Feedback from
MAN B&W MC engines in service has shown average liner wear rates as low as
0.05 mm/1000 hours and cylinder overhauling intervals of 6000–8000 hours.
R&D targeting enhanced durability and extended mean times-between-overhauls
(TBOs) has focused on:
Engine structure: experience has shown that the
jacket cooling water around the lower part of the cylinder frame may in some
cases result in a corrosion attack and thus lead to higher wear rates. Cooling
water in the lower part of the liner has therefore been omitted, thus
increasing the temperature of the liner surface. The modification has been
introduced on the latest engine versions (Figure 10.19). Cracks were
experienced in the horizontal support for the forward main bearing in the
thrust bearing housing. Investigations revealed that, under special service
conditions, the design seemed to leave too narrow a margin; improper welding
quality was also a contributory factor. While such rib cracks do not imply any
risk with respect to the service of the ship in the short term, they should be
rectified at the first available opportunity. For engines in service, the
designer has evolved a rectification method which, in addition to remedying the
cracks, also increases the margins. Repairs have been successfully carried out
using this procedure. The standard design has also been changed to address the
above conditions.
Piston rings: Investigation showed that the overhauling interval could
be prolonged by introducing approximately 30 per cent higher piston rings in
the two uppermost ring grooves. Such rings of a special alloyed grey cast iron
were therefore introduced as standard on new engines. A further extension of
maintenance intervals can be secured by plasma (ceramic) coating or chrome
plating the surface of the uppermost piston ring. These coated rings are
available as options. Experience showed that in some cases the normal
oblique-cut piston ring gap led to a high heat input to the liner when all cuts
coincided. An improved design of S-seal ring (double-lap joint) was developed
and tested to achieve a completely tight upper ring and then reduce the
pressure drop across this ring by introducing a number of small oblique slots
to secure a controlled leakage. The temperature distribution and pressure drop
measured across the ring were promising and very clean ring lands resulted.
Rings of vermicular iron have also shown promising results. The higher piston
topland and higher top rings (the uppermost featuring a pressure-balanced
design) were introduced on engines with higher mean effective pressures to
enhance reliability and extend times between overhauls (Figure 10.5).
Cylinder liner: factors to be considered when
designing cylinder liners include material composition, strength, ductility,
heat transfer coefficient and wear properties. The original MC-type liners with
cast in pipes unfortunately suffered from cracks but the introduction of the
bore-cooled liner solved the problem (Figure 10.20). Such liners have been
fitted to all new engines produced in recent years. The change to a bore-cooled
liner made of grey cast iron for engines originally supplied with liners of the
cast-in cooling pipe design, however, could not be easily effected. Various
parallel developments were therefore initiated, based on the use of stronger
materials, and improvements to the original design introduced. A modified
design of the cast-in pipes, in combination with a tightening-up of the
production and quality specifications, has led to considerably improved
reliability from this type of liner. The modified liner is the standard spare
part supply for older MC engines. The standard specification for new engines is
a bore-cooled type made of Tarkalloy C grey cast iron.
The inside surface temperature of the liner greatly
influences the general cylinder condition. Traditionally, the cooling system
has been laid out to match the maximum continuous rating load but there is an
advantage in controlling the inside liner surface temperature in relation to
the load. MAN B&W has investigated and tested different solutions for
load-dependent cylinder liner cooling. One system simply adjusted the cooling
water flow through the original cooling ducts in the liner but the results were
not promising.
Another system features different sets of cooling
ducts in the borecooled liner, the set deployed depending on the engine load.
At nominal power and high loads the inner row of ducts is used to cool the
liner, yielding the highest cooling intensity. In the intermediate load range
the cooling function is shifted to the next set of ducts which are located
further away from the inner surface; this means that the cooling intensity is
reduced and the liner surface temperature is kept at the optimum level. At very
low loads both rows of cooling ducts are bypassed in order to further reduce
the cooling intensity. Tests showed that the optimum liner temperature could be
maintained over a very wide load range and that this system was feasible but
the added complexity had to be weighed against the service advantages.
