Woodyard D. (ed.) Pounders Marine diesel engines and Gas Turbines

Подождите немного. Документ загружается.

satisfactory effectiveness and a minimum pressure drop on the charge air side

and on the water side, the coolers are designed for air speeds of around 11 m/s

and water speeds in the tubes of 0.75 m/s. Charge air cooler effectiveness is

deﬁned as the ratio of charge air temperature drop to available temperature

drop between air inlet temperature and cooling water inlet temperature. This

ratio is approximately 0.8.

sCaVenging

It is essential that each cylinder should be adequately scavenged of gas before

a fresh charge of air is compressed; otherwise the fresh charge is contaminated

by residual exhaust gases from the previous cycle. Furthermore, the cycle tem-

perature will be unnecessarily high if the air charge is heated by mixing with

residual gases and by contact with hot cylinders and pistons.

In the exhaust turbocharged engine, the necessary scavenging is obtained

by providing a satisfactory pressure difference between the air manifold and

the exhaust manifold. The air ﬂow through the cylinder during the overlap

period has a valuable cooling effect; it helps to increase the volumetric efﬁ-

ciency and to ensure a low cycle temperature. The relatively cooler exhaust

also allows a higher engine output to be obtained before the exhaust tempera-

ture imposes a limitation on the satisfactory operation of the turbine blades.

In two-stroke engines, the exhaust/scavenge overlap is necessarily limited

by the engine design characteristics. In Figure 7.4 a comparison of the exhaust

and scavenge events for poppet valve engines and opposed-piston engines is

given. In the poppet valve engine, the camshaft lost motion coupling enables

the exhaust pre-opening angle to be 52° ahead and astern. In the opposed-pis-

ton engine, the exhaust pre-opening angle is only 34° ahead and 20° astern.

Against this, however, the rate of port opening in opposed-piston engines is

quicker than that in poppet valve engines. It should be noted, however, that

opposed-piston low-speed engines are no longer in production, and poppet

valves are used in the majority of new designs.

scavenging 179

TDC

BDC

Crank

Eccentric

Crank

Poppet valve Opposed piston

10° eccentric lead

Opposed piston

7° eccentric lead

TDC

Ahead

Astern

Ahead

Astern

Ahead

Astern

aus

Figure 7.4 single-acting two-stroke engine timing

180 Pressure Charging

In the four-stroke engine, the substantial increase in power per cylinder

obtained from turbocharging is achieved without the increase of cylinder tem-

perature. In the two-stroke engine, the augmented cylinder loading—if any—is

without signiﬁcance.

A strange fact is that the engine exhaust gas raises the blower air to a

pressure level greater than the mean pressure of the exhaust gas itself. This

is because of the utilization of the kinetic energy of the exhaust gas leaving

the cylinder and the energy of the heat drop as the gas passes through the tur-

bine. In Figure 7.5 the blower pressure gradient A exceeds the turbine pressure

gradient G by the amount of the scavenge gradient. The design of the engine

exhaust pipe system can have an important inﬂuence on the performance of

turboblowers.

The test results of turbocharged engines of both two- and four-stroke types

will show that there is an increase in the temperature of the exhaust gas between

the cylinder exhaust branch and the turbine inlet branch, the rise being some-

times as much as 95°C. The reason for this apparent anomaly is that the kinetic

energy of the hot gas leaving the cylinder is converted, in part, into additional

heat energy as it adiabatically compresses the column of gas ahead of it until,

at the turbine inlet, the temperature exceeds that at the cylinder branch. At the

turbine some of the heat energy is converted into horsepower, lowering the gas

temperature somewhat, with the gas passing out of the turbine to atmosphere or

to a waste heat recovery boiler for further conversion to energy. Though the last

thing to emerge from the exhaust parts or valves is a slug of cold scavenge air

which can have a cooling effect on the recording thermometer, adiabatic com-

pression can still be accepted as the chief cause of the temperature rise.

Turbine

Inlet

Blower

Exhaust

Inlet

Outlet

Figure 7.5 scavenge gradient

MaTChing oF TurBoBlowers

The correct matching of a turboblower to an engine is extremely important.

With correct matching, the engine operating point should be close to optimum

efﬁciency, as shown by the blower characteristic curve (Figure 7.6). In this dia-

gram the ordinates indicate pressure ratio; the abscissae show blower capacity.

