Simulation study of acceleration and crah-stop manoeuvres for ship machinery sytems

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Simulation study of

Acceleration and crash-stop manoeuvers

for ship machinery systems

graduation study

K.H.E. Deleroi student no.: 177105

February 95

OEMO report no.: 95/09

Technische Universiteit Delft

Faculteit der WerIctuiglcunde en Maritieme Techniek Vakgroep: Maritieme Techniek

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Foreword

This graduation study is written in the department EAT of the company Motoren-und Turbinen-Union Friedrichshafen GmbH. The author is a mechanical engineering student at the Delft University of Technology and specializing in marine engineering.

I want to thank especially Prof ft. J. Klein Woud from the Delft University of Technology that he gave me the opportunity to do my graduation work at MTU in Friedrichshafen.

Special thanks go to Dipl.-Ing. H. Klotz, the head of the department,

Dr.-Ing. H. Zellbeck and Dr.-Ing. V.T. Do for their support and help on this subject.

My final thanks go to all colleagues of the department EAT, who have helped me with this study.

Friedrichshafen, February 95

Klaus RE. Deleroi

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Table of contents

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72 . 5., . .

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Summary

MTU diesel engines of the series 396 are very often used on fast ships, like pleasure yachts, fast ferries and patrol boats. Besides the stationary performance also the dynamics of the ship and machinery are of interest for the owner, which means acceleration and crash stop manoeuvres In order to be able to provide this information MTU is in need of a simulation program. A

simulation program of a diesel engine, called MOMOS, already exists. This is extended with a subroutine in which the ship and propeller are modelled, so that also ship manoeuvres can be

simulated. But due to the complex set up, a simulation takes a lot of computer time. For quick

calculations and parameter studies, a more simple and quicker simulation program is needed. The purpose of this study is to set up a simple model of the diesel engine, gearbox, ship and propeller, and to write a computer simulation program for it. After a verification of the simulation

results a parameter study of a crash stop manoeuvre is carried out. In this study the dynamics of

an existing ship propelled by a 396-series diesel engine is taken as basis.

The simulation program is set up modular and consists of four modules.

governor and actuator

diesel engine gearbox and clutch propeller and ship

K-0(-06c12(2.

The main part is the diwl engine module, in which a number of simplifications are made. The diesel engine ispodulate4 with the use of three stationary performance maps. The dynamics of the model consists of the turbocharger model. Its speed is calculated out of the integral of the difference of the actual and the set point of the charge air pressure, and used for determination of

the transient engine torque.

A comparison of the simulation results with measurements of a load acceptance test of a 396-series diesel engine shows that the simulation program produces usable results. Additionally a simulation is carried out for an acceleration and crash stop manoeuvre and the results are

compared with those simulated with the more accurate simulation program. The course of the

engine and propeller speed and engine torque as well as the ship speed and stopping distance show a satisfactory correspondence.

In the parameter study the influence of the following parameters on the dynamics of the ship and machinery during a crash stop are investigated.

delay time between clutch disengagement and reverse engagement set speed increase just before engagement

clutch characteristic ship speed

ship mass

ship resistance curve propeller lay-out

page:v 4.

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Summary

The speed drop of the diesel engine and the released friction heat of the clutch are the main limiting factors for a crash stop manoeuvre.

The minimum diesel engine speed should not undergo the limit of nm = 400 [l/min, otherwise the

engine will stall. The friction heat of a clutch has a defined maximum, which is not easy to measure. Therefore a slipping time of tslip = 2 [s] is defined as limit. For higher values the clutch gets damaged and becomes unusable.

A delay time between disengagement of the clutch for forward mode and the engagement of the

clutch of astern mode is used to decrease the ship speed. For a lower ship speed the propeller

shaft speed is also lower, so that the speed drop of the diesel engine during engagement is reduced.

In order to shorten the crash-stop time and distance the delay time has to be reduced. This can be

achieved with a set speed increase of the diesel engine just before the engagement or with an adjustment of the clutch characteristic. A set speed increase results in a higher kinetic energy of the diesel engine, so that engagement without stalling is possible at higher propeller speeds. A disadvantage is the increase of the friction heat of the clutch, so that the set speed increase should not be chosen too high. The clutch characteristic can also be changed, which means the gradient

of the transmitted torque. A low gradient results in a low speed drop, but simultaneously the

friction heat in the clutch increases.

The influence of specific ship parameters is also regarded. A reduction of the basic ship speed has

the expected positive effect. The speed drop and the friction heat are reduced, so that it is possible

to engage the clutch for the astern mode with a short delay time.

A decrease of the ship mass or an increase of the resistance curve results also in an improved stopping performance, reduced speed drop and slipping time.

The propeller lay-out is only varied regarding the pitch ratio. This has consequences for the propeller diameter and reduction factor of the gearing, which results in a changed inertia of the propulsion system. The changed propellers show a great influence on the speed drop of the diesel

engine, due to the changed inertia.

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Chapter 1: Introduction

Chapter I: Introduction

High speed turbocharged diesel engines are used for a great deal on high speed ships, like fast

ferries, patrol boats or pleasure yachts. Commonly the lay-out of the machinery on board of those

ships is done on basis of stationary performance diagrams of the diesel engine and the

speed/power curve of the vessel. But for high speed ships the dynamic behaviour of the ship and

machinery during acceleration and stopping manoeuvres are also of great interest. In order to

advice clients properly and to optimise the lay-out of the machinery, MTU wants to be able to calculate and simulate the dynamic behaviour of a ship, propeller and machinery system MTU is in possession of a simulation program for the calculation of the dynamics of a diesel

engine, called MOMOS. This computer program consists of a detailed diesel engine model. The thermodynamic gas process in each cylinder is calculated for every degree of the crankshaft. This model is extended with a ship and propeller model, so that also a ship propulsion system can be simulated. Because of the complex set up of the diesel engine model a simulation takes a lot of

computer time. In order to be able to carry out quick simulations and parameter studies, a more

simple and quicker computer program is needed.

The purpose of this study is therefore to set up a simple model of a system consisting of a ship,

propeller, gearbox and diesel engine and to write a simulation program, using the computer language Fortran. After the verification with the reality a parameter study on the basis of a MTU

16 V 396 series diesel engine and an existing ship should be carried out. The influence on the dynamics of the following parameters should be investigated.

delay time between clutch disengagement and reverse engagement

set speed increase just before engagement

clutch characteristic ship's speed

ship's mass

ship's resistance curve

propeller lay-out

The contents of this report are as follows. In chapter 2 the lay-out of the model is described. After that the results of a load acceptance simulation are compared with measurement data in chapter 3.

Additionally the results of an acceleration and crash-stop simulation are verified with simulation

results of the more complex simulation program MOMOS. In chapter 4 the results of a parameter

study are presented, which shows the influence of the ship's speed, ship's mass, ship's resistance and the propeller lay-out on the dynamic behaviour of the system ship and machinery.

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Chapter 2: Description of the model

Chapter 2: Description of the model for the simulation program

PACS

The developed simulation program is called PACS (Program for Acceleration and Crash-stop

Simulation). The underlying model has a modular set up, in order to have a clear structure of the

computer program. In this chapter the set up is described, starting with the overall system (2.1). The model is divided into four modules, the governor and actuator (2.2), the diesel engine (2.3), the clutch with gearbox (2.4) and the ship with the propeller (2.5).

