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LNG Carriers with ME-GI Engine and High Pressure Gas Supply System

Introduction ......................................................................... 3

Propulsion Requirements for LNG Carriers with Dual-Fuel Gas Injection ...................................................... 4

Fuel Gas Supply System – Design Concept ...................... 5

Fuel Gas Supply System – Key Components .................... 6

Capacity Control – Valve Unloading .................................. 9

Compressor System Engineering – 6LP250-5S ................. 10

ME-GI Gas System Engineering .......................................... 11

ME-GI Injection System ....................................................... 12

High-Pressure Double-Wall Piping ..................................... 13

Fuel Gas System - Control Requirements ......................... 15

Machinery Room installation – 6LP250-5S ....................... 18

Requirements for Cargo Machinery Room Support Structure ..................................................... 19

Requirements for Classifi cation ......................................... 20

Actual Test and Analysis of Safety when Operating on Gas .......................................... 20

Main Engine Room Safety ................................................. 20

Simulation Results .............................................................. 21

Engine Operating Modes ................................................... 22

Launching the ME-GI .......................................................... 23

Machinery Concepts Comparion ........................................ 24

Concluding Remarks ................................................................... 28

References ................................................................................... 28

Appendices: I, II,III, IV, V, VI,VII ................................................... 28

MAN Diesel A/S, Copenhagen, Denmark

Contents:

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LNG Carriers with ME-GI Engine and High Pressure Gas Supply System

Introduction

The latest introduction to the marine market of ship designs with the dual- fuel low speed ME-GI engine has been very much supported by the Korean shipyards and engine builders, Doosan, Hyundai, Samsung and Daewoo.

Thanks to this cooperation it has been possible to introduce the ME-GI en-gines into the latest design of LNG car-riers and get full acceptance from the Classifi cation Societies involved.

This paper describes the innovative de-sign and installation features of the fuel gas supply system for an LNG carrier, comprising multi-stage low temperature boil-off fuel gas compressor with driver and auxiliary systems, high-pressure piping system and safety features, controls and instrumentation. The paper also extensively describes the operational control system required to provide full engine availability over the entire transport cycle.

The demand for larger and more energy effi cient LNG carriers has resulted in rapidly increasing use of the diesel en-gine as the prime mover, replacing tra-ditional steam turbine propulsion plants.Two alternative propulsion solutions have established themselves to date on the market:

low speed, heavy fuel oil burning die-sel engine combined with a relique-faction system for BOG recovery

medium speed, dual-fuel engines with electric propulsion.

A further low speed direct propulsion alternative, using a dual-fuel two-stroke engine, is now also available:

high thermal effi ciency, fl exible fuel/gas ratio, low operational and instal-lation costs are the major benefi ts of this alternative engine versionthe engine utilises a high-pressure gas system to supply boil-off gas at pressures of 250-300 bar for injection into the cylinders.

Apart from the description of the fuel gas supply system, this paper also discusses related issues such as re-quirements for classifi cation, hazardous identifi cation procedures, main engine room safety, maintenance requirements and availability.

It will be demonstrated that the ME-GI based solution has operational and economic benefi ts over other low speed based solutions, irrespective of vessel size, when the predicted criteria for relative energy prices prevail.

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Propulsion Requirements for LNG Carriers with Dual-Fuel Gas Injection

In 2004, the fi rst diesel engine orderwas placed for an LNG carrier, equip-ped with two MAN B&W low speed 6S70ME-C engines. Today, the order backlog comprises more than 90 en-gines for various owners, mainly oil companies, all for Qatar gas distribution projects.

While the HFO burning engine is a well known and recognised prime mover, the low speed dual-fuel electronically controlled ME-GI (gas injection) engine has not yet been ordered by the market.

Although the GI engine, as a mechani-cally operated engine, has been avail-able for many years, it is not until now that there is real potential. Cost, fuel fl exibility and effi ciency are the driving factors.

The task of implementing the two-stroke ME-GI engine in the market has focused on the gas supply system, from the LNG storage tanks to the high-pressure gas compressor and further to the engine. A cooperation between the shipyard HHI, the compressor

LNG carrier size (cum)

Recommendedtwo-stroke solution

Propulsion power (kW)

Propulsion speed (knots)

Beam/draft ratio

Estimated gain in effi ciencycompared to a single propeller

145,000-150,000

2 x 6S60ME-GI2 x 5S65ME-GI

2 x 14,2802 x 14,350

19-21 3.8 5%

160,000-170,000

2 x 5S70ME-GI 2 x 7S60ME-GI

2 x 16,3502 x 16,660

19-21 4.0 > 5%

200,000-220,000


2 x 17,2202 x 19,620

19-21 4.2 9%

240,000-270,000


2 x 20,0902 x 21,770

19-21 4.5 > 9%

manufacturer Burckhardt Compression, AG (BCA), the classifi cation society and MAN Diesel has been mandatory to ensure a proper and safe design of the complete gas distribution system, including the engine. This has been achieved through a common Hazid / Hazop study.

Confi guration of LNG carriers utilising the boil-off gas

The superior effi ciency of the two-stroke diesel engines, especially with a directly coupled propeller, has gained increasing attention. On LNG carriers, the desired power for propulsion can be generated by a single engine with a single propeller combined with a power take home system, or a double engine installation with direct drive on two pro-pellers. This paper concentrates on the double engine installations

2 x 50 %, which is the most attractive solution for an LNG carrier of the size 145 kcum and larger. By selecting a twin propeller solution for this LNG car-rier, which normally has a high Beam/draft ratio, a substantial gain in propeller effi ciency of some 5 % for 145 kcum and larger, and up to 9 % or even more for larger carriers is possible.

Redundancy in terms of propulsion is not required by the classifi cation socie-ties, but it is required by all operators on the LNG market. The selection of the double engine ME-GI solution results not only in redundancy of propulsion, but also of redundancy in the choice of fuel supply. If the fuel gas supply fails, it is possible to operate the ME-GI as an ME engine, fuelled solely with HFO.

For many years, the LNG market has not really valued the boil-off gas, as this has been considered a natural loss not accounted for.

Today, the fuel oil price has been at a high level, which again has led to con-siderations by operators on whether to burn the boil-off gas instead of utilis-ing 100 % HFO, DO or gas oil. Vari-ous factors determine the rate of the boil-off gas evaporation, however, it is estimated that boil-off gas equals about 80-90 % in laden voyage, and in ballast voyage 40-50 % of the energy needed for the LNG vessel at full power. There-fore, some additional fuel oil is required or alternative forced boil-off gas must be generated. Full power is defi ned as a voyage speed of 19-21 knots. This speed has been accepted in the market as the most optimal speed for LNG car-

Table I: Two-stroke propulsion recommendations for LNG carriers in the range from 145-270 kcum

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riers when both fi rst cost investment and loss of cargo is considered.

To achieve this service speed, a two-stroke solution for the power require-ment for different LNG carrier sizes is suggested in Table I.

With the high-pressure gas injection ME-GI engine, the virtues of the two-stroke diesel principle are prevailing. The thermal effi ciency and output re-main equivalent to that obtained when burning conventional heavy fuel oil. The high-pressure gas injection system offers the advantage of being almost independent of gas/oil fuel mixture, as long as a small amount of pilot oil fuel is injected for ignition.

LNG Tank Compressor Oxidiser

ME-GI

ME-GI

ME-GI PSC Clutch

FPP

Compressor

High pressure gas

Fig. 1: LNG carrier with the recommended ME-GI application.

In order for the ME-GI to achieve this superior effi ciency of 50 % (+/− 5 %fuel tolerances) during gas running, the gas fuel requires a boost to a pressure of maximum 250 bars at 100 % load. At lower loads the pressure required decreases linearly to 30 % load, where a boost pressure of 150 bars is re-quired. To boost this pressure, a high-pressure compressor solution has been developed by BCA, which is presented in this paper.

Fig. 1 shows an example of an LNG carrier with the recommended ME-GI application.

Fuel Gas Supply System – Design Concept

The basic design concept of the fuel gas supply system presented in this paper considers the installation of two 100 % fuel gas compressors. Full redundancy of the fuel gas compressor has been considered as a priority to satisfy classifi -cation requirements (see Fig. 2).

Each compressor is designed to deliver the boil-off gas at a variable discharge pressure in the range of 150 to 265 bar g (15–26.5 MPa g), according to required engine load to two 50 % in-stalled ME-GI engines A and B. The selected compressor runs continuously, and the standby compressor is started manually only in the event of malfunc-tion of the compressor selected.

The amount of boil-off gas (BOG), and hence the tank pressure, varies consid-erably during the ship operating cycle. The design concept therefore requires that the compressors be able to oper-ate under a number of demanding con-ditions, i.e. with:

a wide variation of BOG fl ow, as ex-perienced during loaded and ballast voyage,

a variation in suction pressure ac-cording to storage tank pressure,

a very wide range of suction tempera-tures, as experienced between warm start-up and ultra cold loaded opera-tion, and

a variable gas composition.

The compressor is therefore fi tted with a capacity control system to ensure gas delivery at the required pressure to the ME-GI engine, and tank pressure con-trol within strictly defi ned limits. These duty variables are to be handled both simply and effi ciently without compro-mising overall plant reliability and safety.

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The compressor is designed to effi -ciently deliver both natural boil-off gas (nBOG) and, if required, forced (fBOG) during the ballast voyage.

Finally, the basic design concept also considers compressor operation in alternative running mode to deliver low pressure gas to the gas combustion unit (GCU). Operation with gas delivery simultaneously to both GCU and ME-GI is also possible.

Alternative fuel gas supply system concepts, employing either 2 x 50 % installed compressors and a separate supply line for the GCU, or 1 x 100 % compressor in combination with a BOG reliquefaction plant, are currently being considered by the market.