The operating condition of cylinder liners and
piston rings is to a great extent a function of the temperature along the
liner. The upper part is particularly important and a triple-fuel valve
configuration (see section below) reduces thermal load while, at the same time,
the pressure-balanced piston ring and high topland ensure an appropriate
pressure drop across the ring pack and control the temperature regime for the
individual piston rings. For monitoring the temperature of the upper part of
the liners MAN B&W offers embedded temperature sensors and a recorder.
Alarm and slowdown temperature settings allow the operator to take
proper action to restore proper running conditions if, for example, a piston
ring or fuel valve is temporarily or permanently out of order. Other features,
such as an uncooled cylinder frame, serve to increase slightly the wall
temperature on the lower part of the cylinder liner while, at the same time,
reducing production costs (Figure 10.19). The rise in wall temperature is aimed
at counteracting the tendency towards cold corrosion in the lower part of the
liner.
Exhaust valves: a high degree of reliability is claimed for the MC
exhaust valve design, and the recommended mean TBOs have generally been
achieved. The average time between seat grinding can also be considerably
increased for a large number of engines. Examples of more than 25 000 running
hours without overhaul have been logged with both conventional and Nimonic-type
exhaust valves. The Nimonic spindle is now standard for the 600 mm bore engines
and above; and a steel bottom piece with a surface-hardened seat is specified
to match the greater hardness of the Nimonic spindle at high temperatures. The
combination of spindle and bottom piece fosters a mean TBO of a minimum 14 000
hours.
Some cases of wear of the spindle stem chromium
plating may be related to the sealing air arrangement. A new sealing air system
was therefore designed, incorporating oil mist and air supply from the exhaust
valves air spring (Figure 10.21). In-service testing gave promising results:
completely clean sealing air chambers and virtually no wear of either the spindle
stem or the sealing rings. The system is now standard for new engines and can
be easily retrofitted to those in service.
Cold corrosion of the exhaust valve housing gas
duct led to lowerthan-expected lifetimes for a number of valves, particularly
those installed in large bore engines. The corrosion attack occurs adjacent to
the spindle guide boss and in the duct areas at the cooling water inlet
positions (Figure 10.22). The problem has been addressed by new housings
designed with thicker gas walls which are now standard fitments for new engines
and spares (Figure 10.23). For engines in service, the following repair methods
and countermeasures have proved effective in dealing with corrosion attacks in
the exhaust valve housing: high velocity sprayed Diamalloy 1005 coating in the
gas duct; and repair welding with gas metal arc welding (MIG-type), preferably
in conjunction with the sprayed coating.
Bearings:
there should be no problems related to production or materials since the
manufacture of whitemetal-type main and crankpin bearings is well controlled
(Figures 10.24 and 10.25). Some cases of production-related problems were
traced to the use of copper- or lead-polluted whitemetal or the lack of proper
surface treatment of the steel back before casting the whitemetal. Based on
service feedback, the main bearing design has been modified to secure a wider
safety margin. The modifications include a larger bore-relief in order to
prevent the mating faces from acting as oil scraping edges in case of displacement,
and an increase of the side and top clearances to raise the oil flow to the
bearing, thus keeping the temperature at a level which sustains the fatigue
strength of the whitemetal. The modified bearing design was introduced as the
standard for new engines and also made available as spares for engines in
service. It is fully interchangeable with existing bearings but the shells must
always be replaced in pairs.
∑
Fuel injection system: the fuel valve used on MC engines operates without any
external control of its function. The design has worked very well for many
years but may be challenged by the desire to maintain effective performance at
very low loads as the opening pressure has to be increased with increasing
maximum combustion pressure to prevent the blow-back of gases into the fuel
system at the end of the injection cycle. At low loads this high opening
pressure might lead to irregular injection, speed variations and a risk of
fouling of the engine gasways.