2.1 Model of the ship and propulsion arrangement

The model of the system ship, machinery and propeller consists of four separate modules, which have the following names:

module 1: governor and actuator module2: diesel engine

module3: clutch and gearbox

module41 ship and propeller

The principle lay-out of the model is shown as a block diagram in figure 2.1. The direct input is

the position of the control lever Lp [%]. Modulel calculates out of the value of Lp, the charge air

pressure pii [bar] and the speed of the engine nm [rev/min] the position of the actuator-rod xr [mm]. The actuator-rod position is the input signal into the fuel injection pump and is a degree for the amount of injected fuel. As a second variable the control parameters of the clutch staf [-] and staa [-] are defined, depending on the control lever position.

The diesel engine itself is modelled in module 2, which has as input signals the actuator-rod position xr and the engine speed nm. Module 2 calculates from these parameters the delivered engine torque Mm [Nm] and the actual charge air pressure pH.

The delivered engine torque Mm and the absorbed propeller torque Q [Nm] are the input signals into module 3, which consists of the multiple disc clutch and the gear unit. Additionally the

control parameters of the clutch staf and staa are used to calculate the speed of the propeller shaft

and engine shaft.

In module 4 the ship and propeller are modelled. As input only the speed of the propeller is needed. The absorbed torque Q of the propeller as well as the speed of the ship vs [m/s] and the sailed distance s [m] are calculated.

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Chapter 2: Description of the model Lp Ph him governor & xr actuator moduloI staLstaa module 2 ship & propeller module 4

tY.1

The dynamics of the whole system are described by differential equations. These equations, which are defined in each module, are solved in an extra integration module at the end of each

integration step. This integration module chosen out of a FORTRAN library [NAG, 1993]. This routine solves the defined first-order differential equations using a Runge-Kutta-Merson method with a variable step width.

2.2 Governor and actuator

As already stated in paragraph 2.1, module 1 consists mainly of the governor and actuator. Additional the static and dynamic limitation of the actuator-rod is included. Besides the control parameters for the clutch are determined as a function of the position of the control lever. In figure 2.2 the block diagram of module 1 is shown.

figure: 2.1 Block diagram of the simulation program PACS

him

Want:

Lp staf, , staa

tun

module 1: governor and actuator

11 governor block no.: 6 LDA Xrra2 Xr staf staa Xr 0.

figure: 2.2 Block diagram of governor and actuator model

page: 3

u

N

t

block no.: 5 block no,: 7 blockno :4'

Xrs

DBR MIN actuator

Min clutch & gearbox

module 3

block no.: block no.: 1 Lp Met An Ian

j

diesel-engine

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The user defined input variable is the control lever position Lp, like it is in reality. From the bridge the speed of the vessel is controlled by the control lever. The speed of the ship is a function of the diesel engine revolutions, which means that the control lever controls in reality the set point of the engine speed nset [rev/min]. In block no. 1 the set point of the diesel engine speed is determined. The function from which defines the relation between the set point of the diesel engine speed nset and the lever position Lp is shown in figure 2.3.

2500 2000 1500

t

1000 u' co) 50 -100 -50 0 50 100 Lp [-A

forward clutch engaged

astern clutch engaged

Lp [%]

50 100

figure 2.4 Clutch status as a function of the lever position

This common variable status can occupy three different values, status = +1 means forward clutch

engaged, status = -1 means astern clutch engaged and status = 0 means both clutches are open

page: 4

figure: 2.3 Set point of the engine speed as a function of the lever position

Until Lp = ±5 [%] the engine speed is kept idle and is after that direct proportional with the engine speed.

The position of the control lever Lp is also needed for the determination of the status of the clutch

staff-] and staa [-]. These control parameters determine if the clutch is engaged or disengaged. In this model two clutches are used, one for the forward gear and one for the astern gear. Staf is the control parameter for the forward clutch and staa for the astern clutch.

At a control lever position of Lp = 0 to ±51%] both clutches are disengaged and at Lp > +5 [%] the clutch for the forward gear is engaged and at Lp < -5 IN the clutch for the astern gear is

engaged. In figure 2.4 the relation between the control lever position and the common clutch variable status [-] is shown.

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This variable is now converted into a status set point for the forward and astern clutch, as shown In table 2.1. FCS [-] means forward clutch set point and ACS [-] means astern clutch set point

STAF/STAA = I ::

STAF/STAA =2:

STAF/STAA = 3 :: STAF/STAA = 4 STAF/STAA = 5 : STAF/STAA =6 :.

table 2.1 Definition of the state set points for the clutches

With these set points the control parameter staf and staa.of the clutches can be determined. Staf

and staa can contain the following values:

TN Tv

clutch is disengaged

waiting time, clutch is still disengaged clutch is slipping towards engagement clutch is engaged

waiting time, clutch is still engaged clutch is slipping towards disengagement

If staf or staa has a value larger than one, the other parameter is kept on value 1, in order to make sure, that not both clutches can engage at the same time. The conversion from the set points to

the control parameters is described in paragraph 2.4.

For MTU 396-series diesel engines, which are regarded in this study, commonly a

BARBER-COLMAN governor and actuator are used. The governor works electronically and is fitted with a PM-regulator circuit (proportional, integrational and differential) for getting a quick reaction and

stable control of the diesel engine speed for load changes. The input value in the governor is the difference between the set point and the actual 'diesel engine revolutions.

An = nse, (2.1)

The governor calculates from An [rev/min] the electronical set point of the actuator position

xrs [mg. The differential equation for the PD' controlcircuit looks as follows:

=Klan

+ if An dt +T,

(2.2)

In block no. 5 the actuator-rod position is limited according to the DBR-limitation (Drehzahl Begrenzung) . This is a stationary restriction, depending on the diesel engine speed. The purpose' is to secure the diesel engine of overspeed and overtemperature. Also it is secured that the turbocharger does not get into the surge area. The output of this block is called xni [mm].

gain factor of the governor

: reset time of the governor : rate time of the governor

C\Aun

Led

gic

page: 5

'

status j

FCS ACS description

' +1 I 1 0 forward gear engaged

0 1 0 0 both clutches disengaged

' -1 i 0 1 astern gear engaged

:

K [mm s] [s] [s]

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Block no. 6 is another limitation block for the actuator-rod position, the so called LDA-limitation (Ladeluftdruckabhangige Rillungsbegrenzung). The maximum actuator rod position xrres2 [mm] depends on the actual charge air pressure. This is a dynamic fuel rack restriction, which depends on the available combustion air in the cylinder. As feedback signal for the combustion air the

charge air pressure is taken. This restriction is used for restraining themio)emissions. This

restriction is not used for the 396-series diesel engines, but for the universality of the model a possibility to use this limitation is included. The 396-series diesel engines have another device with the same effect, but situated in the control unit. There the gradient of the set speed nset is limited. Both limitations, the DBR and the LDA, are in reality programmed in the MTU engine control system (ECS).

Block no.7 is a so called minimum block, which takes the variable with the smallest value as

output. The output is the actual governor output signal xrs.

The BARBER-COLMAN actuator is basically a simple proportional working electromagnet with a sliding armature, where the electromagnetic force is proportional to the input current.