During the initial shop trials of a new engine, with the turbocharger recom-

mended by the manufacturer, test data are recorded and analysed. If the tur-

boblower is correctly matched, nothing more needs to be done. Should the

matching be incorrect, however, the turboblower will supply charge air at either

too low or too high a pressure, or surging may occur at the blower. Mismatching

can usually be corrected by a change of turbine capacity and/or blower diffuser.

Compressors can be designed that maintain a high efﬁciency at a constant

pressure ratio over a wide range of air mass ﬂows by providing alternate forms

of diffuser for any one design of impeller. This range can be further extended

by the use of different impeller designs, each with its own set of diffusers,

within a given frame size of turboblower.

TurBoCharger surging

Surging is a common hazard which can be initiated by many events that result

in a shifting of the turbocharger operating point towards a surge line. Among

the possible reasons for surging are insufﬁcient engineroom ventilation; clog-

ging of the turbocharger intake silencer, compressor, scavenge air cooler, tur-

bine gas inlet protection grid, turbine nozzle ring and exhaust gas economizer

and silencer; deposits on the turbocharger turbine blades; damage to scavenge

Turbocharger surging 181

(

)

(

)

Pressure ratio

2.0

1.8

1.6

1.4

1.2

Correctly matched

Incorrectly matched

Engine operating line

50%

60%

70%

80%

90%

100%

76%

78%

1.0

Efficiency

Capacity ratio

Power increase = 50%

R.P.M. impeller

77%

Max. efficiency curve

Figure 7.6 Turboblower characteristics

182 Pressure Charging

air ﬂaps; damage or wear of fuel injection valves; excessive lubricating oil feed

rate; and wear of turbocharger components such as nozzle ring, turbine blades,

cover ring or shroud ring. Careful attention to engine maintenance and turbo-

charger cleaning is thus the best prevention against surging.

Too low an air-mass ﬂow at a given speed, or pressure ratio, will cause the

blower to surge, while too high a mass ﬂow causes the blower to choke, result-

ing in the loss of pressure ratio and efﬁciency at a given speed. The blower

impeller, as it rotates, accelerates the air ﬂow through the impeller, and the air

leaves the blower with a velocity that is convertible into a pressure at the dif-

fuser. If, for any reason, the rate of air ﬂow decreases, then its velocity at the

blower discharge will also decrease; thus there will come a time when the air

pressure that has been generated in the turboblower will fall below the delivery

pressure. There will then occur a sudden breakdown of air delivery, followed

immediately by a backward wave of air through the blower, which will con-

tinue until the delivery resistance has decreased sufﬁciently for air discharge to

be resumed. ‘Surging’ is the periodic breakdown of air delivery.

In the lower speed ranges surging is manifested variously as humming,

snorting and howling. If its incidence is limited to spells of short duration, it

may be harmless and bearable. In the higher speed ranges, however, prolonged

surging may cause damage to the blower, as well as being most annoying to

engineroom personnel. Close attention to the surging limit is always necessary

in the design and arrangement of blowers.

If one cylinder of a two-stroke engine should stop ﬁring, or is cut out for

mechanical reasons, when the engine is running above say 40–50 per cent of

the engine load, it is possible that the turboblower affected may begin to surge.

This is easily recognized from the repeated changing in the pitch of the blower

noise. In these circumstances the engine revolutions should be reduced until

the surging stops or until ﬁring can be resumed in all cylinders.

TurBoCharging sysTeMs

Four basic turbocharging arrangements are now used on marine diesel engines

(Figure 7.7):

Figure 7.7 exhaust turbocharging methods: 1, constant pressure; 2, impulse; 3,

pulse converter; 4, multi-pulse system

Constant pressure method—All the exhaust pipes of the cylinders of an

engine end in a large common gas manifold to reduce the pressure pulses to a

minimum with a given loading. The turbine can be built with all the gas being

admitted through one inlet, and therefore a high degree of efﬁciency is reached.

For efﬁcient operation the pressure generated by the compressor should always

be slightly higher than that of the exhaust gas after the cylinder.

Impulse turbocharging method—The exhaust gas ﬂows in pulsating form

into the pipes leading to the turbine from where it ﬂows out in a continuous

stream. Each of the gas pulses from the separate cylinders is fed to a corre-

sponding nozzle ring segment of the expansion turbine. By overlapping the

opening times of inlet and exhaust valves after the pulse decay, efﬁcient scav-

enging of the cylinders is possible.