Various
solutions have been investigated, among them a fuel valve with variable opening
pressure (Figure 10.26) which allows low load operation to be improved by
reducing the opening pressure in the relevant load range. The opening pressure
is controlled by a spring and the actual fuel oil injection pressure. At low
load the opening pressure is controlled by the spring alone but, when the
injection pressure increases at higher loads, this higher pressure adds to the
spring force and the opening pressure increases (a small amount of fuel oil enters
the space between the upper spring guide and the slide; the spring that
controls the opening pressure is thereby compressed and the spring force
increased). Under decreasing load, the leakages will reduce the opening
pressure to an appropriate level within the course of a few injections.
Burning of the
neck of the fuel valve nozzle on engine types equipped with three fuel valves
per cylinder (B&W 90GFCA and 90GB models) has been noted with nozzles made
from standard precision-cast Stellite which, in other respects, has yielded
excellent service and lifetimes well above 15 000 hours. Several improved
production processes and new materials have been investigated. A promising
alternative to casting was found in the hot isostatic pressure (HIP) method
which delivers a more uniform material composition. Fuel nozzles produced by
the process achieved a considerable reduction in the neck-burning rate when
tested on the engines concerned
Research also
produced a new patented material composition with the same excellent wear
properties as those of Stellite but additionally offering an ability to
withstand hot corrosion neck burning. A truly heavy duty fuel valve nozzle is
thus available when required.
MAN B&W
has addressed the potential for enhancing the reliability of combustion chamber
components under increasing mean effective pressure conditions by rearranging
the fuel valves in the cylinder cover. Originally, the 90-type large bore
engine was provided with three fuel valves per cylinder. A number of tests were
made in the early 1980s using only two valves which, with the contemporary mean
effective pressures and maximum pressures, showed an advantage in the form of a
slightly reduced specific fuel consumption and only a minor increase in the
heat load. A two-valve configuration was consequently introduced. Rising mean
effective pressures in the 1990s, however, encouraged further testing with
three fuel valves per cylinder, the measurements showing a reduced temperature
level as well as a more uniform temperature distribution. A three-valve
configuration was subsequently introduced on the K80MC-C and K90MC/MC-C models
to enhance engine reliability, while the S90MC-T and K98MCC models were
designed from inception with such an arrangement.
Many owners
today prefer to have their fuel equipment overhauled by a shoreside workshop
but this means that engineroom staff lose familiarity with some vital
components, particularly the fuel valves. In a number of cases MAN B&W has
found that nozzles with different spray hole sizes and spray patterns have been
mixed. Up to four different types have been found on the same engine and even
different types on the same cylinder unit. Such mixing may have an adverse
effect on the running condition. A problem noted on many ships is incorrect
tightening of the fuel valve spring packs—in some cases caused by faulty spring
packs. This can lead to malfunctioning of the fuel valve (such as sticking) due
to deformation of the components inside the valve. The consequences have been
jumpingfuel valves and cracked high pressure injection pipes. It is therefore
important, MAN B&W Diesel warns, to follow instructions covering the
correct tightening of the spring packs.
Fuel pumps: the lifetime of the MC fuel pump plunger and barrel has proved to
exceed 50 000 hours. Only a few cases of sticking pump plungers have been
reported, and these generally resulted from foreign particles entering the
system. The introduction of puncture valves at the top of the pump improved
crash stop performance and increased engine shutdown safety. Some problems were
experienced with the sealing rings mounted around the plunger. The rings are
intended to prevent fuel oil from leaking into the camshaft oil and lowering
its flashpoint. Different sealing rings were tested after the introduction of
the MC engine, the original ring—with an O-ring as a spring element—having been
used on the previous engine type with relatively good results. Under attack
from fuel oil, however, the O-rings became hard and lost their tension too quickly,
making the sealing ring ineffective. Work on improving the material quality as
well as the sealing ring design resulted in spring-loaded lip rings which are
now standard. A new so-called umbrellatype fuel oil pump design (Figure 10.27)
features a sealing arrangement which eliminates the risk of fuel oil
penetrating the camshaft lube oil system (a separate camshaft lube oil system
is therefore no longer necessary). The uni-lube oil system introduced as
standard allows reductions in installation costs, maintenance and space over
the separate systems previously used: tanks, filters, pumps and piping for the
camshaft system are eliminated.