Proportional to the current of the coil in the electromagnet the armature slides on anti friction

bearings against a restoring spring, whereby a movement without hysterese loss is achieved. This linear movement is changed into a rotative shaft motion, which is the input for the fuel injection pump. A sketch of such a actuator is shown in figure 2.5.

page: 6

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figure 2.6 Block diagram of diesel engine model

page: 7

Because the actuator and the attached linkage consist of masses and /springs the dynamic

behaviour can be described with a differential equation of a higher pOwii, which is derived from measurements. It is often sufficient to model the actuator as time-lag of quadratic order including the derivative of the input signal xis.

+ 2 -Ds CO 0 kr +CO 20 xr = Ks -CO 20 + Ks 20 (2.3)

K5= 1.0 [-] gain factor Ds = 2.0 [-] damping factor coo 100.0 [1/s] eigenfrequency Ts = 0.001 [s] time constant

In this differential equation the variable x the set point and the value xr is the actual

actuator-rod position. The used equation and its parameters are taken from [Grzegorzek, 1993]. This

differential equation is only an approximation of the reality. It is complicated to model the actuator completely, because of its complex behaviour.

2.3 Diesel engine

The diesel engine is modelled in module 2, which calculates out of the actuator-rod position xr and the actual diesel engine speed nm the delivered engine torque Mm and the actual charge air pressure pll [bar]. The diesel engine is mainly defined by using three separate performance maps and one integration block for the turbocharger speed calculation. This model has the advantage that it can be used for all kinds of diesel engines, if the needed performance maps are available. In figure 2.6 the block diagram of the engine model is shown.

nm

no 8

block no. 11'

module 2: diesel engine

block na: 9

blockna: 10

14 0.4444

pme.1

blakno.: 14 blcckno.: 15

Pme.iva = f (pil) blockno.: II Apt C nteh PA Mm 4 4..44 Vs 627t limitation biodc no,: 13 = is

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The inputs in this engine model are the actuator rod position xr and the engine speed nm

These inputs signals are used in block no. 8 in order to calculate the stationary mean effective pressure pmes [bar] . The mean effective pressure is a general comparison parameter of diesel

engines and is defined as:

5 2 - 7C

Pme = Mm

VH ZZ

: parameter for type of diesel engine,

6= 2 for a four stroke engine 5 = 1 for a two stroke engine

VH _[m3] : volume of one cylinder zz _[-] : number of cylinders Pine [bar] : mean effective pressure

The calculated mean effective pressure is a stationary value, which the engine will deliver, when

the engine operation has reached the steady state. The used performance map is set up by the test

department, by measuring the stationary engine torque as a function of the fuel rack position and the engine speed. The original performance map cannot be published, but the used digitised performance map is included in appendix 3. The engine torque is after that recalculated into the mean effective pressure following formula 2.4.

In block no. 9 the stationary mean effective pressure pmes and the engine speed rim are used to calculate the stationary charge air pressure pus [bar]. The charge air pressure is also laid down in a performance map as a function of the engine speed and the mean effective pressure, which is also set up by measurements of the test department. The digitised version can be found in appendix 3.

In block no. 10 the same diagram is used, but now the input signals are the engine speed nm and the transient mean effective pressure pmet [bar]. Because both input signals are transient therefore the output signal pil is also a transient value.

The dynamics of the diesel engine in this simple engine model is expressed by the integration

block for the turbocharger speed determination. When the actuator-rod position is changed to higher fuel injection, the engine cannot produce the required torque at once, because of the charge air defect. In order to increase the charge air pressure the compressor has to accelerate.

The charge air pressure is assumed to be directly proportional to the turbocharger acceleration. The speed of the turbocharger can then be calculated by:

ntd, = C

pn dt (2.5)

ntch [rev/min] : turbocharger revolutions

[l/(min 2 bar)]: parameter for the turbocharger acceleration

The factor C in this formula must be adjusted, so that the turbocharger acceleration fits with the real turbocharger behaviour, which is done in chapter 3.

This is only an approximated model of the turbocharger. A better way would be the use of turbine and compressor performance maps, but this would make the model too complex.

page: 8

H;

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With the known turbocharger revolutions Nett and the actual engine speed rim the transient mean effective pressure pmet can be calculated, which is done in block no. 12. The relation between ntch and nm is also laid down in a stationary engine performance map, of which the used inputfile is

shown in appendix 3.

A natural aspirating diesel engine can produce its maximum mean effective pressure at any time.

There is always enough combustion air in the cylinder to burn the injected fuel completely. Turbocharged diesel engines however show a different behaviour. At maximum load the inlet air has a higher pressure than the atmospherical pressure. That means that the amount of combustion air is higher and that the diesel engine is able to burn a larger amount of fuel and produce more

engine torque, or in other words a higher mean effective pressure. But the required charge air pressure has to be generated by the turbocharger, so that themaximum mean effective pressure

cannot be produce at any time. The turbocharger has first to accelerate to its maximum speed in

order to produce the required charge air pressure.

But the mean effective pressure of a corresponding diesel engine without a turbocharger can also be generated by the turbocharged engine. This parameter is called in this report pmena [bar]. At each load the surplus of combustion air is high enough to guarantee a complete combustion proces.

The surplus of combustion air can be expressed by:

total used amount of air per kg fuel theoretically used amount of air per kg fuel

[-] : excess-air factor

Betz [1985] defined a combustion efficiency rum H, which can be set up against the excess- air

factor X, what is shown in figure 2.7. At a X >1.3 the combustion will be complete, so that the combustion efficiency is neon = 1. For a lower excess-air factor the injected fuel cannot burn complete, because of the missing oxygen, so the combustion efficiency is lower than one.

X. =

figure 2.7 Combustion efficiency as function of the excess-air factor

page: 9 11 corn

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In figure 2.8 a block is drawn which corresponds with the maximum mean effective pressure of a turbocharged diesel. The hatched block corresponds with the mean effective pressure of the same diesel engine but without turbocharger.

ca. 21 [bar]

ca. 7 [bar]

maximal available pme

pine of a natural

aspirating engine

figure 2.8 Presentation of the mean effective pressure of a diesel engine

In the engine model, shown in figure 2.6 three mean effective pressures are calculated, which are summed up below.

Prn

= f(xr,

pmc, = f(ntch , nm

P. =

f(p11)

The first two variables are already discussed before and the last one is the already mentioned mean effective pressure for a natural aspirating diesel engine.

During steady state the stationary and the transient mean effective pressure are equal. Only in dynamic conditions these variables are different. Possible combinations are illustrated in figure 2.9. Block no. 14 decides, which mean effective pressure of the three is taken as the output value.

LIZ Oef

/

A4.4 447 G(22-e-c4.. pe

'7

...c -1

(2.6)

r

Z .4.

74_,/,

pme = pmes pine = pmena pme = pmet pine = pmet

figure 2.9 Possible combinations of the mean effective pressure

fi

page: 10 0 (1) (2) pries (3) pmet (4)

pars pmet _pars

pmena _pmena prnena proem 77;1 pmes _pmet PAW' pmet

7 0-e

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Chapter 2. Description of the model

In case one the stationary and the transient mean effective pressure are lower than the pmena, but

the stationary pmea has a higher value than the transient one That means that the actuator-rod is set on a position, so that more fuel is injected and the diesel engine can accelerate Because the

stationary pines is underneath the pmena limit, the pines can be delivered at once, because enough combustion air is available. The transient pmet value will correspond with pines in the next step.. In case two the pules is higher than the pmena, but the transient one is lower than the pmena. Here

the engine also accelerates, but the fuel injection has such a high rate that there occurs a

combustion air defect to burn all fuel. The available mean effective pressure is the pmena. The diesel engine can generate only the corresponding engine torque until the turbocharger ha&

reached the required speed for the needed charge air pressure.