Pulse converter method—Interactive interference limits the ways in which

the normal impulse system can be connected to groups of cylinder exhaust

pipes. However, with a pulse converter such cylinder groups can be connected

to a common ejector. This prevents return ﬂows and has the effect of smooth-

ing out the separate impulses. It also improves turbine admission, increases

efﬁciency and does not mechanically load the blading as much as normal

impulse turbocharging.

Multi-pulse method—This is a further development of the pulse converter. In

this case, a number of exhaust pipes feed into a common pulse converter together

with a number of nozzles and a mixing pipe. With this form of construction the

pressure waves are fed through with practically no reﬂection because the turbine

nozzle area is larger. The multi-pulse method makes a noticeable performance

increase possible compared with the normal impulse turbocharging method.

TurBoCharger ConsTruCTion

An example of turbocharger construction is provided by ABB Turbo Systems’

(Brown Boveri) VTR design, which remained unchanged for many years

(Figure 7.8).

The diesel engine exhaust gases enter through the water-cooled gas inlet

casing (50), expand in the nozzle ring (30) and supply energy to the shaft (20)

by ﬂowing through the blading (21). The gases exhaust to the open air through

the gas outlet casing (60), which is also water cooled, and the exhaust piping.

The charge air enters the compressor through an inlet stub (82) or the silencer

ﬁlter (80). It is then compressed in the inducer and the impeller (25), ﬂows

through the diffuser (28) and is fed to the engine via the pressure stubs on the

compressor casing (74).

Air and gas spaces are separated by the heat insulating bulkhead (70). In

order to prevent exhaust gases from ﬂowing into the balance channel (2) and

the turbine side reservoir, barrier air is fed from the compressor to the turbine

rotor labyrinth seal via channel X.

The rotor (20) has easily accessible bearings (32, 38) at both ends,

which are supported in the casing with vibration damping spring elements.

Turbocharger construction 183

184 Pressure Charging

Either roller or plain bearings are used, but for the most common construction

using roller bearings a closed loop lubrication system with an oil pump directly

driven from the rotor is used (47, 48). These pumps are ﬁtted with oil centri-

fuges to separate out the dirt in the lubricating oil. The bearing covers are each

ﬁtted with an oil ﬁlter, an oil drain opening and an oil gauge glass. On models

with plain bearings, where the quantity of oil required is large, these are fed

from the main engine lubricating oil system.

A key feature of this VTR design was the modular construction to match a

wide range of diesel engine types. The separate modules of the turbocharger,

as shown in Figure 7.8, are the silencer ﬁlter (80), the air inlet casing (82), the

compressor housing (74), the gas outlet casing (60), the outlet casing feet (680)

and the gas inlet casing (50). The ﬁxing screws are placed so that the radial

position of all the separate casings can be arranged in any position relative to

the other.

Compressor wheels for MAN Diesel turbochargers are almost exclusively

milled from an aluminium-forged blank. Cast aluminium alloys were once

used as well but superseded as mechanical strength requirements increased.

Titanium alloys may also be exploited for applications imposing high air

outlet temperatures at high pressure ratios. Radial turbine wheels are pur-

chased as casting blanks of a nickel base alloy. ‘Built’ rotors are used only

for axial turbine wheels (the cast or forged individual blades are retained in

the disc by a ﬁr-tree foot). For smaller diameter ranges, progress in casting

technology has yielded axial turbine wheel castings incorporating the blades

(integral wheels).

Figure 7.8 Cross-section of Brown Boveri type VTr 501 turbocharger

TurBoCharger PerForManCe and deVeloPMenTs

A typical pressure–volume curve for a turbocharger is shown in Figure 7.9,

with reference to the operating characteristics of a two-stroke diesel engine.

The operating curve runs almost parallel to the surge line. Every value of the

engine output corresponds to a point on the curve, and this point, in turn, corre-

sponds to a particular turbocharger speed which is derived automatically. There

is, consequently, no need for any system of turbocharger speed control.