Chain
drives: the chain drives for the camshaft and moment compensators have
functioned well and, if the chains are kept tightened to MAN B&W
instructions, the rubber guide bars will hardly ever need replacement. Since
the chain tightening procedure on some ships might prove difficult for crews,
automatic chain tightening was introduced as an aid to reducing the onboard
maintenance workload. The tightening force is applied by a hydraulic piston to
which oil is fed via the engines lubricating/ cooling oil system. The hydraulic
chain tightener is also available for the front second-order moment compensator
drive. Positive service experience encouraged the fitment of the automatic
tightener as standard on 50–90 bore MC engines.
Installation
refinements: a simplified type of hydraulically-adjustable top bracing
introduced for vibration control in 1989 was designed to cope with hull
deflections giving rise to excessive stresses in the traditional mechanical
type of top bracing. The system features a hydraulic cylinder which is a
self-contained and does not need an external pump station. Instead, a pneumatic
accumulator is built into the piston. The hydraulic cylinder is pre-tightened
by the air pressure in the accumulator. When the engine is started the oil
pressure in the cylinder increases because of the influence of the guide force
moments, and the hydraulic cylinder adjusts its length to cope with slow hull
deflections and thermal influence (Figure 10.28).
The system
worked satisfactorily after teething problems were solved but a drawback of the
hydraulically-adjustable top bracing is its higher first cost over the
traditional mechanical bracing. An improved mechanical top bracing,
incorporating two beams, was designed and tested in service from early 1997
(Figure 10.29). This is rigid in the athwartship direction and sufficiently
flexible in the longitudinal direction to adapt to movements between engine and
hull.
Simplification
has also benefited engine seating arrangements. Feedback from shipyards
indicated a preference for epoxy supporting chocks; epoxy chocks therefore
became the standard in an effort to reduce installation costs, with the
previously used cast iron supporting chocks as an option. Normal holding-down
bolts are used in association with epoxy chocks but with the tightening force
substantially reduced. The lower force is due to the higher coefficient of
friction as well as the need to comply with the permissible specific surface
pressure of epoxy. The general application of epoxy thus permits the use of
thinner bolts which—following successful testing in service—have been
introduced as standard. Larger engines (L42MC and upwards) have had one side
chock per crossgirder per side. Investigations and experience showed that this
number could be halved without any problems. An even simpler solution proposed
features a side chock design that allows epoxy to be used as a liner.
Installation man-hours are saved since the location of the side chock itself is
much less demanding in terms of accuracy (Figure 10.30).
POST-1997 SERVICE EXPERIENCE
A
service experience report covering mainly larger bore MC and MCC engines
commissioned after 1997 indicated that a significant increase in reliability
and longer times-between-overhauls had resulted from continuous development and
updating. The following contributions were highlighted:
Cylinder condition
A
number of measures have contributed to positive developments in the cylinder
condition in recent years, notably the Oros combustion chamber,
high topland pistons, the piston cleaning
(PC) ring, controlled pressure relief (CPR) piston rings, alu-coating of piston
rings and wave-cut cylinder liners.
With the Oros configuration, the combustion
air is concentrated around the fuel nozzles and the distance from the nozzles
to the piston top is increased. This results in a lower heat load on the piston
top and a basically unchanged heat load on the cylinder cover and exhaust
valve. The higher topland and the PC ring have proved very beneficial in
avoiding a build-up of lube oil-derived deposits on the topland. Such deposits
(via the hard face sponge effect) scrape off and absorb the oil film, leaving
the naked liner wall vulnerable to extensive wear and/or scuffing. A high
topland piston also means that the mating surfaces between the cylinder liner
and the cover are lowered, thus reducing the thermal load on the liner and
improving the conditions for lubricating the liner. (This was taken into
account before introducing the Oros configuration.)