The third and forth combination happen when accelerating or decelerating the diesel engine which is already operating on a load higher than the pmena limit. In this case another phenomenon

occurs, which is illustrated in figure 2.10.

proono.dt)

puMto+dt) Pa3a(10) = Pne(tO)

X(to) = 2

plat. = 7 [brill

to :time when actuator-rod position is increased :. time when diesel engine operates in steady state

t

figure 2.10 Mean effective pressure response to an increase of fuel injection

When the diesel engine operates in steady state and the transient and stationary mean effective pressure are both higher than the pmena value, for instance pme = pmet = 12 [bar]. The

actuator-rod position is suddenly increased, which results in a larger amount of injected fuel The stationary

mean effective will then lie at pines = 16 [bar". This mean effective pressure cannot be reached at once. But the combustion excess-air factor is at the stationary point still higher than A, = 1.3, so that a part of the more injected fuel can be burnt at once. The mean effective pressure will therefore jump immediately to a higher pressure, for instance pmet = 14 [bar]. But to burn the

whole amount of the injected fuel more combustion air is needed. Therefore the turbocharger has to accelerate in order to produce the needed charge air pressure. Due to the higher energy in the exhaust gas, the turbine is able to accelerate. This proces takes a couple of seconds until the

required speed and charge air pressure is reached. At that time the stationary mean effective

pressure will be equal with the transient mean effective pressure again, which is shown in figure 2.10

In order to take this effect into consideration in this engine model, the direct available mean effective pressure pm is set to be a function of the charge air pressure, see figure 2.11. The

page: 11 pn(max) 2,1 [bar] Por [bad; 16 bar pay* 14!bar 12 bar = to

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course of the line is determined by drawing a line through two known points. The first one is the point of the charge air pressure of pll = 1.0 [bar]. At this pressure the engine will react like a natural aspirating engine, the corresponding mean effective pressure is assumed to be

pmena = 7.0 [bar]. The second point is the origin of this graph. At p11= 0 [bar] no combustion can take place that means pme = 0.0 [bar]. For a higher charge air pressure than Pu = 1.0 [bar] the direct available mean effective pressure pmena increases proportionally with the charge air pressure. A pMe [bar] 25 20 pine of natural

aspirating diesel engine

15 10 5 2 3 _{p11 [bar]} page: 12 (2.7) figure 2.11 Direct available mean effective pressure

depending on the charge air pressure

In block no. 13 the direct available pmena is defined, following the curve of figure 2.11. After the limitation block no. 14, where the decision is made which mean effective pressure is taken as output value, block no. 15 calculates the diesel engine torque following the equation 2.7.

MM =

(V'4

- z

PITIC

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2.4 Clutch and gearbox

The clutch and gearbox are modelled in module3, its block diagram is shown in figure 2.12. With the input signals of the delivered diesel engine torque Mm, the absorbed propeller torque Qi [Nm]

and the control parameters for the clutches ,staf and staa the diesel engine speed nm [rev/min] and the propeller speed_xi are calculated.

staf _staf block no.: 16 Mc = f(t) or Mc = Mm Node no.: 19 Tin block no.: 20 MB =f(vs) shaft brake

module 3: clutch and gearbox

Mc block Mc no Is Q. block no. 1" 1 J/, block no.: 18

figure 2.12 Block diagram of the clutch and gearbox model

The clutch and the gearbox are thought to be arranged, as schematically shown in the next figure.

engine shaft

Mc FC

AC

Mc

ditch reduction gear lull fongsrd/ a:urn forward larern

11p propeller shaft

figure 2.13 Model oldie gearbox with integrated clutches

page: 13 staa staa.

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The engine torque works through the input shaft on either the forward or the astern clutch. Both

clutches can transmit only a certain torque, which is called the transmitted torque Mc [Nm]. The clutch is built up of multiple discs fixed on the input shaft and discs fixed on the outputshaft.

These discs are pressed together hydraulically, when the clutch is closed. The transmitted torque depends on the hydraulic force, which presses the disks of the input and output shaft together. Behind the clutches the reduction gear units are placed, which are represented by the reduction

factors iF and ip ,for the forward and astern gear units. The output shaft of the gearbox is direct

coupled with the propeller shaft.

and 12 are so called integration points, where the two differential equations, one for the

propeller shaft speed and one for the engine shaft speed are based on. Both equations are only

solved when both shafts operate independently, that means both clutches are open or one is slipping.

The differential equations as below:

(

MCF iF 4- MCA MB 1 Thr 12: :nm

=1[21

t'.7r

M M" MCA

dt EI 2 [-] iA _[-]

MCF [Nm]

MCA [Nm]

MB [Nm] Q [Nm] Titr [-] page: 14 dt

: reduction factor of the forward gear : reduction factor of the astern gear

: transmitted torque of forward clutch : transmitted torque of astern clutch : torque of the shaft brake

: absorbed propeller torque : transmission efficiency

These equation have to be valid for every thinkable operation. Therefore the signs of the torque's

and reduction factors are defined as follows:

: positive.

negative. Mc-F : positive.

MA : positive.

MB : positive while propeller shaft rotating direction is positive. negative while propeller shaft rotating direction is negative.

Mm : positive.

positive while forward operation. negative while astern operation.

22t

_EJ1

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Chapter 2: Description of the model diesel engine forward clutch astern clutch forward JFG gear

/

iF astern gear AG propeller np =

n = f

M 1 A

2 n

EJ,

page: 15 dt dt (2.9) figure 2.14 Definition of the used moments of inertia

The moments of inertia are summarised in the following manner:

ZJI = JP ±-/Fa +JAG IJ2 = JM + JCI

JP [kg m2] moment of inertia of the propeller and propeller shaft

Jm [kg m2]: moment of inertia of the diesel engine

JFG [kg m2] moment of inertia of the forward gear, plus output side of forward clutch

JAG [kg m2] : moment of inertia of the astern gear, plus output side of astern clutch

JcI [kg ni?] : moment of inertia of the input side of both clutches.

If one clutch is engaged then only the differential equation of the integration point II is solved and the engine speed is calculated out of the propeller speed and reduction factor. The differential equation looks then as it is shown in the formula underneath, depending which clutch is engaged.

1\41,4*iF

(Z)

1

2-7c D3

or (2.10)

(21)

The moments of inertia are now summarised in a different manner. If the forward clutch is closed the total inertia is:

EJ3 =JP +JFG +JAG ±(iF)2 Vivf +JO] (2.11) Or if the astern clutch is engaged:

EJ3 = Jp+JFG +JAG ±(1)02 +JC11 (2.12)

The diesel engine speed can then be calculated with the reduction factor of the gear unit. If the forward gear is engaged the diesel engine speed is:

= iF np (2.13)

Or if the astern gear is engaged.

nm =iA (2.14)

The decision if two differential equation have to be solved or only one depends on the status of the clutch. That is controlled by the clutch control parameter staf (status forward clutch) and staa

(status astern clutch).