If, for example, at a speciﬁc turbocharger speed, the charge air pressure is

lower than normal, the conclusion may be drawn that the compressor is con-

taminated. By spraying a certain amount of water into the compressor, the

deposit can be removed, provided it has not already become too hard. A special

duct for water injection is arranged on ABB’s VTR turbochargers.

The operating curve and the behaviour of the turbocharger may vary,

depending on whether constant pressure or pulse turbocharging is employed

and also on how any volumetric (i.e. mechanically driven) compressors and

turbochargers are connected together. By observing changes in behaviour, it

is usually possible to deduce causes and prescribe the measures necessary to

rectify the problems.

Turbocharger design and charging system reﬁnements continue to under-

write advances in diesel engine fuel economy and speciﬁc output. Turbocharger

developments also contribute to improved engine reliability (lower gas and

Turbocharger performance and developments 185

4 6 8 10 12 14 16 18

Air intake volume m

3.0

2.0

1.5

1.0

Pressure ratio (for 15 °C air intake temp.)

+ 50%

+ 75%

+ 100%

+ 110%

Engine load

9500

8800

8200

7500

6500

5000

Turbocharger speed (rev/min)

2.5

Figure 7.9 Typical pressure–volume curve

186 Pressure Charging

component temperatures) and operational ﬂexibility (e.g., better part-load per-

formance). Rising engine mean effective pressure ratings call for an increase

in charge air pressures, with turbocharger efﬁciencies remaining at least at the

same level (Figure 7.10).

Two-stroke engines typically operate at bmeps of up to 19.5 bar, with tur-

bocharger efﬁciencies of just over 68 per cent and pressure ratios of up to 4.

Raising the bmep to 21 bar calls for an increase in turbocharger efﬁciency to

70 per cent and a pressure ratio rise to 4.2. A higher bmep requires a higher

pressure ratio to keep the air excess ratio at the same level; at the same time,

the turbocharger efﬁciency must also be increased to maintain the air mass

ﬂow and cylinder mass purity at desired values.

A valuable development in recent years in matching engines and turbo-

chargers was the ‘low port’ concept, which took advantage of the availability

of new turbochargers with higher efﬁciencies (around 68 per cent instead of the

previous 65 per cent). As the required scavenge air mass ﬂow could be deliv-

ered using less exhaust gas energy, the height of the scavenge air ports could be

reduced and the opening of the exhaust valve delayed—lengthening the effec-

tive working expansion stroke and thus reducing speciﬁc fuel consumption. An

important consideration was that this beneﬁt could be obtained without raising

key component temperatures. At the same time, the exhaust temperature was

not reduced too low, so retaining the potential for waste heat recovery from an

economizer.

Suitable turbocharger layouts are achieved through a well-established pro-

cedure. An appropriate turbocharger speciﬁcation laid down during the engine

design process is based on the calculation of certain parameters, notably

scavenge air pressure and full-load and part-load turbocharger efﬁciencies.

The desired safety margin against surging—usually 15 per cent—is also spec-

iﬁed. In subsequent engine shop tests, the scavenge air pressure is adjusted,

based on checking compliance with deﬁned requirements such as the level of

Figure 7.10 Trends in engine bmep with respect to turbocharger overall eciency

and pressure ratio

turbocharger efﬁciency, efﬁciency curve shape and related tolerances. If neces-

sary, the turbine nozzle ring is changed.

Tests are also made to check the engine’s stability against turbocharging

surging, which can be achieved in one of the three ways: the exhaust back pres-

sure can be increased at full engine power; fuel can be cut off to one engine

cylinder close to the turbocharger while the engine is running at part load, and

the dynamometer load can be suddenly reduced by around 25 per cent from

full load.

While satisfying increasing demands for higher efﬁciencies and charge air

pressures, turbocharger designers must also address the requirement for com-

pact conﬁgurations, reliability and minimal maintenance: qualities fostered by

designs with fewer components. All servicing and overhaul work should ide-

ally match the intervals allowed for the major engine components. In addition,

safe engine operation dictates a maximum width of the compressor map in

order to avoid compressor surging (liable to occur due to the changes in oper-

ating conditions inﬂuenced by hull fouling, which has the same effect as speed

reduced at constant torque).

Demand for higher charge air pressures necessitates higher circumferential

velocities of the compressor wheel and raises the sound level; this in turn has

promoted the development of more effective intake silencers for turbochargers.

