysishas been performed covering all cylinders. Onewould expect that the TBN of the drain oil from thetwo test cylinders would be reduced relative to theother cylinders as the dosage in test cylinders isrelatively low. This is not the case, as seen in fig.15. It is not possible to identify the test cylindersbased on the TBN of the drain oil. For the timebeing no explanation can be offered on thisphenomenon.

A parameter of great influence on absolute cylinderwear measured over a period of time is the engineload. This has previously been discussed by

Dragsted [4], who suggested a so-called loadprofile number, LPN to be used in the evaluation ofthe severity of the engine operation (LPN = theaverage load in percent of MCR + the number ofrunning hours at a load higher than 80% MCR inpercent of all running hours).

The LPN for the MAN B&W 12K90MC engineduring the discussed test period has been 132,which according to previous experience isconsidered as rather high even for container vesselengines. The influence of engine load on piston ringwear will be demonstrated in the following section.

Radio-Nuclide measurements

To determine the influence of engine load on thewear at the piston ring/cylinder liner system underreal operation conditions a sophisticated wearmeasurement was carried out at one of thecylinders equipped with SIP lubrication.

The measuring technology used was the Radio-Nuclide Thin Layer Difference (RN-TLD)-Method,which works without any disturbance of thetribologic system. The RN-TLD-Method is explainedin [5], The typical measurement set-up chain isshown in fig. 16. In addition to previous

measurements which could be carried out onlywhen the engine was stopped (off line), the latestdevelopments in the RN-TLD-Method allow so-called online measurements, i.e. measurementswith the engine running at arbitrary speed and load.The basic concept of the RN-TLD-Method consistsof the generation of radioactivity in the surface ofthe component which is to be investigated (i.e. thepiston ring). To accomplish this, the plasma coatedCPR-ring was exchanged with a standard, non-coated, oblique cut piston ring. Is this surfaceexposed to abrasive wear, a loss in radioactivity

can be measured using highly sensitive probes.The actual wear rate can be determined bycomparing the measured loss in activity with acalibration curve and taking into consideration thenatural loss of activity as well as the geometry andsurface roughness of the investigated component.The necessary levels of radioactivity are extremelylow allowing measurements aboard vessels as wellas at test beds.

0

5

10

15

20

25

30

35

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1 2 3 4 5 6 7 8 9 10 11 12

Cylinder Number

TBN

Fig. 15,Analysis of Cylinder Drain OilFig. 16, Schematic View of WearMeasurement at a Piston Ring

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Results

Online measurements were carried out throughoutthe whole voyage, off-line measurements weremade during port times. The results of bothmeasurements were in good agreement. Only littlerunning in effects were observed.

- The measured wear rates of the top piston ringvaried between 0.01mm/1000h and0.065mm/1000h. Tendencies between oilinjection timing and wear, oil injection dosageand wear and load and wear could bemeasured. Fig. 17 shows a summary of theresults of all measurements.

- Measurements with similar engine operationconditions led to similar wear rates confirming areproducible functionality of the system.

- The cylinder oil timing chosen for the entire

service test (standard timing) showed agenerally low wear rate, which increased withincreasing engine load (hatched area in fig 17).

- The measurements with changed injectiontiming confirmed the fact that the SIP lubricationsystem is not sensitive to such changes,considering the big steps of plus/minus 15degrees crank angle. Generally, the variationsin the injection timing led only to small changesin the wear rates which, however, all are at avery low level. The earlier timing led to

increased wear at higher loads whereas thelater injection timing did not show any loaddependence.

- The oil dosage was reduced in several steps.The measured wear rates did not increase evenfor the lowest dosage of 0.44 g/BHPh (MCR).

The potential of the system to reduce lubricating oilconsumption without trade-off in liner wear wasconfirmed.

During all measurements no abnormal wearphenomenon were recorded. This was confirmedby the final inspection of the liner at the end of themeasurements. The wear rates remained at a verylow level during all the tests.

CONCLUSION

The experiments and tests carried out with a newlubrication system, based on the new technology ofinjecting cylinder oil into the air swirl in the engine

cylinder, prove that the system can work asexpected with low cylinder and piston ring wear

rates even with very low cylinder oil dosages. Thescope for cylinder oil savings will for most users beclose to 50%. The duration of the tests onboarddoes of course not correspond to normal serviceexpectations, but it is also the intention to continuethe tests and further to increase the number ofengines to be equipped with the system producedaccording to normal manufacturing standards (nottest equipment).

REFERENCES

[1] Patent application PCT/DK99/00599Lubricating system for large dieselengines

[2] Pedersen, P. S., The Intelligent Engine One Solution to NOx Reduction, ISME2000-PD-22.

[3] Mikkelsen, U., & Bryndum, L., Large BoreEngines Problems and Countermeasures,The Motor Ship Marine PropulsionConference 2000.

[4] Dragsted, J. Engine Cylinder Condition,from delivery and some 10 years ahead.CIMAC Congress London 1993, paper D15.

[5] Bludszuweit, S., Schwarte, J., Prescher, K.& Richter, B., A Super-Sensitive Methodfor Wear Measurements Inside Ship

Engines, CIMAC Congress Copenhagen1998.

Fig. 17, Measured Wear Rates dependenton Engine Load, Lubrication Oil InjectionTiming and Oil Dosage.

cimac 2001

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