Cracked cylinder liners are
now rarely reported thanks to the successful introduction of countermeasures,
such as bore-cooled liners and (for smaller bore engines) slim liners. For
engines originally specified with cast-in cooling pipes in the liners, the
later design with oval pipes has stopped the occurrence of cracks.
Safe and
stable running-in is secured by the semi-honed surface of the liner in
conjunction with alu-coated piston rings. A thin layer of alu-coating is worn
off the rings during the first 1000–2000
hours, making it possible to reduce the breaking-in and running-in time as well
as the cylinder oil feed rate during most of the breaking-in period. This
benefits the enginebuilder, who can cut the delivery time of an engine, and the
operator, who receives an engine with improved running-in conditions.
The
surface of MAN B&W cylinder liners is described as semihoned. The
semi-honing process cuts off the tops of the wave-cut,thus reducing the
necessary breaking-in between rings and liner surface, while still retaining
circumferential pockets for lube oil. The alu-coated piston rings remove the
remaining broken or damaged cementite from the liner surface during the initial
wear period (effectively performing a "free of charge" full honing).
Low and uniform
cylinder wear rates and a good liner condition are revealed by inspections of
S-MC, S-MC-C, K-MC and K-MC-C engines. While initial (running-in) wear is
naturally higher, the wear rate of the largest bore (90 and 98) models is
reduced to less than 0.05 mm/ 1000 hours after around 1500 hoursrunning, which
is considered very satisfactory.
Initial
wear is part of the running-in of liners and piston rings, and high wear is
expected during this period. Experience with the Alpha Lubricator (see above)
indicated that there was significant potential for cylinder lube oil reduction
while retaining a fully acceptable wear rate and mean time-between-overhauls.
The top ring design of the K98 engine was upgraded to increase the safety
margin against breakage in response to a number of failures. The production
process at the sub-suppliers was also changed to reduce such incidents. The
upgrading involved a number of changes, including relocation of the controlled
leakage grooves, reduction of the number of grooves from six to four (the same
leakage area is achieved by applying wider grooves) and modified surface
machining of the grooves to avoid fine cracks from the outset (Figure 10.31).
Bearings
A decrease in the number of reported main bearing failures has been noted
since 1998, when a number of features were introduced to the bearing design,
bearing adjustment and engine/shaftline installation. These major updates
included the introduction of the Optimum Lemon Shape (OLS)-type main bearing as
an evolution of the Mark 5 bearing type, featuring reduced top and side
clearances. Service experience has confirmed the efficiency of the new bearing
type, reflected in a significantly reduced number of reported failures. Main
bearing damage can sometimes still occur, however: for example, due to poor
bonding of the bearing metal. In almost all cases, main bearing damage is
initiated from a fatigue crack at the edge of the bearing, the aft
edge/manoeuvring side being the most common point of initiation. Geometrical
non-conformities, often involved in these cases, further increase the damage
frequency as margins established during the design phase are reduced.
Calculations, combining the dynamics of the complete crankshaft with the
hydrodynamic and elastic properties of the bearing, have provided detailed
information on the mechanisms leading to local loading of the main bearing
edges. The calculations have indicated that a slight radial flexibility of the
bearing edge will significantly increase the overall minimum oil film thickness.