Turbocharger development also seeks to raise the speciﬁc ﬂow rate for a given

frame size to achieve as compact a unit as possible for cost efﬁciency, space

saving and ease of installation.

Turbocharger designers have pursued higher pressure ratios, overall ther-

modynamic efﬁciencies and speciﬁc volumetric ﬂow rates, as well as lower

noise levels, from more simple and compact modular conﬁgurations with

uncooled casings, inboard plain bearings lubricated directly from the engine

oil circuit and signiﬁcantly fewer parts than earlier generations. Reﬁnements

have sought easier servicing and enhanced reliability and durability.

Certain trends in engine parameters can be identiﬁed that will dictate

matching developments in turbocharger technology, Wärtsilä highlighting

higher speciﬁc power output to achieve less weight per kilowatt, higher engine

efﬁciency, lower exhaust emissions, improved engine reliability and longer

times between overhauls and lower manufacturing costs. These will call for

further turbocharger development addressing higher scavenge air pressures,

higher turbocharger efﬁciency, more compact dimensions and competitive

prices.

Bmep ratings increased from around 17 bar for two-stroke engines deliv-

ered in 1990 to 19 bar in 2000. Ratings of 22 bar are not unreasonable to

expect by 2010. Such a level dictates an increased pressure ratio and efﬁciency

from turbochargers because it is important that certain boundary conditions

are maintained to sustain engine reliability, curb smoke levels during part-load

running and retain current speciﬁc fuel consumption rates. The wider use of

variable turbine geometry (VTG), two-stage turbocharging and turbochargers

with auxiliary drives can also be anticipated.

Turbocharger performance and developments 187

188 Pressure Charging

TurBoCharging and eMissions/Miller CyCle

High turbocharger efﬁciency contributes to reduced carbon dioxide emis-

sions by improving the engine efﬁciency. The same turbocharger, combined

with a ‘smart’ turbocharging system that is able to guarantee an optimum air/

fuel ratio under all conditions, would contribute even more to the control of

soot emissions.

ABB Turbo Systems explains that the role that turbocharging can play in

reducing NOx emissions is not easily recognized. This is because NOx is pro-

duced during combustion at very high temperatures—something which can

hardly be inﬂuenced by changes in the turbocharging system since the ﬂame

temperature depends on local conditions in the cylinder and not on the glo-

bal mean air/fuel ratio. Nevertheless, an improvement can be achieved through

a joint development programme involving both the turbocharging system

and the engine, aimed at reducing the temperatures of the working cycle in

the cylinders.

The ﬁrst idea was turbocooling in which the charge air is cooled in a proc-

ess that makes use of a special turbocharger. If the pre-compressed air is fur-

ther compressed in a second-stage compressor, then cooled and expanded

through a turbine, very low temperatures can be obtained at the cylinder inlet.

ABB Turbo Systems reports that ﬁrst evaluations revealed that the available

turbocharger efﬁciencies for this process were not high enough for reasonable

engine efﬁciencies.

The Miller cycle promises much better results. The idea is similar to that on

which turbocooling is based. The charge air is compressed to a pressure higher

than that needed for the engine cycle, but ﬁlling of the cylinders is reduced by

suitable timing of the inlet valve. Thus the expansion of the air and the conse-

quent cooling take place in the cylinders and not in a turbine. The Miller cycle

was initially used to increase the power density of some engines (see Niigata

engines, Chapter 30). Reducing the temperature of the charge allows the power

of a given engine to be increased without making any major changes to the cyl-

inder unit. When the temperature is lower at the beginning of the cycle, the air

density is increased without a change in pressure (the mechanical limit of the

engine is shifted to a higher power). At the same time, the thermal load limit

shifts due to the lower mean temperatures of the cycle.

Promising results were obtained on an engine in which the Miller cycle

was used to reduce the cycle temperatures at constant power for a reduction

in NOx formation during combustion: a 10 per cent reduction at full load was

achieved, while fuel consumption was improved by around 1 per cent. This was

mainly due to the fact that with the Miller cycle—at the same cylinder pressure

level—the heat losses are reduced due to the air/fuel ratio being slightly higher,

and the temperatures lower.

The signiﬁcantly higher boost pressure for the same engine output with the

same air/fuel ratio has mitigated against widespread use of the Miller cycle.

If the maximum achievable boost pressure by the turbocharger is too low in