There are 6 different state's of a clutch possible STAF/STAA =1 : STAF/STAA =2 : STAF/STAA =3 : STAF/STAA =4 : STAF/STAA =5: STAF/STAA =6: clutch is disengaged

waiting time, clutch is still disengaged clutch is slipping towards engagement clutch is engaged

waiting time, clutch is still engaged clutch is slipping towards disengagement

Staf /staa = 1 is a stable condition, that means the regarded clutch stays open. These status values correspond with the set point FCS/BCS = 0. When the set point is set to one (FCS = 1 or

BCS = 1) then the staf or staa values are set to 2.

Stagstaa = 2 means the clutch is still disengaged, until a delay time is over; then the staf/staa value is switched over to status 3. The delay time before engagement has to be defined by the user in the input file. During this time the engine revolutions can be increased, if required.

Staf/staa = 3 means that the clutch closes, this results in a transmitted torque between the input and output shaft. The course of the transmitted torque is not easy to measure. It can be derived

from the oil pressure of the clutch hydraulic system. The transmittable torque is a function of the friction between the input and output discs. The friction itself depends on the position of the discs to each other, which is controlled by the hydraulic oil pressure.

The curve of the derived transmitted torque is as shown in figure 2.15. It is not a linear curve, but

with sections of different gradients. This is in order to let the diesel engine increase its load gradually

(22)

page: 17'

figure 2.15 Transmitted torque of the clutch. as a function of the time

When both shafts are synchronised, the clutch status is switched to staf/staa = 4, which means

that it is engaged and both shaft speeds are correlated with each other.

A clutch can only transmit a maximum torque, Mc(max). If the delivered or absorbed torque get too high, then the clutch will start slipping again. The program controls the transmitted torque, if Q or Mm are too high the status is returned to staf/staa = 3. Otherwise the clutch stays closed until a signal that the clutch has to disengage. Staf/staa =4 corresponds with the set points

FCS/BCS = 1.

The clutch status is set to Staf/staa = 5 when the set points are set to zero (FCS = Cl or BCS = 0); that means the clutch has to disengage. After a user defined delay time or when the diesel engine reaches a certain value the clutch status is set to staf /staa = 6. The delay time can be used in crash-stop, manoeuvres to set the actuator-rod in idling position.

,Staf/staa = 6 means that the clutch disengages and the transmitted torque decreases rapidly. The transmitted torque is read from a file. From this moment on both shafts operate independently and

two differential equations have to be solved. If the torque reaches the value zero, then the clutch is fully disengaged and the clutch status is set to staf/staa = 1, what corresponds with the set point FCS or BCS = 0.

Only in the clutch status for staf/stab = 4 or 5 the single differential equation has to be solved. It is

not possible to dis- or engage the forward and the astern clutches at the same time. When a crash-stop is simulated the process of disengagement of the forward clutch has to be fulfilled fully before the astern clutch can start its engagement procedure.

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2.5 Ship and propeller

With the knowledge of the propeller rotational speed np, the ship speed vs [kn] and its position

s [m] plus the absorbed torque Q canbe calculated. Therefore the open water propeller diagram in a four quadrant presentation, has to be known, in which thrust and torquecoefficients are laid

down as a function of the water velocity and rotational speed of the propeller. Additionally the ship resistance curve or at least the resistance value at the nominal ship speed is needed.

The block diagram of module 4 is shown in the next figure.

module 4: ship and propeller

blodc no.,21 Q 11P I T,Q = f(v.,no) VI blodc no.:22 (1-t) biodc no.:23 1-w Toff blodc no.:24 ttlo + Eclw No& na.: 25 Kvt) 1-11-1. V V a tan(p) a U 0.7 7I D

va [m/s] : water intake velocity at the propeller [m] : propeller diameter

[us]

: propeller speed

page: 18

(2.15) figure 2.16 Block diagram of the ship and propeller

In block no. 20 the delivered thrust and the absorbed propeller torque are calculated out of the

input variables propeller speed np and water intake velocity at the propeller va [m/s]. Therefore the 4 quadrant curve of the propeller is required, in which the non-dimensional thrust and torque factors, CT [-] and CQ [-], are defined as a function of the angle of attack of the water 13 [-]. This angle is defined, as shown in formula 2.15 and figure 2.17. The variable u [m/s] is the

circumferential speed of the propeller blade at a height of 0.7 times the diameter of the propeller, D [m].

(24)

figure 2.17 Definition of the angle of attack 13 of a propeller blade

With the calculated angle 0 the thrust and torque coefficients CT and CQ can be read out of the 4-quadrant open water diagram of the propeller. An example of such a diagram is shown in

figure 3.6 in chapter 3..

The propeller thrust and torque can therefore be calculated.

T= CT dva,24-(0.7-

-OP-1Y)

2 4 (2.16) \ .7( Q = CQ-(-1-p [v2 ±(O.7

..n

D)3] 2 a 4 T [N] : propeller thrust Q [Nm] : propeller torque

p [kg/m3] : density of the water

The non-dimensional coefficients CT and CQ are determined by thrust and torque measurements of model propellers. These measurements are done by dragging the propeller through the water, while driving the propeller by an electromotor. For every combination of dragging speed and revolutions of the propeller shaft the thrust and the torque are measured and recalculated in the coefficients. If the propeller is mounted under a ship the water-flow around the propeller is influenced by the ship's hull. The water-flow and the delivered thrust are reduced. The thrust reduction is expressed by the thrust deduction factor t [-]. The effective thrust is determined in block no. 21.

Ter = (1 t)-T

(2.17)

Teff [N] : effective propeller thrust : thrust deduction

page: 19

t: [-J

(25)

Also the water intake velocity at the propeller is reduced and is determined out of the ship speed in block no. 21.

v. = (1 w) vs

(2.18)

vs [kn] : ship speed

[-] : wake fraction

This reduction factor in this case is called the wake fraction. The wake fraction and the thrust deduction are commonly a function of the ship speed.

The difference between the effective thrust of the propeller and the resistance of the ship is the acceleration force of the vessel. If the force is divided by the mass of the ship and the surrounding

water, the ship acceleration can be determined. This equation is integrated one time to get the

vessel's speed and a second time to get the sailed distance.

Ter R

a =

ms + mu,

s=.1.v,-dt=1.1(

Tf R

ef

Ms +My,/

dt dt

as [rn/s2] : acceleration of the ship

ms [kg] : mass of the ship

m,, [kg] : mass of the surrounding water

Dien [1973] uses as assessment for the mass of the surrounding water that itis 10% of the ship mass.

= 0.1- ms (2.20)

In block no. 24 the resistance of the ship is calculated, which depends mainly on the ship's speed. In the computer program the user has two possibilities to calculate the ship resistance.

The first method follows the assumption, that the resistance of a mono-hull ship is in first

approximation proportional to the ship's speed to the second power. Only the resistance of the

ship at the nominal speed has to be known. The program draws a parabolic function through this nominal point and through the origin. With this curve the resistance of the vessel can be

interpolated and extrapolated for every ship speed.

A better and more general method is to use a measured resistance curve of the ship. While the first method can only be used roughly for normal mono-hull ships, the second method provides the possibility to simulate other ship types, for example catamarans.

page:* 20

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Chapter 3: Verification of the model

A verification of the simulation program PACS is carried out, in order to adjust some specific

engine parameters and to see if the results correspond with reality.