These alternative concepts are not de-scribed further in this paper.

The fuel gas compressor with the des-ignation 6LP250-5S_1 is designed to deliver low-temperature natural or forced boil-off gas from atmospheric tank pressure at an inlet temperature as low as −160°C, up to a gas injection pressure in the range of 150 to 265 bar. A total of fi ve compression stages are provided and arranged in a single verti-cal compressor casing directly driven by a conventional electric motor. The guiding principles of the compressor design are similar to those of API 618 for continuous operating process com-pression applications.

The compressor designation is as follows: 6LP250-5S_16 number of cranks L labyrinth sealing piston, stages 1 to 3P ring sealing piston, stages 4 to 5250 stroke in mm 5 number of stages S cylinder size reference 1 valve design

A unique compressor construction al-lows the selection of the best applicable cylinder sealing system according to the individual stage operating tempera-ture and pressure. In this way, a very high reliability and availability, with low maintenance, can be achieved.

Oil-free compression, required for the very cold low pressure stages 1 to 3, employs the labyrinth sealing principle, which is well proven over many years on LPG carriers and at LNG receiving terminals. The avoidance of mechanical friction in the contactless labyrinth cylin-der results in extremely long lifetimes of sealing components (see Appendix 1). The high-pressure stages 4 and 5 em-ploy a conventional API 618 lubricated cylinder ring sealed compressor tech-nology (see Fig. 3).

Fig. 2: Basic design concept for two compressor units 100 %, type 6LP250-5S_1

Fuel Gas Supply System – Key Components

Fuel gas compressor 6LP250-5S_1

The compression of cryogenic LNG boil-off gas up to discharge pressures in the range of 10-50 barg (1.0 to 5.0 MPa g) is now common practice in many LNG production and receiving terminals installed world wide today.

Compressor designs employing the highly reliable labyrinth sealing prin-ciple have been extensively used for such applications. The challenge for the compressor designer of the ME-GI application is to extend the delivery pressure reliably and effi ciently by add-ing additional compression stages to achieve the required engine injection pressure. In doing so, the compressor’s physical dimensions must consider the restricted space available within the deck-mounted machinery room.

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Fig. 3: Highly reliable cylinder sealing applied for each compression stage

Fig. 4: Main constructional features of the 6LP250-5S compressor

Six cylinders are mounted on top of a vertical arranged crankcase. The double acting labyrinth compression stages 1 to 3 are typical of those employed at an LNG receiving terminal.

The single acting stages 4 and 5 are a design commonly used for compres-sion of high-pressure hydrocarbon process gases in a refi nery application (Fig. 4).

The two fi rst-stage labyrinth cylinders, which are exposed to very low tem-peratures, are cast in the material GGGNi35 (Fig. 5). This is a nodular cast iron material containing 35 % nickel, also known under the trade name of Ni-Resist D5.

This alloy simultaneously exhibits re-markable ductility at low temperatures and one of the lowest thermal expan-sion coeffi cients known in metals.

The corresponding pistons are made of nickel alloyed cast iron with laminar graphite. Careful selection of cylinder materials allows the compressor to be

Ring piston – lubricated design

Stage 3 Stage 2 cooled cooled

Stage 4/5 Stage 4/5cooled cooledlube lube

Stage 1not cooled

Heat barrierstage 1 only

Ps= 1.03 bar a Ps= 1.03 bar

Pa= 265 bar a Pa= 265 bar a

Labyrinth piston – oil-free compression

Stage 1not cooled

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started at ambient temperature condi-tion and cooled down to BOG tempera-ture without any special procedures.

Second and third stage labyrinth cylin-ders operate over a higher temperature range and are therefore provided with a cooling jacket. Cylinder materials are nodular cast iron and grey cast iron re-spectively.

The oil lubricated high-pressure 4th and 5th stage cylinders are made from forged steel and are provided with a coolant jacket to remove heat of com-pression.

In view of the smaller compression volumes and high pressure, the piston and piston rod for stages 4 and 5 are integral and manufactured from a single forged steel material stock. Compres-sion is single acting with the 4th stage arranged at the upper end and the 5th stage at the lower end and arranged in step design. Piston rod gas leakage of the 5th stage is recovered to the suc-tion of the 4th stage (see Fig. 6).

Fig. 5: Cylinder block

Fig. 6: Sectional view of the lubricated cylinder 4th and 5th stage

Valve portsCylinder gas nozzie

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Motion work – 6LP250-5S

The 6-crank, 250 mm stoke compres-sor frame is a conventional low speed, crosshead design typically employed for continuous operating process du-ties. The industry design standard for this compressor type is the American Petroleum Industry Standard API 618 for refi nery process application.

The forged steel crankshaft and con-necting rods are supported by heavy tri-metal, force lubricated main bear-ings. Oil is supplied by a crankshaft driven main oil pump. A single distance piece arranged in the upper frame sec-tion provides separation between the lubricated motion work and the non-lubricated compressor cylinders. The passage of the crankshaft through the wall of the crankcase is sealed off by a rotating double-sided ring seal immersed in oil. Thus, the entire inside of the frame is integrated into the gas containing system with no gas leakage to the environment (see Fig. 7).

Capacity Control – Valve Unloading

Capacity control by valve unloading is extensively employed at LNG terminals where very large variations in BOG fl ows are experienced during LNG transfer from ship to storage tank. The capacity of the compressor may be simply and effi ciently reduced to 50 % in one step by the use of valve unloaders. The nitrogen actuated un-loaders (see Fig. 8) are installed on the lower cylinder suction valves and act to unload one half of the double-acting cylinders.

Additional stepless regulation, required to control a compressor capacity cor-responding to the rate of boil-off and the demand of the engine, is provided by returning gas from the discharge to compressor suction by the use of bypass valves. The compressor control system is described in detail later in this paper.

Fig. 8: Cylinder mounted suction valve unloader

Cylindergas nozzle

Valve seat

Valve disc

Compressed gas Suction gas

Diaphragm actuator

N2 control gas inlet/outlet

Fig. 7: Design principle of vertical gas-tight compressor casing

Piston rod guidingPiston rod guidance is provided at the lower crank end by a heavy nodular cast iron crosshead an at the upper end by an additional guide bearing . Both these components are oil lubricated and water cooled.

These key guiding elements are therefore subjected to very little wear.

Heat barrierThe cold fi rst-stage cylinders are separated from the warm compressor motion work by means of a special water jacket situated at the lower end of the cylinder block. This jaket is supplied with a water/glycol coolant mixture and acts as a thermal heat barrier.

Compression section, oil-free or lubricated

Distance piece provides separation

Piston guide system lubricated

Double-acting labyrinth or ring

Cylinder

Packing oil-free or lubricated

Heat barrier

Oil shield

Gulde bearing

Crosshead

Gas-tight casing

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Compressor System Engineering – 6LP250-5S

A compressor cannot function correctly and reliably without a well-designed and engineered external gas system. Static and dynamic mechanical analy-sis, thermal stress analysis, pulsation analysis of the compressor and auxiliary system consisting of gas piping, pulsa-tion vessels, gas intercoolers, etc., are standard parts of the compressor sup-plier’s responsibility.

A pulsation analysis considers upstream and downstream piping components in order to determine the correct sizing of pulsation dampening devices and their adequate supporting structure.

The compressor plant is designed to operate over a wide range of gas suc-tion temperatures from ambient start-up at +30°C down to −160°C without any special intervention.

Each compressor stage is provided with an intercooler to control the gas inlet temperature into the following stage. The intercooler design is of the conventional shell and tube type. The fi rst-stage intercooler is bypassed when the suction temperature falls below set limits (approx −80°C).

Volume LNG tanker cum 210,000

Max. BOG rate LNG tanker % 0.15 per day and liquid volume

Density of methane liquid at 1.06 bar a kg/m3 427 assumed basis for design

BOG mass fl ow kg/h 5,600

LNG tank pressure low / high bar a 1.06/1.20

Temperature BOG low °C −140 during loaded voyage

Temperature BOG high °C −40 during ballast voyage

Temperature BOG start up °C +30

Delivery P to ME-GI pressure low / high bar a 150/265

Temperature NG delivery to ME-GI °C +45

Compressor shaft power kW 1,600

Delivery P to GCU bar a 4.0 to 6.5

The P&I diagram for the compressor gas system is shown in Appendix III.

Bypass valves are provided over stage 1, stages 2 to 3, and stages 4 to 5. These valves function to regulate the fl ow of the compressor according to the engine set pressure within defi ned system limits. Non-return valves are provided on the suction, side to prevent gas back-fl ow to the storage tanks, between stages 3 and 4, to maintain adequate separation between the oil-free and the oil lubrication compres-sor stages, and at the fi nal discharge from the compressor.

Compressor safety

Safety relief valves are provided at the discharge of each compression stage to protect the cylinders and gas system against overpressure. Stage differen-tial relief valves, where applicable, are installed to prevent compressor exces-sive loading.

Pressure and temperature instrumenta-tion for each stage is provided to en-sure adequate system monitoring alarm and shutdown. Emergency procedures allow a safe shutdown, isolation and venting of the compressor gas system.

Table II: Rated process design data for a 210 kcum carrier

The design of the gas system com-prising piping, pulsation vessels, gas intercoolers, safety relief valves and ac-cessory components follows industry practices for hydrocarbon process oil and gas installations.

Process duty – compressor rating

The sizing of the fuel gas compressor is directly related to the “design” amount of nBOG and, therefore, to the capacity of the LNG carrier.

The fuel gas system design concept considers compressor operation not only for supplying gas to the ME-GI en-gine, but also to deliver gas to the gas combustion unit (GCU) in the event that the engine cannot accept any gas.