At the same time, the maximum oil film pressure will be reduced. A bearing
design with flexible forward and aft edges of the bearing shell was
successfully tested, the flexibility achieved by removing the contact between
the shell and bedplate at the end portions of the shell. The unsupported width
of the shell is equal to the shell thickness. Apart from the flexible edges,
the properties of the bearing are similar to those of the OLS-type and the
design provides a larger safety margin in the event of geometrical
non-conformities. A thin shell bearing design has been introduced to the latest
engine types. The main bearings of the small and medium bore models (S46MCC to
S70MC-C) are lined with AlSn40 and provided with a PTFE runningin coating as
standard. The bearings of the large bore models are lined with white metal. Few
damage incidents to the thin shell main bearings have been reported. A revised
engine installation recommendation—including an updated shaftline alignment
procedure and a differentiated bearing height in the aft end of the
engine—resolved cases of repeated damage to the aft-end bearings. This was
presumably caused by missing static load, particularly in the second aftmost
main bearing during normal operating conditions. The new alignment procedure,
exploiting pre-calculated bedplate sagging as well as vertical offsets to the
main bearing saddles, achieved a significant drop in the number of reported
damage incidents to the aft-end bearings. The crosshead bearings of both MC and
MC-C engines generally perform very satisfactorily, but cases of wiping have
been observed. Such wiping is of a cosmetic nature but can sometimes cause
blockage of the oil-wedges that normally build up the oil film to the
padsinside the bearing. Disturbance of this oil film build-up could result in
slight fatigue damage just behind the blocked area of the oil-wedge. If
observed at an early stage during inspections, however, the problem can be
solved by removing the wiped lead from the oil-wedge. The crankshaft thrust
bearing introduced on Mark 5 engines has solved the problem of cracks in the
horizontal support plates. By making this plate (which connects the fore and
aft thrust bearing supports) in one piece, and shaping it like a calliper, a
significantly wider design margin is derived even though engine outputs (and
hence propeller thrust) have been increased. No cracks have been reported on
engines with this so-called Calliper design thrust bearing (Figure 10.32). The
bearing saddles have remained free of cracks, in compliance with pre-calculated
stress levels.
Exhaust valves
Nimonic exhaust valve spindles are well accepted
now that operators have become acquainted with the long-lasting seat
performance, despite dent marks. Nimonic spindles are standard for 50MC and
60MC/MCC engine models upwards; and Stellite spindles are standard for smaller
models. Corrosion in the valve housing was effectively minimized by introducing
the optimized cooling water system; this raises the wall temperature in the
housing above the critical level for the formation of acid on the gas side of
the duct. The high temperature level dictates a cast iron spindle guide
bushing.
Wear of the
previously chrome-plated spindle stem has been effectively reduced by the
HVOF-based cermet coating introduced in 1997. Furthermore, wear and corrosion
problems at the spindle guide/ spindle/seal area, caused by combustion
products, have been minimized over the years by design changes to the sealing
air system. Since the stem seal had difficulty in reaching lifetimes similar to
those of the valve seats, however, it was decided to replace the sealing air
system by a lubricating device tested with good results in long term service.
Reduced wear of the spindle stem (HVOF coating) and of the long spindle guide (grey
cast iron) has been achieved but the lifetime of the stem seal itself is still
sometimes too short. Tests with oil as the sealing medium instead of air showed
very low wear rates on the seals, along with a high cleanliness level on the
surfaces of the spindle stem and spindle guide. A system delivering the
necessary dosage of only approximately 1 kg/cylinder/day was developed for the
medium and large bore engines. It is located in the top of the exhaust valve
and fed with oil from the valves hydraulic system; oil is fed to the spindle
guide via a small pipe. The sealing oil is taken from the circulating oil and
is therefore part of the necessary minimum oil consumption for keeping the
system oil viscosity and base number (BN) level at the prescribed equilibrium.
The latest design of exhaust valve on the small and medium bore engines
occasionally suffered from cooling water leaks at the lowermost O-ring between
the bottom piece and cylinder cover. Investigations resulted in this O-ring
being replaced by a special Teflon seal with spring back-up (U-seal).
Tests showed
that the best way of increasing valve seat lifetime was by altering the seat
geometry of the bottom piece to the patented Wseat configuration, now standard
on all MC/MC-C engines (Figure 10.33). Results are even better when used with
the new type of slide fuel valves.