First of all the model of the diesel engine, which is the most important part of the program is checked (3.2). This is done by comparing the results of a simulation run with measurements of a

load acceptance test run.

After that a simulation of an acceleration (3.3) and crash-stop manoeuvre (3.4) are carried out.

The results are compared with results of a more complex engine simulation program, which is

called MOMOS (Modulare Motor Simulation).

3.1 Comparison with a load acceptance test run

In order to check the dynamics of the diesel engine and to adjust specific engine parameters a load acceptance test is simulated. The simulation results are compared with measurements of a

generator load acceptance test run of a 396-series diesel engine The basic data of the test run are [5]:

diesel engine: continuous power: speed droop: alternator: load increase: 16 V 396 TE 34 1675 (kW] at 1500 [rev/min] 4% AEG, DKB 4564104 25%-50%-75%-100%

In the simulation program PACS some changes have to be made in order to simulate a load acceptance.

The propeller and ship module has to replaced by a module for the generator load. It consists

only of a file, in which the absorbed torque values are listed as a function of the time.

The generator load cannot be directly measured, it must be determined out of the alternator electric power. Assuming a constant alternator efficiency of rio = 0.95 the engine torque can be calculated otiiiifilTe-erectrapower and thethe actual alternator revolutions. The absorbed

generator torque is illustrated as a function of the time in figure 3.1.

(27)

Chapter 3: Verification of the model load A [%] 100 75

50

-/5 _ 0 page: 22 1800 1600 1400 1200

t

1000 800 o 600 400 200 0 A 5 10 IS 20 25 30 time[4] 35

figure 3.1 Load increase of the acceptance test run

In the governor and actuator module a speed droop function block has to be added. The engine speed of a generator set is a function of the load. In this case a speed droop of 4% is

used. This means that the diesel engine speed at zero load is 4% higher than at full speed. The load-engine speed relation is illustrated in figure 3.2.

figure 3.2 Load-engine speed relation

(28)

Because the load of the engine cannot be measured easily during operation the actuator

position is taken as a degree of the load. The set speed of the diesel engine can be determined following formula 3.1.

1

i(xo

xr)

rise, =

+P

nseto (3.1)

kXr0 Xr.idle)

[-] speed droop

[min] : actuator position at nominal speed

icue : actuator position at idling speed

Another change has to be made concerning the determination of the clutch status.

Generator-sets do not have a switchable clutch between alternator and diesel engine. Therefore the clutch stays engaged during the simulation.

After the described changes module 1 looks then as follows:

block no.: 2

stet', ,erne

module 1: governor and actuator

pa

Xr

staf staa

Xr

Additionally the required performance maps of the diesel engine type 16V 396 TE 34 have to digitised. The original and the digitised perfromance maps are given in the appendix 3.

With these changes simulations of load acceptance tests can be carried out. For the simulation a

step width of H = 0.01 [s] is used, which brought satisfying results. It is not advisable to use a

larger step width, because then strong inaccuracies appear.

page: 23

figure 3.3 Adjusted block diagram of module?. governor and actuator

The reduction factors of the gear units have to set to unity, because the regarded

generator-set consists of no gear.

block no block no.: 3

governor 11/11/14

bloc& co 5 block no 7 Wk no; 4 Xrs BR to MIN actuator block no.: 6 LDA = 1.01 = 1.0

(29)

The comparison with measurement data is also used to calibrate the parameter C, which is a

degree for the acceleration of the turbocharger. For a load increase it is important that the turbocharger accelerates quick enough, so that the needed charge air pressure can be provided.

The factor C is adjusted in a way, so that the course of the diesel engine speed, charge air pressure and engine torque correspond optimally with the measurement data. The simulation shows the best results for:

C = 850 [1/(min2*bar)]

The governor reaction has also to be adjusted, so that the diesel engine reacts as quick as in

reality. In this model only a PI governor is used, the differation part is switched off. The main

parameter is the gain factor K, which determines the output value. The speed of a reaction can be

adjusted with the reset time TN. The following parameters are chosen:

xr, =

KlAn+i--i-Andt]

(3.2)

K = 0.85 [mm s] : gain factor of the governor

TN = 2.0

[s] : reset time of the governor

In figure 3.4 the diesel engine revolutions rim, the absorbed engine torque Mm and the charge air

pressure pii are set up against the time. In figure 3.5 the calculated and measured actuator-rod

position are presented as a function of the time.

Although the developed simulation program is based on stationary engine performance maps and

a simple model of the turbocharger, the results show a good correspondance with the

measurement data. Especially the revolutions and the engine torque are nearly equal, which are the output variables with the most significance for an acceleration and crash-stop simulation. The course of the charge air pressure shows some deviation, mainly in the high load region. It

shows the course is too steep just after a load increase. But for steady state the charge air

pressure seems to correspond with the real values. This deviation can be explained by the simple

set up of the turbocharger model, although the factor C is adjusted optimally. For a better

turbocharger model performance maps of the compressor and the turbine should be used, but this

would make the engine model too complex.

Another reason for the deviations is a possible inaccuracy of the interpolation of the performance maps. The interpolation points are chosen with an interval of nm = 200 [rev/min] and between these points a linear interpolation is carried out, which can result in inaccuracies.

But for an acceleration and crash-stop simulation program a correct course of the engine torque and engine speed is important and the charge air pressure is only a tool for a correct engine torque

calculation.

(30)

The actuator position calculated by the simulation program shows a great deviation in comparison

with the measured values. The explanation for this difference is that the governor and the actuator

are not of the same type as used during the test run. It was not possible to change the governor and actuator, because the used type is unknown. Another reason is that the performance maps are based on a basic actuator with a range of xr = 0 -17 [mm]. But the test data show a maximum actuator position of xr = 13 [mm].

A quantitative comparison between these two signals is therefore not possible, but qualitatively it

can be concluded that the course of the actuator position of the model is correct.

(31)

Chapter 3: Verification of the model 0 7 1500 1450 DIESEL ENGINE: 16V 396 TB 34 dM/dt = 25Z-507-75Z-100Z ; dt = 8 Cs] speed droop MEASUREMENT: 1600 SIMULATION (PACS) . 1550 671e .1 ek time Cs] 16 20 24 28 32 time Cs]

Lrndc.-.

page: 26 ut (it Li CD cr cT, 0 12000 10000 8000 6000 4000 2000 OX 257. 12 SOZ 16 20 24 time Cs] 757. 28 100% 32 0 ' ' 1'8 ' 210 ' 24 ' 26 32

figure 3.4 Comparison between simulation and measurement of a load acceptance;

Engine speed. absorbed engine torque and charge air pressure as function of time

=

(32)

Chapter 3: Verification of the model 20' $6, DIESEL ENGINE: 16V 396 TB' 34 dM/dt = 251 - 50% - 75X, - 100X dt = 8 Es: speed

droop

= 4% MEASUREMENT: SIMULATION (FAGS): '0

figure 3.5 _{Comparison between simulation and measurement of}_{a load acceptance-,}

Actuator-rod position as function of time

page:. 27' 116 20 24 2a

time CC

0 0

-18 -8 6 2 0 114 12 -10 -1 I 8 32

(33)

3.2 Comparison of simulation runs between PACS and MOMOS

In the previous paragraph only the diesel engine model of the simulation program PACS is checked on its validity. In this paragraph an acceleration and a crash-stop simulation are carried

out with the whole model. The results are compared with these of a moreprecise simulation program, called MOMOS.