The compressors are therefore rated to handle the maximum amount of natural BOG defi ned by the tank system sup-plier and consistent with the design rat-ing of GCU.

Design nBOG rates are typically in the range of 0.135 to 0.15 % per day of tanker liquid capacity. During steady-state loaded voyage, a BOG rate of 0.10 to 0.12 % may be expected.

Carrier capacities in the range 145 to 260 kcum have been considered, re-sulting in the defi nition of 3 alternative compressor designs which differ according to frame rating and compressor speed.

Rated process design data for a carrier capacity of 210 kcum are as shown in Table II.

The rating for the electric motor driver is determined by the maximum com-pressor power required when consider-ing the full operating range of suction temperatures from + 30 to −140°C and suction pressures from 1.03 to 1.2 bar a.

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ME-GI Gas System Engineering

The ME-GI engine series, in terms of engine performance (output, speed, thermal effi ciency, exhaust gas amount and temperature, etc.) is identical to the well-established, type approved ME en-gine series. The application potential for the ME engine series therefore also ap-plies to the ME-GI engine, provided that gas is available as a main fuel. All ME en-gines can be offered as ME-GI engines. Since the ME system is well known, the following description of the ME-GI engine design only deals with new or modifi ed engine components.

Fig. 9 shows one cylinder unit of a S70ME-GI, with detail of the new modi-fi ed parts. These comprise gas supply double-wall piping, gas valve control block with internal accumulator on the (slightly modifi ed) cylinder cover, gas in-jection valves and ELGI valve for control of the injected gas amount. In addition, there are small modifi cations to the ex-haust gas receiver, and the control and manoeuvring system.

Apart from these systems on the en-gine, the engine and auxiliaries will comprise some new units. The most important ones, apart from the gas supply system, are listed below, and the full system is shown in schematic form in Appendix IV

The new units are:

Ventilation system, for venting the space between the inner and outer pipe of the double-wall piping.

Sealing oil system, delivering sealing oil to the gas valves separating the control oil and the gas.

Inert gas system, which enables purging of the gas system on the engine with inert gas.

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Fig. 9: Two-stroke MAN B&W S70ME-GI

The GI system also includes:

Control and safety system, compris-ing a hydrocarbon analyser for check-ing the hydrocarbon content of the air in the double-wall gas pipes.

The GI control and safety system is designed to “fail to safe condition”. All failures detected during gas fuel running includ-ing failures of the control system itself, will result in a gas fuel Stop/Shut Down, and a change-over to HFO fuel operation. Blow-out and gas-freeing purging of the high-pressure gas pipes and the complete gas supply system follows. The change-over to fuel oil mode is always done with-out any power loss on the engine.

The high-pressure gas from the com-pressor-unit fl ows through the main pipe via narrow and fl exible branch pipes to each cylinder’s gas valve block and accumulator. These branch pipes perform two important tasks:

They separate each cylinder unit fromthe rest in terms of gas dynamics, utili-sing the well-proven design philoso-phy of the ME engine’s fuel oil system.

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They act as fl exible connections be-tween the stiff main pipe system and the engine structure, safeguarding against extra-stresses in the main and branch pipes caused by the inevitable differences in thermal expansion of the gas pipe system and the engine structure.

The buffer tank, containing about 20 times the injection amount per stroke at MCR, also performs two important tasks:

It supplies the gas amount for injec-tion at a slight, but predetermined, pressure drop.

It forms an important part of the safety system.

Since the gas supply piping is of com-mon rail design, the gas injection valve must be controlled by an auxiliary control oil system. This, in principle, consists of the ME hydraulic control (system) oil sys-tem and an ELGI valve, supplying high-pressure control oil to the gas injection valve, thereby controlling the timing and opening of the gas valve.

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Gas valves

Large volumeaccumelator

Cylinder cover withgas valves and PMI

ELGI valve

High pressuredouble wallgas pipes

Exhaust reciever

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ME-GI Injection System

Dual fuel operation requires the injection of both pilot fuel and gas fuel into the combustion chamber.

Different types of valves are used for this purpose. Two are fi tted for gas injection and two for pilot fuel. The aux-iliary media required for both fuel and gas operation are as follows:

High-pressure gas supply

Fuel oil supply (pilot oil)

Control oil supply for activation of gas injection valves

Sealing oil supply.

The gas injection valve design is shown in Fig. 10. This valve complies with traditional design principles of com-pact design. Gas is admitted to the gas injection valve through bores in the cylinder cover. To prevent gas leakage between cylinder cover/gas injection valve and valve housing/spindle guide, sealing rings made of temperature and gas resistant material are installed. Any gas leakage through the gas sealing rings will be led through bores in the gas injection valve and further to space between the inner and the outer shield pipe of the double-wall gas piping sys-tem. This leakage will be detected by HC sensors.

The gas acts continuously on the valve spindle at a max. pressure of about 250 bar. To prevent gas from entering the control oil activating system via the clearance around the spindle, the spindle is sealed by sealing oil at a pressure higher than the gas pressure (25-50 bar higher).

The pilot oil valve is a standard ME fuel oil valve without any changes, except for the nozzle. The fuel oil pressure is constantly monitored by the GI safety

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system, in order to detect any malfunc-tioning of the valve.

The designs of oil valve will allow oper-ation solely on fuel oil up to MCR. lf the customer’s demand is for the gas engine to run at any time at 100 % load

on fuel oil, without stopping the engine, this can be done. If the demand is pro-longed operation on fuel oil, it is recom-mended to change the nozzles and gain an increase in effi ciency of around 1% when running at full engine load.

Cylinder cover

Gas inlet

Gas spindle

Sealing oil

Control oil

Connection to the

ventilated pipe system

Sealing oil inlet

Cylinder cover

Gas inlet

Gas spindle

Sealing oil

Control oil

Connection to the

ventilated pipe system

Sealing oil inlet

The system provides:Pressure, timing, rate shaping,main, pre- & post-injection

200 bar hydraulic oil.

Common withexhaust valve actuator

Low pressure fuel supply

Fuel return

Position sensor

Measuring andlimitingdevicePressure booster(800 - 900 bar)

Injection

ELFI valve

ELGI valve

Gas

800

600

400

200

00 5 10 15 20 30 3525 40 45

Bar abs

Pilotoil pressure

Control oil pressure

Deg. CA

Fig. 10: Gas injection valve – ME-GI engine

Fig. 11: ME-GI system

Cylinder cover

Gas inlet

Sealing oil inlet

Connection to the ventilated pipe systemControl oilSealing oil

gas spindle

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As can be seen in Fig. 11 (GI injection system), the ME-GI injection system consists of two fuel oil valves, two fuel gas valves, ELGI for opening and clos-ing of the fuel gas valves, and a FIVA valve to control (via the fuel oil valve) the injected fuel oil profi le. Furthermore, it consists of the conventional fuel oil pressure booster, which supplies pilot oil in the dual fuel operation mode. This fuel oil pressure booster is equipped with a pressure sensor to measure the pilot oil on the high pressure side. As mentioned earlier, this sensor monitors the functioning of the fuel oil valve. If any deviation from a normal injection is found, the GI safety system will not allow opening for the control oil via the ELGI valve. In this event no gas injec-tion will take place.

Under normal operation where no mal-functioning of the fuel oil valve is found, the fuel gas valve is opened at the cor-rect crank angle position, and gas is injected. The gas is supplied directly into an ongoing combustion. Conse-quently the chance of having unburnt gas eventually slipping past the piston rings and into the scavenge air receiver is considered to be very low. Monitoring the scavenge air receiver pressure safe-guards against such a situation. In the event of high pressure, the gas mode is stopped and the engine returns to burning fuel oil only.

The gas fl ow to each cylinder during one cycle will be detected by measur-ing the pressure drop in the accumu-lator. By this system, any abnormal gas fl ow, whether due to seized gas injection valves or blocked gas valves, will be detected immediately. The gas supply will be discontinued and the gas lines purged with inert gas. Also in this event, the engine will continue running on fuel oil only without any power loss.

High-Pressure Double-Wall Piping

A common rail (constant pressure) gas supply system is to be fi tted for high-pressure gas distribution to each valve block. Gas pipes are designed with double-walls, with the outer shielding pipe designed so as to prevent gas outfl ow to the machinery spaces in the event of rupture of the inner gas pipe. The intervening space, including also the space around valves, fl anges, etc., is equipped with separate mech-anical ventilation with a capacity of approx. 30 air changes per hour. The pressure in the intervening space is below that of the engine room with the (extractor) fan motors placed outside the ventilation ducts. The ventilation inlet air is taken from a non-hazardous area.

Gas pipes are arranged in such a way, see Fig. 12 and Fig 13, that air is suck- ed into the double-wall piping system from around the pipe inlet, from there into the branch pipes to the individual gas valve control blocks, via the branch

supply pipes to the main supply pipe, and via the suction blower into the at-mosphere.

Ventilation air is exhausted to a fi re-safe place. The double-wall piping system is designed so that every part is ven-tilated. All joints connected with seal-ings to a high-pressure gas volume are being ventilated. Any gas leakage will therefore be led to the ventilated part of the double-wall piping system and be detected by the HC sensors.

The gas pipes on the engine are de-signed for 50% higher pressure than the normal working pressure, and are supported so as to avoid mechanical vibrations. The gas pipes are further-more shielded against heavy items fall-ing down, and on the engine side they are placed below the top-gallery. The pipes are pressure tested at 1.5 times the working pressure. The design is to be all-welded, as far as it is practicable, using fl ange connections only to the ex-tent necessary for servicing purposes.