Engines
featuring the Oros combustion chamber (see section above) leave little distance
between the piston top and the underside of the exhaust valve spindle. The
usual extended lift system for releasing high hydraulic pressure cannot
therefore be applied. A safety valve located in the actuator is used instead.
Unfortunately, a few cases of damaged exhaust valves and camshaft sections were
experienced due to different external factors, including an insufficient
release action of the safety valve. The valve was subsequently redesigned so
that, once activated, it implements a special function to keep it open for
around 20 seconds. In addition, a disc spring was introduced in the exhaust
valve on top of the spindle guide to avoid damage to the air piston in the
event of over-shoot/extended lift of the valve spindle. Fuel injection system
In general, the fuel pumps work well and without difficulties, although a few
incidents have been experienced and addressed:
A combined puncture and
suction valve used on the S60MC-C, S70MC-C, S90MC-C and K98MC-C engines
originally featured a bellow as a substitute for the conventional sealing rings
in order to benefit from a component needing little or no maintenance. The
reliability of the bellow was not satisfactory,
however, and a new design eliminating this element was introduced for
the above engine models.
Fuel pumps without shock absorbers were introduced
on these same engine types as a cost reduction measure but this resulted in
annoying (although not damaging) pressure fluctuations in the fuel supply
system. Even though these measured fluctuations were within MAN B&Ws and
the classification societiesdesign limits, shipowners experienced problems with
shipyard-installed equipment such as pumps, filters and pre-heaters. The
problems were avoided by re-introducing a shock absorber on the fuel pumps of
these engines.
A new type of gasket between the fuel pump housing
and the top cover was introduced to cure leakage experienced on some K98MC/MC-C
and S90MC-C engines. The primary sealing is secured with a viton ring,
protected against corrosion attack by a steel bushing. A soft iron plate of the
same shape as the original seal forms the groovefor the square viton ring.
Fuel pump top cover fractures were experienced on
small and medium bore engines, initiated at the position where the inclined
drillings for the high pressure pipes intersect with the central bore. The
cause of the failure in all cases was related to roundings that did not fulfil
the MAN B&W specification. A design change has improved the safety margin
against failures, simplified manufacture and made the component less sensitive
to minor tolerance deviations (Figure 10.34).
Fractured fuel
valve nozzles were found on large bore engines where slide-type valves had been
standard for some years. The main reason for the cracks was residual stress
from machining, but the high temperature of the valve nozzle itself also
contributed to fracturing because of the consequential high mean stress. The
problem was cured by optimizing the production parameters.
|
MC
|
MCC
|
|
Wt
|
Weight
reduced by 10%
|
|
Power to
weight ratio increased
|
|
|
Long stroke
3.8:1
|
Super long
stroke 4.2:1, obtain high prop eff
|
|
Sfoc
167gm/bhp/hr
|
171
|
|
Mep 18
|
19
|
|
Peak pr 140
|
150
|
|
Mean piston
speed 8.02m/sec
|
8.46
|
|
Governor
mech
|
Epectro
mechanical
|
|
Starting air
distr axial
|
Radial type
|
|
Holding down
bolt access from db tank
|
Easy access
|
|
T/c eff low
|
High
|
|
Chain
smaller
|
Bigger
|
|
Emissioms
|
Appreciabke
emissioms
|
|
Single stay
bolt
|
Twin stay
bolts without pinching screw
|
|
Thck shell
brg
|
Thin shell
brg
|
|
Journal dia
less
|
Increased
|
|
Crosshead
heavy and telescopic pipe connected to pin
|
Simple n
telescopic pipe connected on guide shoe.
|
|
Liner thick
|
Slim
|
|
Piston top
with Inconel coating
|
No
|
|
Bottom ring
grooves thin chrome layer
|
Incr by
0.5mm
|
|
Top piston
ring
|
Cpr ring,
alu coat
|
|
Pistom land
low
|
High top
land with oros piston and the interface btw liner n head is lowered.
|
|
Fuel pump
dia Smaller
|
Adjusted
larger as per the more fuel reqd
|