3.2.1 Ship and machinery data

cet-k

A good contl of the simulation results would be a comparison with test trial data of an

acceleration and crash-stop manoeuvre. But unfortunately this is not possible, because of lack of complete test trial data of a ship powered with 396-series diesel engines. Therefore it is decided to compare the results of PACS with simulation results of MOMOS. The latter simulation program can calculate the diesel engine more accurate, due to the calculation of the gas proces in each cylinder per degree of the crankshaft.

The simulation of the acceleration and crash-stop manoeuvre are both based on the patrol boat

"Neustrelitz" of the Bundesgrenzschutz. This vessel is originally fitted with two 595-series diesel engines on two shafts with fixed pitch (EP) propellers. Because this was the only example of a large vessel with a complete data set about the ship, machinery and propeller, it is taken as a basis for this comparison simulation. The 595-series diesel engines are replaced by 396-series diesel

engines.

The used ship, machinery and propeller data are listed underneath.

shin data: machine data:

page: 28

1 45 [in] Type 2 x 16V 396 TE 94

ms 350 [ton] P(max) 2240 [kW]

Ro 242 [kN] ' nm(max) 2100 [rev/min]

vso 23.2

Pall

nrrodie 600 [rev/min]

w 0.2 [-' Jm 7.257 [kg m2

t

0.102

vi

_{JR-, + JACi} 20.0 [kg m2-:

Jp 142.0 [kg m2-]

Iltr 0.97 {-]

iF / lA 2.641 [-]

table 3.1 Used ship and machinery data

The used performance maps for the 16V 396 diesel engine are given in the appendix 3. As above a PI-governor is used with the following parameters:

K = 0.85 [min s] = 2.0 [s]

The constant for the turbocharger acceleration is the same as used for the load acceptance test, its

value is:

C = 850 [1/(rnin2*bar)]

I

(34)

The simulation program MOMOS and PACS can only simulate one diesel engine. Because this ship has two propulsion diesel engines the ship's mass ins and the resistance Ro have to be divided by a factor 2. rolDeller data: page: 29 z 5 D 1.368 [m] A/A0 0.779 P/D 0.884

0

CD I (--) 2 . 0 1 . 5 1. 0 0 a 5 0 .0

-0.5

-1.0

... 1. 5 i IS i i I I I I di

-,, , A t i

-10

t, , % ''S w CO CT % 1 ,I, 1 I t f ,, , %, t I i t i 1 1 I % 1 VI

ts-,

I t s' ... .. I

-2.0

0

11111EITITIIIIIII-1-111-1-11

30 60 90

120 150 180 210 240 270 300 330 360

0 table 3.2 Used propeller data

From the propeller data the corresponding thrust and torque coefficients, CT and CQ, can be calculated as a function of the angle of attack 0. This is done with the program CTCQ, of which a description is included in appendix 5. As input only the number of propeller blades z, the area ratio A/A0 and the pitch ratio P/D is needed. In figure 3.6 the used 4 quadrant propeller diagram is shown.

(35)

The clutch for the forward and astern gear unit is a multiple disc clutch. In figure 3.7 the course of the transmitted torque versus the time is shown. The data is taken from [Ridder, 1992].

MA

[kNin] 45 40 35 30 25 20 15 10 5 7 page: 30

figure 3.7 Transmitted torque as function of time; Lolunann & Stolterfoht clutch

0.0 1.0 2.0 3.0 4.0 5.0 6.0 t [s]

(36)

3.2.2 Acceleration manoeuvre

An acceleration manoeuvre for the above ship and machinery is carried out with the Simulation

programs PACS and MOMOS. The regarded manoeuvre is carried out as follows: t = 0 [s] 'diesel engine, runs with rim =1550 [rev/min],

all 'clutches disengaged;

ship stands still; vs =0 [al]

ti= 4 Is], : signal for engagement of forward clutch ship stands still; vs =0 [kri]

t> 410

.: forward clutch engages

engine is accelerated in 30 [s]l to maximum speed

Figure 3.8 shows the engine set and actual speed plus the propeller speed as a function of the time

for the whole acceleration manoeuvre. The thick lines are the results calculated by PACS and the thin lines by MOMOS. Both simulations results of the diesel engine and propeller speed are over the whole range very similar. Notable differences occur only at the switching point of the

turbocharger, which can be recognized by the oscillations at around nm =1500 [rev/min]. At that point a second turbocharger is switched in. The diesel engine model within PACS is modelled

with performance maps, in which a discontinuity occurs at the switching point. This results of

course in deviations compared with reality. The engine model modelled in MOMOS uses two turbochargers, of which the aerodynamic and thermodynamic proces is calculated more exactly.

The most interesting interval of the acceleration manoeuvre is the moment when the engagement of the forward clutch takes place. This interval is indicated by a circle and is illustrated enlarged in figure 3.9. Both simulation results are also in the enlarged view quite similar. Only a greater overshoot over the set point can be noticed. At nm = 400 [rev/nun] a line is drawn which

represents the diesel engine speed limit. At an engine speed under this limit the diesel engine will stall. A lack of combustion air occurs, the diesel engine receives insufficient air to burn all the 'injected fuel. The converted energy is not enough to take the load on the engine. But this load still.

increases and forces the engine speed too a lower speed until standstill. The lowest possible engine speed is originally rim = 350 [rev/min]', but for security reasons the minimum engine speed is defined to be nm = 400 [rev/min] for this study. At this low engine speed it can also occur that the oil pressure gets under a critical value, which results in an emergency stop of the diesel

engine, which is released by the ECS (Engine Control System).

In figure 3.10 the path of the actuator rod position xr, mean effective pressure pm, and the charge

air prssure pu, are shown. The results of both simulation programs differ a lot concerning the

actuator position, shown in the upper plot. But that is because MOMOS uses an actuator with a greater range for the fuel rack, xr = 0 23 [mm], where PACS only calculates with xr = 0 -17 [mm] Therefore this result can only be compared qualitatively, but the actuator position calculated with PACS shows the same behaviour. At the switching point of the turbocharger a greater oscillation of the actuator-rod position occurs, due to possible deviations in the

performance diagrams or in the interpolation routine.

The mean effective pressure pm, and the charge air pressure pll show a similar course. The start and the end value are the same. Differences occur only at the switching point of the turbocharger.

page: 31

(37)

Figure 3.11 shows the ship speed and the sailed distance as function of the time. The curves calculated by PACS correspond very well with these of MOMOS. But this is not surprising, as the engine speed corresponds also very well with the MOMOS results.

It can be concluded that an acceleration manoeuvre can be simulated exactly over the whole speed region. Also at the critical point, the engagement of the clutch for the forward mode, the engine and propeller speed show a very similar course.

(38)

CD -ao0o- S'oo 1000 0

magnifiCation

in figure 3.9

ins° = 350 Ct]; RI = 242 CI1N] at vs0 = 23.2 Ckn], acceleration from vs 0 to 23.2 Ckn] in dt = 30 Cs]

0

set speed Aset engine speed rim

Propeller speed apsi

2 ₀ 0 2590 B. "fp-' 30 LIP 50' 110 20' Urns Cs] 500 MOMOS

(39)

-Lao 800 200 0 ?. 0 3,

ndlin

\ Pncs 4 v0 I

I

I I

I

/

5.0

time CIO T.0

diesel ngin*

speed limit

8:0

8.