Ventilation air

Ventilation air

Fuel Gas flow

Fig. 12: Branching of gas piping system

Protective hose

High pressure gas pipe

High pressure gasVentilation air

Outer pipe

Soldered

Ventilation air

Bonded sealVentilation air

Fuel Gas fl ow

13

5510-0026-00ppr.indd 13 2007-10-01 13:51:04

One way valve Fuel gas inlet

Purge

valves

Fuel gas

Nitrogen

Ventilation air

Control air

Control Oil

Gas stop

valve

Cylinder

cover

Fuel gas

accumulator

volume

Control oil buffer

volume

Fig. 13: Gas valve control block

The branch piping to the individual cylinders is designed with adequate fl exibility to cope with the thermal ex-pansion of the engine from cold to hot condition. The gas pipe system is also designed so as to avoid excessive gas pressure fl uctuations during operation. For the purpose of purging the system after gas use, the gas pipes are con-nected to an inert gas system with an inert gas pressure of 4-8 bar. In the event of a gas failure, the high-pressure pipe system is depressurised before automatic purging. During a normal gas stop, the automatic purging will be started after a period of 30 min. Time is therefore available for a quick re-start in gas mode.

14

5510-0026-00ppr.indd 14 2007-10-01 13:51:09

Fuel Gas System - Control Requirements

The primary function of the compres-sor control system is to ensure that the required discharge pressure is always available to match the demand of the main propulsion diesel engines. In do-ing so, the control system must ade-quately handle the gas supply variables such as tank pressure, BOG rate (laden and ballast voyage), gas composition and gas suction temperature.

If the amount of nBOG decreases, the compressor must be operated on part load to ensure a stable tank pressure, or forced boil-off gas (fBOG) added to the gas supply. If the amount of nBOG increases, resulting in a higher than acceptable tank pressure, the control system must act to send excess gas to the gas combustion unit (GCU).

Tank pressure changes take place over a relatively long period of time due to the large storage volumes involved. A fast reaction time of the control sys-tem is therefore not required for this control variable.

The main control variable for compres-sor operation is the feed pressure to the ME-GI engine, which may be subject to controlled or instantaneous change. An adequate control system must be able to handle such events as part of the “normal” operating procedure.

The required gas delivery pressure var-ies between 150-265 bar, depending on the engine load (see Fig. 14 below).

The compressor must also be able to operate continuously in full recycle mode with 100 % of delivered gas returned to the suction side of the compressor. In addition, simultaneous delivery of gas to the ME-GI engine and GCU must be possible.

When considering compressor control, an important difference between cen-trifugal and reciprocating compressors should be understood. A reciprocating compressor will always deliver the pres-sure demanded by the down-stream user, independent of any suction con-ditions such as temperature, pressure, gas composition, etc. Centrifugal com-pressors are designed to deliver a cer-tain head of gas for a given fl ow. The discharge pressure of these compres-sors will therefore vary according to the gas suction condition. This aspect is very important when considering transient starting conditions such as suction temperature and pres-sure. The 6LP250-5S_1 reciprocating compressor has a simple and fast start-up procedure.

Compressor control – 6LP250-5S_1

Overall control concept

Fig. 15 shows a simplifi ed view of the compressor process fl ow sheet. The system may be effectively divided into a low-pressure section (LP) consisting of the cold compression stage 1, and a high-pressure section (HP) consisting of stages 2 to 5.

Gas

pre

ssure

Set

poin

t(b

ar)

Control of gas delivery pressure

Engine Load (% of MCR )

General Data forGas Delivery Condition:

Pressure:Nominal 250 barMax. value 300 barPulsation limit ± 2 barSet point tolerance ± 5%

Temperature :Approx. 45 oC

Quality:Condensate free, without oil/waterdroplets or mist, similar to thePNEUROP recommendation 6611‘‘Air Turbines ’’

Fig. 14: Gas supply station, guiding specifi cation

The main control input for compressor control is the feed pressure Pset re-quired by the ME-GI engine. The feed pressure may be set in the range of 150 to 265 bar according to the desired en-gine load. If the two ME-GI engines are operating at different loads, the higher set pressure is valid for the compressor control unit.

If the amount of nBOG is insuffi cient to satisfy the engine load requirement, and make-up with fBOG is not foreseen, the compressor will operate on part load to ensure that the tank pressure remains within specifi ed limits. The ME-GI en-gine will act independently to increase the supply of HFO to the engine. Prima-ry regulation of the compressor capac-ity is made with the 1st stage bypass valve, followed by cylinder valve unload-ing and if required bypass over stages 2 to 5. With this sequence, the compres-sor is able to operate fl exibly over the full capacity range from 100 to 0 %.

If the amount of nBOG is higher than can be burnt in the engine (for example during early part of the laden voyage) resulting in higher than acceptable suction pressure (tank pressure), the control system will send excess gas to the GCU via the side stream of the 1st compression stage.

15

5510-0026-00ppr.indd 15 2007-10-01 13:51:10

In the event of engine shutdown or sud-den change in engine load, the com-pressor delivery line must be protected against overpressure by opening by-pass valves over the HP section of the compressor.

During start-up of the compressor with warm nBOG, the temperature con-trol valves will operate to direct a fl ow through an additional gas intercooler after the 1st compression stage.

The control concept for the compres-sor is based on one main control mode which is called “power saving mode”. This mode of running, which minimises the use of gas bypass as the primary method of regulation, operates within various well defi ned control limits.

The system pressure control limits are as follows:

Fig. 15: Simplifi ed fl ow sheet

Pmin suction Prevents under-pressure in compressor inlet manifold - tank vacuum.

Phigh suction Suction manifold high- pressure - system safety (GCU) on standby.

Pmax suction Initiates action to reduce inlet manifold pressure. Pmax Prevents overpressure of ME-GI feed compressor discharge manifold. A detailed description of operation with-in these control limits is given below.

Power saving mode

Economical regulation of a multi-stage compressor is most effi ciently executed using gas recycle around the 1st stage of compression. The ME-GI required set pressure Pset is therefore taken as control input directly to the compressor

1st stage bypass valve, which will open or close until the actual compressor discharge pressure is equal to the Pset. With this method of control, BOG de-livery to the ME-GI is regulated without any direct measurement and control of the delivered mass fl ow. If none of the above control limits are active, the con-troller is able to regulate the mass fl ow in the range from 0 to 100 %.

The following control limits act to over-rule the ME-GI controller setting and initiate bypass valve operation:

Pmin suction (tank pressure below set level)

The control scenario is falling suction pressure. If the Pmin limit is active, the 1st stage recycle valve will not be permitted to close further, thereby preventing fur-ther reduction in suction pressure. If the pressure in the suction line continues to decrease, the recycle valve will open governed by the Pmin limiter.

16

5510-0026-00ppr.indd 16 2007-10-01 13:51:10

burned simultaneously in the GCU. No action is taken in the ME-GI control system.

Pmax ME-GI feed

The control scenario is a reduction of the engine load or closure of the ME-GI supply line downstream of the com-pressor. The pressure will rise in the delivery line. Line overpressure is pre-vented by a limiter, which acts to direct-ly open the bypass control valve around stages 2 to 5. As a consequence, the controller will also open the 1st stage recycle valve.

The control range of the compressor is 0 to 100 % mass fl ow.

GCU-only operating mode

The control scenario considers a situa-tion where gas injection to the ME-GI is not required and tank gas pressure is at the level of Phigh.

The nBOG is compressed and delivered to the GCU by means of a gas take-off after the 1st stage.

The following actions are initiated:

• manual start of the GCU

• closing of the bypass valve around 1st stage

• fully opening of the bypass valves around stages 2-5.

In this mode, the compressor is oper-ating with stages 2-5 in full recycle at a reduced discharge pressure of approxi-mately 80 bar. The pressure setting of the GCU feed valve is set directly by the GCU in the range 3 to 6 bar a.

There is no action on the ME-GI controller.

Action of Pressure will fall at theME-GI control compressor dischargesystem: requiring the HFO injection rate to be increased.

fBOG: If a spray cooling or forced vaporizer is installed, it may be used for stabilising the suction pressure and thereby increase the gas mass fl ow to the engine. Such a system could be activated by the Pmin suction pressure limit.

Phigh suction (tank pressure above set level)

The control scenario is increasing suc-tion pressure due to either reduced engine load (e.g. approaching port, manoeuvring) or excess nBOG due to liquid impurities (e.g. N2).

The control limiter initiates a manual start of the GCU (the GCU is assumed not to be on standby mode during nor-mal voyage).

There is no action on the compressor control or the ME-GI control system.

Pmax suction (tank pressure too high)

The control scenario is the same as de-scribed above, however, it has resulted in even higher suction pressure. Action must now be taken to reduce suction pressure by sending gas to the GCU.

The high pressure alarm initiates a manual sequence whereby the 1st stage bypass valve PCV01 is closed and the bypass valve PCV02 to the GCU is opened. When the changeover is completed, automatic Pset control is transferred to the GCU control valve PCV02. The gas amount which can-not be accepted by the ME-GI will be

17

5510-0026-00ppr.indd 17 2007-10-01 13:51:11

Fig. 16: Typical layout of cargo machinery room

Machinery Room Instal-lation – 6LP250-5S

The layout of the cargo handling equip-ment and the design of their supporting structure presents quite a challenge to the shipbuilder where space on deck is always at a premium. In conjunction with HHI and the compressor maker, an optimised layout of the fuel gas com-pressor has been developed.

There are many factors which infl uence the compressor plant layout apart from limited space availability. (See Fig. 16.) External piping connections, adequate access for operation and maintenance, equipment design and manufacturing codes, plant lifting and installation are just a few. The compressor together with acces-sory items comprising motor drive, auxiliary oil system, vessels, gas cool-ers, interconnecting piping, etc., are manufactured as modules requiring minimum assembly work on the ship deck. Separate auxiliary systems pro-vide coolant for the compressor frame and gas coolers.

If required, a dividing bulkhead may separate the main motor drive from the hazardous area in the compressor room. A compact driveshaft arrange-ment without bulkhead, using a suitably designed ex motor, is however pre-ferred. Platforms and stairways provide access to the compressor cylinders for valve maintenance. Piston assemblies are withdrawn vertically through man-holes in the roof of the machinery house (see Fig. 17).

Fig. 17: Fuel gas compressor with accessories

Dischargeline

Suction line

Compensator E. mortor

Oil System

15m

27m

34m

18

5510-0026-00ppr.indd 18 2007-10-01 13:51:11

Requirements for Cargo Machinery Room Support Structure

Fig. 18 shows details of the compressor base frame footprint and requirement for support by the ship structure.

Reciprocating compressors, by nature of movement of their rotating parts, exhibit out-of-balance forces and mo-ments which must be considered in the design of the supporting structure for acceptable machinery vibration levels.

As a boundary condition, the structure underneath the cargo machinery room must have adequate weight and stiff-ness to provide a topside vibration level of (approximately) 1.2 - 1.5 mm/s. Sat-isfactory vibration levels for compressor frame and cylinders are 8 and 15 mm/s respectively (values given are rms – root mean square).

Fig. 18: Compressor base frame footprint

Foundation defl ection due to ship movement must, furthermore, be con-sidered in the design of the compressor plant to ensure stress-free piping termi-nations.

Maintenance requirements - availability/reliability

The low speed, crosshead type com-pressor design 6LP250-5S, like the ME-GI diesel engine, is designed for the life time of the LNG carrier (25 to 30 years or longer). Routine maintenance is limited purely to periodic checking in the machinery room.

Maintenance intervention for dis-mantling, checking and eventual part replacement is recommended after each 8,000 hours of operation. Annual maintenance interventions will normally require 50-70 hours work for checking and possible replacing of wearing parts.

Major intervention for dismantling and bearing inspection is recommended every 2-3 years.

Average availability per compressor unit is estimated to be 98.5 % with best availability approximately 99.5 %. With an installed redundant unit, the com-pressor plant availability will be in the region of 99.25 %.

Any unscheduled stoppage of the 6LP250-5S compressor will most likely be attributable to a mal-function of a cylinder valve. With the correct valve design and material selection (Burck-hardt uses its own design and manu-facture plate valves) these events will be very seldom, however a valve failure in operation cannot be entirely ruled out.

LNG boil-off gas is an ideal gas to com-press. The gas is relatively pure and uncontaminated, the gas components are well defi ned, and the operating tem-peratures are stable once “cool-down” is completed.

These conditions are excellent for long lifetime of the compressor valves where an average lifetime expectancy for valve plates is 16,000 hours. Therefore, we do not expect any unscheduled inter-vention per year for valve maintenance. Such a maintenance intervention will take approx. 7-9 hours for compressor shutdown, isolation and valve replace-ment.

A total unscheduled maintenance in-tervention time of 25 hours, assuming 8,000 operating hours per year, may be used for statistical comparison. On this basis compressor reliability is estimated at 99.7 %.

Our experience in many installations shows that no hours are lost for un-scheduled maintenance. The reliability of these compressors is therefore com-parable to that of centrifugal compres-sor types.

19

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Requirements for Classifi cation

When entering the LNG market with the combined two-stroke and reliquefaction solution, it was discovered that there is a big difference in the requirements from operators and classifi cation socie-ties.

Being used to cooperating with the classifi cation societies on other com-mercial ships, the rules and design re-commendations for the various applica-tions in the LNG market are new when it comes to diesel engine propulsion. In regard to safety, the high availability and reliability offered when using the two-stroke engines generally fulfi l the requirements, but as the delivery and pick up of gas in the terminals is carried out within a very narrow time window, redundancy is therefore essential to the operators.

As such, a two-engine ME-GI solution is the new choice, with its high effi -ciency, availability and reliability, as the traditional HFO burning engines.

Compared with traditional diesel operated ships, the operators and shipowners in the LNG industry generally have different goals and demands to their LNG tankers, and they often apply more strict design criteria than applied so far by the classifi cation societies.

A Hazid investigation was therefore found to be the only way to secure that all situations are taken into account when using gas for propulsion, and that all necessary precautions have been taken to minimize any risk involved.

In 2005, HHI shipyard, HHI engine builder, BCA and MAN Diesel therefore worked out a hazard identifi cation study that was conducted by Det Norske Ver-itas (DNV), see Appendix V.

Actual Test and Analysis of Safety when Operating on Gas

The use of gas on a diesel engine calls for careful attention with regard to safety. For this reason, ventilated dou-ble-walled piping is a minimum require-ment to the transportation of gas to the engine.

In addition to hazard considerations and calculations, it has been necessary to carry out tests, two of which were carried out some years ago before the installation and operation of the Chiba power plant 12K80MC-GI engine in 1994.

A crack in the double-wall inner pipe

The fi rst test was performed by intro-ducing a crack in the inner pipe to see if the outer pipe would stay intact. The test showed no penetration of the outer pipe, thus it could be concluded that the double-wall concept lived up to the expectations.

Pressure fl uctuation

The second test was carried out to investigate the pressure fl uctuations in the relatively long piping from the gas compressor to the engine.

By estimation of the necessary buffer volume in the piping system, the stroke and injection of gas was calculated to see when safe pressure fl uctuations are achieved within given limits for optimal performance of the engines. The piping system has been designed on the basis of these calculations.

Main Engine Room Safety

The latest investigation, which was recently fi nished, was initiated by a number of players in the LNG market questioning the use of 250 bar gas in the engine room, which is also located under the wheel house where the crew is working and living.

Even though the risk of full breakage happening is considered close to neg-ligible and, in spite of the precautions introduced in the system design, MAN Diesel found it necessary to investi-gate the effect of such an accident, as the question still remains in part of the industry: what if a double-wall pipe breaks in two and gas is released from a full opening and is ignited?

As specialists in the offshore industry, DNV were commissioned to simulate such a worst case situation, study the consequences, and point to the appro-priate countermeasures. DNV’s work comprised a CFD (computational fl uid dynamics) simulation of the hazard of an explosion and subsequent fi re, and an investigation of the risk of this situa-tion ever occurring and at what scale.

As input for the simulation, the volume of the engine room space, the position of major components, the air ventilation rate, and the location of the gas pipe and control room were the key input parameters.

Realistic gas leakage scenarios were defi ned, assuming a full breakage of the outer pipe and a large or small hole in the inner fuel pipe. Actions from the closure of the gas shutdown valves, the ventilation system and the ventilation conditions prior to and after detec-tion are included in the analysis. The amount of gas in the fuel pipe limits the duration of the leak. Ignition of a leak causing an explosion or a fi re is further-more factored in, due to possible hot

20

5510-0026-00ppr.indd 20 2007-10-01 13:51:12

spots or electrical equipment that can give sparks in the engine room.

Calculations of the leak rate as a func-tion of time, and the ventilation fl ow rates were performed and applied as input to the explosion and fi re analyses.

Simulation Results

The probability of this hazard happen-ing is based on experience from the offshore industry.

Even calculated in the worst case, no structural damage will occur in the HHI LNG engine room if designed for 1.1 bar over pressure.

No areas outside the engine room will be affected by an explosion.

If this situation is considered to repre-sent too high a risk, unattended machin-ery space during gas operation can be introduced. Today, most engines and equipment are already approved by the classifi cation societies for this type of operation.

By insulation, the switchboard room fl oor can be protected against heat from any jet fi res.

No failure of the fuel oil tank structure, consequently no escalation of fi re.

The above conclusion is made on the assumption that the GI safety system is fully working.

In addition, DNV has arrived at a differ-ent result based on the assumption that the safety system is not working. On the basic in these results DNV have put up failure frequencies and developed a set of requirements to be followed in case a higher level of safety is required.

After these conclusion made by DNV, HHI has developed a level for their engine room safety that satisfi es the requirements from the classifi cation so-cieties, and also the requirements that are expected from the shipowners.

•

•

•

This new engine room design is based on the experience achieved by HHI with their fi rst orders for LNG carriers equipped with 2 x 6S70ME-C and reli-quefaction plant. The extra safety that will be included is listed below:

Double-wall piping is located as far away as possible from critical walls such as the fuel tank walls and switchboard room walls.

In case of an engine room fi re alarm, a gas shutdown signal is sent out, the engine room ventilation fans stops, and the air inlet canals are blocked.

During gas running it is not possible to perform any heavy lifting with the engine room crane.

A failure of the engine room ventilation will result in a gas shutdown.

HC sensors are placed in the engine room, and their position will be based on a dispersion analysis made for the purpose of fi nding the best location for the sensors.

The double-wall piping is designed with lyres, so that variation in tem-peratures from pipes to surroundings can be absorbed in the piping.

In fact, any level of safety can be achieved on request of the shipowner. The safety level request will be achieved in a co-operation between the yard HHI, the engine builder HHI, the classifi -cation society and MAN Diesel A/S.

The report “Dual fuel Concept: Analysis of fi res and explosions in engine room” was made by DNV consulting and can be ordered by contacting MAN Diesel A/S, in Copenhagen.