Set speed engine speed

nm

HOMOS

PACS

Propeller speed npMi

Se

0 0 0

ms0 s 350 cj]; RO A 242 CkNJ et vs0 s 23.2 Bird

U

-nset

6.0 1000 ₆₀₀ ₄₀₀

(40)

Chapter 3: Verification of the model 30 28 26 ri 22 20 LJ 18 16 X 0 12

00

10 c 8 6 w 0 4

Ca

₂ 0 a) LI 25 E 20

00.

a) C-(_ a)C- 15 CD 7 0 10 c (I)

Cc

Ic 5 E 0 LI ms0

350 Et3. RO

242 CkN3 ot vs0 = 23.2 Ckn3

MOM OS: PACS:

1015

20 25 30 35 40 45 50 to IS 20 25 30 35 40

time Cs]

t ime Cs]

45 50

figure 3.10 Comparison of actuator-rod position, mean effective pressure and charge air pressure between PACS and MOMOS during an acceleration manoeuvre

page: 35 45 50 1 0 15 20 25 30 35 40

time Cs]

0 a, = 24 c

(41)

iChapter3: Verification of the model 25`r 2.0 '5

ms0, = 350 CO.. RD =. 242 [WC

vs0

23..2 Ekma

MOMOSr.; PACS: 0 ILO 20J 30

trne

figure 3.11 Comparison of the ship speed and sailed distance between PACS and MOMOS during an acceleration manoeuvre

page: 36 VS SOO 400 300 200 100 -40

Cs]

-0

(42)

3.2.3 Crash-stop manoeuvre

Much more interesting for the operator of a high speed vessel is the dynamic behaviour of its ship and machinery during a crash-stop, which is part of the sea trials program. For thepatrol boat

"Neustrelitz" the stopping time of the quickest possible crash-stop is measured to be

tstop 57 ID]; Butthe vessel is originally fitted with 595-series diesel engines. With PACS and MOMOS a crash-stop for this patrol boat, but now with 396-series diesel engines is simulated The crash-stop manoeuvre is carried out as follows:

t = 0 [s] : ship travels at max speed, vs = 23.2 [1c.n]

t = 5 [s] : signal for crash-stop, control lever is put from +100% to -100%.

The fuel injection is stopped and the friction torque of the diesel engine brakes the propellershaft.

When the idling speed of the engine is reached, in this case nm = 600 [rev/min], the forward clutch is disengaged. The engine stays then during a delay time of dtdeiay = 58 [s] idling and disengaged. For a crash-stop this delay time seems too long, but it is derived from data of the sea trial report. During the sea trial a minimum delay time of dtdday = 40 [s] is used, which is

increased proportionally to the power rating of the diesel engine. Although the delay time can surely be shortened this simulated crash-stop manoeuvre is only used in order to compare the results of both simulation programs.

Half a second before engaging the astern gear the diesel engine speed is increased to

nm = 800 [rev/min]. Then a signal is sent for engaging the clutch and acceleration of the diesel engine. The clutch cannot close immediately, because the oil pressure which forces the discs

together has to be built up first. This takes normally a quarter of a second. After a delay of

dt = 0.75 [s] the diesel engine is accelerated from nm = 800 [rev/min] to 2100 [rev/min] as quickly as possible.

In figure 3.12 the propeller and engine speed of both simulation programs are shown. The thin lines correspond with the results of MOMOS and the thick ones with PACS. It can be seen that

differences occur during the deceleration phase. The engine speed, calculated with MOMOS, drops more rapidly than with PACS. But the acceleration of the propeller shaft after engagement of the astern clutch shows an almost equal course. The engagement procedure of the astern clutch is indicated by a circle and is illustrated enlarged in figure 3.13.

The speed drop during engagement is the same for both simulation programs. The minimum speed, according to MOMOS, is situated before the synchronisation point. The results of PACS show that the synchronisation point is equal with the minimum speed point. These differences are only minor ones and can be accepted

In figure 3.14 the actuator position, the mean effective pressure and the charge air pressure are

presented as a function of the time. The actuator position of both simulations differ a lot, due to the use of a different actuator type. The mean effective pressure shows an almost equal course over the whole range of the crash-stop. The charge air pressure calculated with PACS shows almost the same path than the one calculated by MOMOS. Only at the deceleration phase and acceleration phase of the diesel engine notable differences occur.

In figure 3.15 the stopping performance of the regarded vessel is presented. The curves for the speed and the stopping distance correspond very well between the two simulation programs, due to the good correspondance of the engine and propeller speed.

page: 37

(43)

Chapter 3: Verification oldie model

As a conclusion it can be stated that a crash-stop manoeuvre can be simulated with the program PACS as well as with MOMOS. The stopping distance plus the engine and propeller speed show an equal course, apart from minor deviations, which can be neglected.

(44)

E; 0

ms0 = 350 Et]. RO

= 242 illts13 eat 1,90

2 23.2 Elln]

set speed

nset

engine speed

nm

Propeller speed npmi

eta

j-t1/476i 10 MOMOS PACS --_

megn f ice t ion in

f igure 3.13

.1111"__ maw__ I 20 30 40 50 60 70 80 time Co3 2500 2000 -1500 -1000 . SOO ,

/tat,

N-iFf,"?

.4111.

(45)

800 LI 600 a a 1100 1000 -200 0

ms0 = 350 CtJ, RU

= 242 CkNJ at vs0 . 23.2 CknJ

set speed

nset

engine speed

nm

Propeller speed npxi

71.0 momos PACS 11111, 11=1111 MOM OS 71.5 72.0 72.5 One go]]

-\

/

\

\I

...4,

S. 51,

_/

/

73.0 73.5 diesel on 9 in I tow t =

(46)

Chapter 3: Verification of the model 30 28 26 2L1 in 22 20 Li 18 16 14 12

00

_... 10 8 7 ₆ o 4

ca

₂ 0 (ID L..1 25 .CD E 20

00.

15 c._ c_ CD 7 10 c Ul C 'CD IL 5 E a 0

L a

o CD gi)

00 .r

0 0_ 3 -2 -, 350 CtD. RD = 242 EN

ot

vs0 = 23.2 [Kn3 MOMOS: PACS: 10 20 30 40 50 60 lb 80 90

time Cs3

10 20 30 40 SO 60 70 80 90

time

Cs:

0 0? T lb 2b I 3I0 T I 1 6I T

II

gib 40 so 70 80

t ime Cs]

page: 41

figure 3.14 _{Comparison of the actuator-rod position, mean effective pressure and charge air pressure} between PACS and MOMOS during a crash-stop manoeuvre

(47)

ms0

350 Et], RU = 242 INN: at

vs0

23.2 Ckn3

MOMOS:

PACS:

s(stop)

figure 3.15 Comparison of the ship speed and stopping distance between PACS and MOMOS during a crash-stop manoeuvre

page: 42

time Cs]

- 400 - 300 - 200 _t a3 (,)

25

-20 -15 10 5 0 LO 20 I 30 I

vs

100 0 4b sb sb 713 BO 90 = '500 0) 0. 101