•

•

•

•

•

•

21

5510-0026-00ppr.indd 21 2007-10-01 13:51:13

Fuel oilFuel gas

Fuel oil

100% load

Fuel100%

100% load

Fuel100%Fuel-oil-only mode "Minimum fuel" mode

Min. Pilotoil 5-8%

30% load

Min load for Min. fuel mode

Fig. 19: Fuel type modes for the ME-GI engines for LNG carriers

Engine Operating Modes

One of the advantages of the ME-GI engine is its fuel fl exibility, from which an LNG carrier can certainly benefi t. Burning the boil-off gas with a varia-tion in the heat value is perfect for the diesel working principle. At the start of a laden voyage, the natural boil-off gas holds a large amount of nitrogen and the heat value is low. If the boil-off gas is being forced, it can consist of both ethane and propane, and the heat value could be high. A two-stroke, high-pres-sure gas injection engine is able to burn those different fuels and also without a drop in the thermal effi ciency of the engine. The control concept comprises two different fuel modes, see alsoFig. 19.

fuel-oil-only mode

minimum-fuel mode

The fuel-oil-only mode is well known from the ME engine. Operating the engine in this mode can only be done on fuel oil. In this mode, the engine is considered “gas safe”. If a failure in the gas system occurs it will result in a gas shutdown and a return to the fuel-oil-only mode and the engine is “gas safe”.

The minimum-fuel mode is developed for gas operation, and it can only be started manually by an operator on the

•

•

Gas Main Operating Panel in the control room. In this mode, the control system will allow any ratio between fuel oil and gas fuel, with a minimum preset amount of fuel oil to be used.

The preset minimum amount of fuel oil, hereafter named pilot oil, to be used is in between 5-8% depending on the fuel oil quality. Both heavy fuel oil and marine diesel oil can be used as pilot oil. The min. pilot oil percentage is cal-culated from 100% engine load, and is constant in the load range from 30-100%. Below 30% load MAN Diesel is not able to guarantee a stable gas and pilot oil combustion, when the engine reach this lower limit the engine returns to Fuel-oil-only mode.

Gas fuels correspond to low-sulphur fuels, and for this type of fuel we rec-ommend the cylinder lube oil TBN40 to be used. Very good cylinder condition with this lube oil was achieved from the gas engine on the Chiba power plant. A heavy fuel oil with a high sulphur con-tent requires the cylinder lube oil TBN 70. Shipowners intending to run their engine on high-sulphur fuels for longer periods of time are recommended to in-stall two lube oil tanks. When changing to minimum-fuel mode, the change of lube oil should be carried out as well.

Players in the market have been fo-cussed on reducing the exhaust emis-sions during harbour manoeuvring. When testing the ME-GI at the MAN Diesel research centre in Copenhagen,

the 30% limit for minimum-fuel mode will be challenged taking advantage of the increased possibilities of the ME fuel valves system to change its injection profi le, MAN Diesel expects to lower this 30% load limit for gas use, but for now no guaranties can be given.

22

5510-0026-00ppr.indd 22 2007-10-01 13:51:13

Launching the ME-GI

As a licensor, MAN Diesel expects a time frame of two years from order to delivery of the fi rst ME-GI on the test-bed.

In the course of this time, depending on the ME-GI engine size chosen, the engine builder will make the detailed designs and a fi nal commissioning test on a research engine. This type approv-al test (TAT) is to be presented to the classifi cation society and ship owner in question to show that the compressor and the ME-GI engine is working in all the operation modes and conditions.

In cooperation with the classifi ca-tion society and engine builder, the most optimum solution, i.e. to test the compressor and ME-GI engine before delivery to the operator has been con-sidered and discussed. One solution is

to test the gas engine on the testbed, but this is a costly method. Alternatively, and recommended by MAN Diesel, the compressor and ME-GI operation test could be made in continuation of the gas trial. Today, there are different opin-ions among the classifi cation societies, and both solutions are possible de-pending on the choice of classifi cation society and arrangement between ship owners, yard and engine builder.

MAN Diesel A/S has developed a test philosophy especially for approval of the ME-GI application to LNG carriers, this philosophy has so far been approved by DNV, GL, LR and ABS, see Table III. The idea is that the FAT (Factory Ac-ceptance Test) is being performed for the ME system like normal, and for the GI system it is performed on board the LNG carrier as a part of the Gas Trial Test. Thereby, the GI system is tested in combination with the tailor-made

gas compressor system for the specifi c LNG carrier. Only in this combination it will be possible to get a valid test.

Prior to the gas trial test, the GI system has been tested to ensure that every-thing is working satisfactory.

MAN Diesel Copenhagen

Engine builder testbed

Yard Quay trial

Yard Sea trial

Gas trial

MAN B&W research engine – 4T50MX or similar suitable location

TAT of ME-GI con-trol system and of gas components. Test according to MBD test program. Subject to Class ap-proval.

First ME-GI production engine

Test according to: • IACS UR M51 MBD Factory Acceptance Test program (FAT) for ME engines.

Test according to: • Yard and Engine Builder test program approved by Class

Test according to: • Yard and Engine Builder test program approved by Class

After loading gas, the following tests are to be carried out: • Acceptance test of the complete gas system including the main engine. • Test of the ME-GI control system according to MBD test program approved by Class

Second and follow-ing ME-GI engines

- do - - do - - do - - do -

Engine is tested on: Gas and marine diesel oil

Marine diesel oil Marine diesel oil and/or heavy fuel oil

Marine diesel oil and/or heavy fuel oil

Marine diesel Heavy fuel oil and gas

Table III: MAN B&W ME-GI engines – test and class approval philosophy

23

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Fig. 20: Alternative two-stroke propulsion and power generation machinery systems

Machinery Concepts Comparison

In this chapter, the ME-C and ME-GI engines in the various confi guations will be compared. The comparison will show the most suitable propulsion solu-tion for a modern LNG carrier.

The study is made as objective as pos-sible, however, only MAN Diesel sup-ported systems are compared.

Both the ME-C engine with reliquefac-tion and the ME-GI engine with gas compressor can be used either in twin engine arrangements, coupled to two fi xed pitch (FPP) or two controllable pitch propellers (CPP), or as a single main engine coupled to one FPP.

For LNG carriers, the total electricityconsumption of the machinery on board is higher than usual compared with most other merchant ship types.

Therefore, the electrical power genera-tion is included in the comparison.

Thus, the various main propulsion ma-chinery solutions may be coupled with various electricity producers, such as diesel generators (DG), the MAN Diesel waste heat recovery system, called the Thermo Effi ciency System (TES), or a shaft generator system (PTO).

Applying the propulsion data listed in Tables IV and V, the estimated data for the electrical power consumption in Tables VI and VII, MAN Diesel has cal-culated the investment and operational costs of all the alternative confi gura-tions illustrated in Fig. 20.

The investment and operational costs have been analysed and the results have been compared using the Net Present Value (NPV) method, see Fig. 21.

In order to quantify the effect of the machinery chosen on the total exhaust gas emissions, and thereby bring it directly into the comparison, costs for the various emission pollutants have been assumed and used in some of the calculations, thereby visualising a possible future economic impact of the emissions.

The following emission fees have been used in the calculations:

CO2: 17.3 USD/tonneNOx: 2,000 USD/tonneSO2: 2,000 USD/tonne

It has been assumed that the CO2 fee is to be paid for the complete CO2 emis-sion, whereas the NOx and SO2 fees are to be paid for only 20% of the total NOx and SO2 emissions, since the two latter pollutants are mostly a problem when the ship operates close to the coast line.

Calculations have been made, taking different HFO and LNG prices and dif-ferent time horizons (10, 20 and 30 years) into account, and with and with-out the incorporation of the estimated emission fees.

The calculations have been made for three different sizes of LNG carriers; 150,000, 210,000 and 250,000 m3.

Finally, the Net Present Value results, for each LNG carrier size, have been scaled towards each other in such a way that the highest Net Present Value, which represents the alternative with the highest cost for each combination of fuel prices, time horizon and emis-sion scenario has been nominated to equal 100% cost, whereas the remain-ing Net Present Values within the same category have been listed in percent-ages of the above most expensive confi guration.

ME + DG

ME + TES + DG

ME + PTO + DG

ME + TES + PTO + DG

1 x FPP

2 x FPP

2 x CPP

HFO + reliq.

ME-C

Dual Fuel

ME-GI

LNG

Carrier

NPV formula Each cash infl ow/outfl ow is discounted back to its Present Value. Then they are summed. Therefore:

Wheret - the time of the cash fl ow

n - the total time of the project

r - the discount rate

Ct - the net cash fl ow (the amount of cash) at that point in time.

C0 - the capitial outlay at the begining of the investment time ( t = 0 )

Fig. 21: NPV defi nition

24

5510-0026-00ppr.indd 24 2007-10-01 13:51:13

Table V: Average ship particulars used for propulsion power prediction calculations for LNG carriers of the membrane type

Case Unit A B C

Ship capacity m3 150,000 210,000 250,000

Scantling deadweight

dwt 80,000 108,000 129,000

Scantling draught

m 12.3 12.7 12.7

Average designship speed

knot 20.0 20.0 20.0

Design deadweight

dwt 74,000 98,500 118,000

Light weight of ship

t 30,000 40,000 48,000

Design displace-ment of ship

t 104,000 138,500 166,000

Design draught m 11.6 12.0 12.0

Length overall m 288 315 345

Length between perpendiculars

m 275 303 332

Breadth m 44.2 50.0 54.0

Breadth/design draught ratio

3.81 4.17 4.50

Block coeffi cient, perpendicular

0.720 0.743 0.753

Sea margin % 15 15 15

Engine margin % 10 10 10

Light running margin

% 5 5 5

Case Unit A B C

Ship capacity m3 150,000 210,000 250,000

Design draught m 11.6 12.0 12.0

1. Single propeller

Propeller diameter

m 1 x 8.60 1 x 8.80 1 x 9.00

SMCR power kW 1 x 31,361 1 x 39,268 1 x 45,152

SMCR speed rpm 92.8 91.8 93.8

Main engine (without PTO)

1 x 7K90ME Mk 6

1 x 7K98ME Mk 7

1 x 8K98ME Mk 6

2. Twin-skeg and Twin-propulsion.

Propeller diameter

m 2 x 8.10 2 x 8.40 2 x 8.70

SMCR power kW 2 x 14,898 2 x 18,301 2 x 20,780

SMCR speed rpm 88.1 90.5 88.0

Main engine(without PTO)

2 x 5S70ME-CMk 7

2 x 6S70ME-CMk 7

2 x 7S70ME-CMk 7

Ballast draught m 9.7 9.9 10.3

Average engine load in ballast

% SMCR

68 68 68

Table IV: Results of propulsion power prediction calculations for LNG carriers of the membrane type

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5510-0026-00ppr.indd 25 2007-10-01 13:51:13

150,000 m3 Reliquefac-

tion

Load scenario

Relique-faction (kW)

Other consumers

(kW)

Total electricity consumption

(kW)

Laden voyage

3370 2100 5470

Ballast voyage

800 2100 3065

Loading at terminal

800 4500 5300

Unloading at terminal

800 6400 7200

Manoeuvring laden

3370 3200 6570

Manoeuvring ballast

965 3200 4165


tion

Load scenario


Other con-sumers (kW)


(kW)

Laden voyage

4565 2150 6715

Ballast voy-age

1365 2150 3515

Loading at terminal

1000 4500 5500


1000 7000 8000

Manoeuvring laden

4565 3400 7965

Manoeuvring ballast

1365 3400 4765


tion

Load scenario


Other con-sumers (kW)


(kW)

Laden voyage

5595 2200 7795

Ballast voyage

1595 2200 3795

Loading at terminal

1240 4500 5740


1240 7400 8640

Manoeuvring laden

5595 3600 9195

Manoeuvring ballast

1595 3600 5195

150,000 m3 Dual fuel

Load scenario

Gas com-pressor

(kW)

Other consumers

(kW)


(kW)

Laden voyage

1630 2100 3730

Ballast voyage

1630 2100 3730

Loading at terminal

(0) 4500 4500


(0) 6400 6400

Manoeuvring laden

(0) 3200 3200

Manoeuvring ballast

(0) 3200 3200


Load scenario

Gas com-pressor

(kW)

Other consumers

(kW)


(kW)

Laden voyage

1630 2150 3780

Ballast voy-age

1630 2150 3780

Loading at terminal

(0) 4500 4500


(0) 7000 7000

Manoeuvring laden

(0) 3400 3400

Manoeuvring ballast

(0) 3400 3400


Load scenario

Gas com-pressor

(kW)

Other consumers

(kW)


(kW)

Laden voyage

1630 2200 3830

Ballast voyage

1630 2200 3830

Loading at terminal

(0) 4500 4500


(0) 7400 7400

Manoeuvring laden

(0) 3600 3600

Manoeuvring ballast

(0) 3600 3600

Table VI: Electrical power consumption for reliquefaction Table VII: Electrical power consumption for ME-GI

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An example of the results is illustrated below in Table VIII. Fuel Price 1 corre-sponds to the price level of 2006. The following colour codes apply to Table VII.The analysis shows that, economically, good predictions of the future develop-ment of the HFO and LNG prices relative to each other are absolutely essential for choosing the optimal main propulsion two-stroke engines for the vessel, i.e. the choice whether to use HFO burning main engines or gas burning main en-gines is the single most important deci-sion to make.

However, the result in regard to the fuel price relationship may also be infl uenced by some factors in the business model used, e.g. whether a fi xed amount of LNG is to be shipped by the vessel or a fi xed amount of LNG is to be delivered by the vessel.

After that, the total economy of the LNG carrier purchase and operation can also be infl uenced by choosing the most op-timal machinery confi guration. This goes for the main propulsion plant, but also includes the electricity production plant for the vessel in question which, how-ever, is of smaller signifi cance to the total economy considerations than the main engine fuel type. Single-propeller machinery arrange-ments do not seem to be attractive be-cause of the lower propulsion effi ciency.

The most favourable machinery arrange-ment generally appears to be the twin main engine solution, coupled to two fi xed pitch propellers, and incorporating TES systems for utilisation of the waste heat and supplementary production of electrical power.

For the machinery arrangements based on the dual fuel ME-GI main engines, the selection of two main engines coupled to two fi xed pitch propellers and incor-porating either TES systems alone or a combination of TES systems and PTO

TES = Thermo Effi ciency System, PTO = Power Take Off, DG = Diesel Gas,NPV = Net Present Valve (see Fig. 21)

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5510-0026-00ppr.indd 27 2007-10-01 13:51:13

systems is found to be the most optimal choices, with the latter arrangement be-ing about equal to the fi rst mentioned, or in some cases even better if emission fees are incorporated in the analysis.

Generally, emission considerations fa-vour the selection of the dual fuel ME-GI engine over the HFO engine and, with today’s fuel prices, the dual fuel ME-GI engine is found to be the most optimal choice. However, as already mentioned, project specifi c factors, such as a re-quirement for a fi xed amount of LNG to be delivered, need to be addressed in specifi c cases, and may infl uence the balance between the ME-GI engine al-ternative and the HFO engine alternative.In the ME-GI engine for LNG carriers, any ratio of gas and heavy fuel, from 0% gas and 100% fuel to 95% gas and 5% fuel, can be used at any load above 30% – below it is fuel only. Hence, full fuel/gas fl exibility is ensured, while accept-ing a wide range of variation in sulphur throughput.

Using twin-engine propulsion in a twin skeg arrangement has proven to pro-vide propulsion power savings of 5-8%, compared with the single screw propul-sion alternative for large LNG carriers of 150,000-270,000 m3 capacity, due to their large breadth/draught ratio.

The most optimal and fl exible choice of machinery for the LNG carrier appears to be a twin-engine ME-GI installation in combination with a double Thermo Effi ciency System (one for each main en-gine), a 100% capacity gas compressor plant and a 100% capacity reliquefaction plant.

This propulsion system in combination with suffi cient HFO storage tank ar-rangements would allow a fully fl exible operation of the vessel, optimised for any future business environment.

Concluding Remarks

To enter the market for a demanding ap-plication such as LNG vessels calls for a high level of know-how and careful stud-ies by the shipyard, the engine builder, the compressor maker as well as the engine designer.

A tailor-made ME-GI propulsion solution together with a fuel gas supply system is now available, which optimises the key application issues such as effi ciency, economy, redundancy and safety. This system is based on conventional, proven technology and can be applied with con-siderable benefi t on to LNG carriers in the range of 150 kcum up to 260 kcum.

References

- ‘‘High reliability of the Laby® LNG BOG compressor with the unique sealing system’’, (P. Ernst, Burck- hardt Compression AG)

- ‘‘Dual-fuel concept – analyses of fi res and explosions in engine room’’, (Asmund Huser, DNV Consulting).

- ‘‘Alternative Propulsion for LNG ships by Low Speed ME-C and ME-GI Engines’’, (Niels B. Clausen, MAN Diesel A/S)

- ‘‘LNG Gas Carrier with High-pressure Gas Engine Propulsion Application’’, GasTech 2006, Abu Dhabi, United Arab Emirates, (John Linwood, Burckhardt Compression AG, Switzerland, Jong-Pil Ha, Hyundai Heavy Industries Co., Ltd, Korea, Kjeld Aabo, MAN Diesel A/S, Copenhagen, Denmark), Rene S Laursen, MAN Diesel A/S, Copenhagen, Denmark

Appendices

I Lifetime of compressor parts

II Reference list for LNG boil-off gas installation

III Gas system P&I diagram for fuel gas compressor

IV ME-GI schematic, showing the GI assessment to an ME engine

V LNG carrier voyage illustration

VI Safety aspects

VII Hydrodynamics and vibrations on LNG carriers

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Appendix I

Average lifetime of compressor parts 6LP250B-5S_1

DescriptionQ

uant

ityYear

1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th

Hours* ( x 1,000)

8* 16* 24* 32* 40* 48* 56* 64* 72* 80*

Crank gear

Shaft seal 1

Main bearings 6

Guide bearings (crankshaft) 1

Connecting rod bearings 6

Crosshead pin bearings 6

Crossheads 6

Guide bearings (piston rod) 6

Oil scrapers 6

Piston

Piston rod 1st to 3rd stage 4

Piston rod 4th to 5th stage 2

Piston skirt 1st to 3rd stage 3

Piston 4th to 5th stage 2

Piston rings 4th stage 10

Piston rings 5th stage 10

Piston guide rings 4th stage 2

Piston guide rings 5th stage 2

Packing

Packing 1st to 3rd stage 4

Packing 4th to 5th stage 2

Valves

Suction valves 8

Discharge valves 20

Controlled suction valves

1st to 3rd stage 8

4th to 5th stage 4

Note: The average lifetime is based on a compressor running 8,000 hours per year

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No. Type Commis-sioned

Installation Operator

1 4D300-3E 1986 Adnoc LNG Terminal Abud Dhabi

ADGAS

4 4D300-2C 1990 CPC LNG Terminal Taiwan

Chinese Petroleum Corp.

2 4D300-2D 1993 Pyeong Taek Terminal Korea

Korea Gas Corperation

2 2D250B-2C_1 2003 BILBAO LNG Terminal Spain

Bhaia de Bikaia Gas

3 2D250B-2C_1 2003 SINES LNG erminal Portugal

Transgas-Atlantico

2 2K90-1A_1 2004 BIJLSMA LNG H705 LNG Carrier

Knutsen Os Shipping

2 4D300B-2K 2005 Barcelona LNG Spain

Enagas

3 4D250B-2N_1 2005 SAGGAS Terminal GNL Sagunto, Spain

Regasagunto UTE, Madrid

3 2D250B-2C_1 (2007) Reganosa, Murgados Spain

Regasifi cadora del Noroeste S.A.

Reference List of LNG Boil-off Gas Installationswith Burckhardt Compression labyrinth-piston compressors

Appendix II

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Appendix III

Gas system for fuel gas compressor

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Appendix IV

ME-GI schematic

32

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Appendix V

Fig. App. V: LNG carrier voyage illustration: Top: ME-GI engine load diagram; Middle: 6LP250 compressor; Bottom: BOG tank pressure

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LNG carrier voyage illus-tration

During an LNG carrier voyage, gas is available both on laden and on ballast voyage. It is therefore expected that the fuel-oil-only mode will be used during manoeuvring, canal voyage and, possisly, during a period with an engine failure. At any other voyage situation gas will be used as fuel. BCA and MAN Diesel have worked out a simulation for such a voyage, an example is shown in Fig. App. V.

Below is a summary of the main conclusions from the illustrations:

On low engine load, e.g. 30%, and laden voyage - only HFO is being burned in the engine - BOG gas tank pressure increases until it reaches its max. level, then compressor starts sending the BOG to the GCU.

at 50% engine load and laden voyage - BOG tank pressure high - the com-pressor sends BOG to the engine and to the GCU at the same time.

at 90% engine laden voyage - the en-gine is burning all BOG generated - BOG tank pressure operates within determined limits. If the tank pressure reaches its min. level, the amount of pilot oil is increased.

at 90% load ballast voyage and with the engine running in fuel-oil-only mode, a very slow pressure increase takes place due to the huge BOG buffer volume.

at 90% load ballast voyage - mini-mum fuel mode - a large BOG gas amount is being burned before the BOG tank pressure reaches the tank pressure min. value. At min. pres-sure, add up with pilot oil starts.

•

•

•

•

•

In this example, a boil-off rate of 0.12% is used in laden voyage, and a 0.06% boil-off rate in ballast condition. The tank pressure phigh = 1.17 and plow = 1.03, they may differ from laden to ballast, in this example the pressure range is the same.

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Safety Aspects

Hazard identifi cation (Hazid) process for ME-GI engines for LNG carriers

LNG operators are seriously conside ring the ME-GI propulsion solution for use on LNG carriers and, as described previously, they require hazid considerations in connection with the use of gas in the engine room. A Hazid investigation of the complete gas system, from the gas storage tanks to the engine inlet, has therefore been carried out and completed.

The Hazid study was carried out in cooperation between the Hyundai shipbuilding and engine building divisions, Burckhardt Compression, and MAN Diesel. Det Norske Veritas (DNV) participated as consultant at the meetings as well as being responsible for granting their acceptance of the procedures and, ultimately, fi nal approval for use on board LNG carriers.

Scope of the Hazid study

The scope of the study was minimised to cover only those components and systems relating to gas running opera-tion, i.e.:

LNG storage tanks, producing boil-off gas

forced LNG vaporiser

three 50%, 5-stage reciprocating compressors, fed with BOG and force-vaporised LNG (Burckhardt Compression)

diesel fuel storage and supply system

two MAN B&W ME-GI engines adapt-ed for dual fuel operation (natural gas /diesel)

oxidiser (GCU)

•

•

•

•

•

•

fi re and gas detection, and air ventila-tion systems for enclosures.

For each of the components or subsystems, the Hazid considered possible malfunction of instruments, control systems or equipment failures. To study the ‘worst case’ consequences, the assessment was initially made without consideration for any planned limiting measures.

Once the ‘worst case’ consequences had been identifi ed, the planned limiting measures were considered, and a judgment was made as to whether they were adequate with respect to the identifi ed hazards or to operational problems.

The outcome of the above was a report that was sent to each of the participating companies, and special attention was paid to the recommendations made in the report. Consequently, each of the participating companies had the opportunity to follow the recommendations and upgrade their design to a higher safety standard.

A total of 20 main system items were reviewed in the Hazid workshop, resulting in 22 recommendations. Failure Mode and Effect Analysis (FMEA)

Prior to performing a Hazid study of a projected 210,000 cum LNG carrier equipped with two 6S70ME-GI engines and a gas supply system, layout drawings, system diagrams and an FMEA study were prepared in cooperation between MAN Diesel, Burckhardt, the engine builder and the shipyard. FMEA is a method designed to identify potential failure modes for a product or process before the problems actually occur. The method was developed in the 1950s to identify problems that could arise from

• malfunctions of military systems. Over time, this method has been applied on other business areas as well, because the method has proved to be well suited for reviews of mechanical and electrical hardware systems, like for instance the ME-GI engine.

The FMEA can be described as a method of evaluating and documenting the causes and effect of component failures. The FMEA fi rst considers how a failure mode of each system component can result in performance problems for the overall system, secondly it ensures that appropriate safeguards are available to handle these situations.

Only known failure problems can be handled in the study and, therefore, the study gradually expands as more knowledge of the system is achieved in the course of the development of the system. To perform a full study, detailed knowledge of each component is required, both with respect to design and to operational behaviour. Accordingly, the system covered in the FMEA must be well defi ned before a useful FMEA can be fi nalised.

Normally, the FMEA only examines the effect of a single point failure on the overall performance of a system, but in some cases, where the consequences of two following failures can lead to a catastrophic result, it may be necessary to include double failures.

Risk evaluation methods

The FMEA incorporates a method to evaluate the risk associated with the potential problems identifi ed in the analysis. The method used is called risk priority numbers (RPN), and is de-scribed below.

Appendix V

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To assess risks by using the RPN method, the analysers must:

rate the hazard of each effect of failure

rate the likelihood of occurrence for each cause of failure

rate the likelihood of prior detection for each cause of failure (i.e. the likeli-hood of detecting the problem before it reaches the end user or customer).

calculate the RPN by obtaining the product of the three ratings:

RPN = hazard (B) x occurrence (A) x detection (E) = B x A x E

The RPN can then be used to compare issues within the analysis and to prioritise problems for corrective action, see the Table.

The FMEA study is a part of the requirements for approval from the classifi cation societies. The FMEA therefore formed part of the documentation that was delivered to DNV before the type approval documentation for the ME-GI engine was issued to MAN Diesel. With this approval, the electronically controlled gas injection system is approved for use on MAN B&W ME engines.

The compressor system from BCA reached the system approval from DNV in autumn 2006. Thesteps to reach this level are equal as described above.

•

•

•

•

Factor A

Degree Risk Occurrence A-Factor

1 very often less than 500h OH (operating hours) 10

2 often 500h to 1000h OH 8

3 occasional 1000h to 8000h OH 6

4 seldom 8000 to 24000h OH 3

5 very seldom

more than 24000h OH 1

Factor B

Degree Risk Hazard B-Factor

1 Danger to life. Failure can effect death of person

10

2 hazardous Operation will fail. Injuries of person possible 7

3 major A Operation will fail. No harm to person 5

4 major A Limited operation possible. No harm to person

4

5 minor No or minor effect on operation 1

Factor E

Degree Risk Detection E-factor

1 very rarely The detection of the hazard or failure is almost not possible. Feasibility 30%

10

2 reasonable The detection of the hazard or failure unlikely. Feasibility 60%

6

3 high It is feasible that the hazard or failure will be detected. Feasibility 99%

3

4 very high It is certain that the hazard or failure will be detected. Feasibility 99,9%

1

Table: Rating of hazards

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Hydrodynamics and vibrations on LNG carriers

One question raised by the operators of LNG tankers is dealing with the vibration level initiated by the use of two-stroke engines and the infl uence, if any, on the structure of the insulated LNG storage tanks.

In this connection, investigations have been carried out in cooperation with Det Norske Veritas and various shipyards.

There are various kinds of excitation sources in a ship. The most dominant sources in a general cargo ship are the propeller and the two-stroke diesel engine.

In a traditional LNG carrier with steam turbines, the propeller is the only dominant excitation source, because the turbine rotor or gears are no source of excitation.

Application of two-stroke diesel engines on large LNG carriers with a twin skeg hull design and twin propellers results in reduced propeller loads. Compared with existing single propeller designs, signifi cantly reduced pressure pulses and vibrations are obtained, partly thanks to the reduced cavitations.

The application of twin propeller and two-stroke diesel engines on an LNG carrier may need the below additional anti-vibration and anti-noise analysis and countermeasures. However, owing to the high number of two-stroke diesel engine installations built, a number of standard countermeasures against vibrations have been developed. Furthermore, expertise and various countermeasures have been developed to cope with extraordinary vibrations.

Appendix VII

Hull girder vibration analysis infl uencing cargo containment system due to the 2nd order external moment of main engine. The most effective way to minimise such an external moment of the diesel engine is an application of the well proven 2nd order moment compensator, which neutralize the 2nd order moment.

Fatigue assessment of cargo contain-ment system due to excitation by the guide force moment of the main engine. Analysis needed. Countermeasure: top bracing or electrically driven moment compensator.

Local vibration and noise analyses for engine room. Two-stroke diesel engines and corresponding auxiliary machinery are dominant noise sources in an engine room and, therefore, anti-noise activities considering the optimum arrangement of working spaces and proper insulation should be performed more rigorously than usual for LNG carriers. For this purpose, an extensive noise analysis is required to evaluate the noise levels in accommodation, ECR and working areas. Furthermore, a detailed analysis of the local vibration behaviour is necessary to achieve a good vibration status of large machinery and local structures in the engine room area.

Deckhouse vibration analysis coupled with double-bottom mode. The double-bottom structure between two diesel engines is more fl exible and easy to vibrate with diesel engine excitations. This is checked by analyses of deckhouse coupled with double-bottom, and if necessary structural countermeasures are introduced.

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Technology

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