Performance_Simulation_of_Sequentially_Turbocharged_Marine_Diesel_Engines_With_Applications

8/7/2019 Performance_Simulation_of_Sequentially_Turbocharged_Marine_Diesel_Engines_With_Applications

1/8

Pascal Chesse

Jean-Francois Hetet

Xavier Tauzia

Laboratory of Fluid Mechanics,

U.M.R. 6598 C.N.R.S.,

Ecole Centrale de Nantes,

BP 92101,

44321 Nantes Cedex 3, France

Philippe Roy

Bahadir Inozu

School of Naval Architecture

and Marine Engineering,

University of New Orleans,

911, Engineering Building,

New Orleans, LA 70148

Performance Simulation ofSequentially TurbochargedMarine Diesel Engines With

Applications to Compressor SurgeThis paper presents the SELENDIA code designed for the simulation of marine dieselengines. Various measured and simulated results are compared for the performance of asequentially turbocharged marine diesel engine during a switch from one to two turbo-chargers. The results show a good agreement between measured and simulated data.Surge loops that are experimentally observed in case of an anomaly are analyzed usingsimulated results. Finally, the predictive capabilities of the simulation code are utilized toinvestigate the influence of the inlet manifold volume on the engine and air chargingsystem performance with a special focus on compressor surge. S0742-47950001104-2

1 Introduction

Turbocharger matching is a rather difficult challenge for thedesign of high output diesel engines, particularly for naval appli-cations. Test-bed measurements are usually tedious and expensivedue to the size of the power plant. As a result, engine simulationcan be a valuable tool when used in parallel to actual testing forthe objective of guiding experimental investigations and improv-ing the global understanding of various aspects of the internalcombustion engine. In previous publications, we reported the de-velopment of a simulation code that was used to investigate com-pressor surge and the resulting engine operation limits 1,2. Inthis paper, a new version of this code designed for the optimiza-tion of turbocharging systems under transient conditions is pre-sented. Various measured and simulated results are shown for a

sequentially turbocharged marine diesel engine.

2 The SEMT Pielstick System Applied to the PA6 STC

Engine

S.E.M.T. Pielstick has been designing sequentially turbo-charged diesel engines since the early eighties. Commercializationstarted in 1990 for high speed and medium speed engines 3 5.The system used for the PA6-STC Sequentially TurboChargedengine is presented in this section. The engine is equipped withtwo turbochargers in parallel as shown in Fig. 1. At low andmedium loads, one turbocharger is in use. At high loads, valves Gand A are opened allowing the use of both turbochargers. Due toa better air supply, the engine operating range at low speed andhigh torque is significantly enlarged in comparison with a classicsingle stage turbocharging system.

The advantages of a sequential turbocharging system are obvi-ous at steady state. However, transient performance needs to bethoroughly investigated, especially with respect to the systemswitch from one to two turbochargers. This phase is critical due tothe large inertia of the turbochargers. The system switch, triggeredby a turbocharger speed threshold, is a two step procedure. The Gvalve located on the exhaust gas line shown in Fig. 1 is opened.

Subsequently, the turbine drives the second turbocharger whosespeed progressively increases. After a certain delay, the A valvelocated on the inlet air line is also opened and the second turbo-charger starts supplying air to the engine.

3 The Selendia Simulation Code

3.1 Main Characteristics. The first version of the SELEN-DIA simulation code was designed for the steady state perfor-mance of marine diesel engines 6. The code is based on thefilling and emptying method which conceives the engine as asuccession of control volumes such as the cylinders and manifoldsfor which mass and energy balance equations are applied 7.Despite the variations of the gas composition, the gas is assumedto be homogeneously distributed inside the entire control volume.

Furthermore, fresh air and exhaust gases are assumed to be per-fect. The First Law of Thermodynamics in open systems appliedto the various control volumes allows the determination of the gastemperature and pressure as shown in Fig. 2.

Contributed by the Internal Combustion Engine Division of T HE AMERICAN

SOCIETY OF MECHANICAL ENGINEERS for publication in the ASME JOURNAL OF

ENGINEERING FOR GAS TURBINES AND POWER. Manuscript received by the ICE

Division March 27, 2000; final revision received by the ASME Headquarters April

17, 2000. Technical Editor: D. Assanis.

Fig. 1 Sequential turbocharging system developed byS.E.M.T. Pielstick for the PA6 STC

562 Vol. 122, OCTOBER 2000 Copyright 2000 by ASME Transactions of the ASME

Downloaded 11 Dec 2009 to 199.64.72.252. Redistribution subject to ASME license or copyright; see http://www.asme.org/terms/Terms_Use.cfm


2/8

The cylinder volume as a function of the engine crank angle isderived from simple geometry considerations. The Keenan andKayes tables provide the thermodynamic properties of the exhaustgas and fresh air as a function of temperature and excess air 8.The combustion heat release and all of the parameters related tothe combustion and injection processes are modeled using twoWiebes laws associated with Gaudarts parameters 9,10. Theheat transfer at the cylinder wall is evaluated using Woschnismodel 11. Friction losses are evaluated with a modified Chenand Flynns model 12. Mass flows are derived using Barre deSaint Venants laws. The inlet and exhaust valve effective areasare either calculated based on geometry consideration or interpo-lated from test bed data. Finally, the various parameters of the air

charging system are evaluated based on the compressor and tur-bine maps provided by the turbocharger manufacturer 13.

The SELENDIA simulation code makes use of a specific simu-lation language called A.C.S.L. Advanced Continuous Simula-tion Language 14. This language provides pre-programmed in-

tegration routines among other specific simulation programmingtools as well as extensive graphic capabilities for interactive run-time sessions.

After numerous validations for the simulation of steady stateperformance, the code was modified to allow the transient re-sponse simulation of highly rated marine diesel engines as shownin Fig. 3. The governor model is a Proportional Integrator PImodel. The fuel rack position is provided by the following differ-ential equation:

dRack

dtRackn K p. d diff

dtKi .diff ,

diffNengNrequested

Neng(1)

where Kp and Ki are provided by the governors manufacturer asa function of the engine speed. In addition, the model includes

Fig. 2 Structure of the SELENDIA Code

Fig. 3 Engine block diagram for the simulation of marine diesel engine transientresponse

Journal of Engineering for Gas Turbines and Power OCTOBER 2000, Vol. 122 563



3/8

various fuel rack position limits based on engine speed and boostpressure and aimed at preventing engine over-speed and insuffi-cient excess air. The code was validated and utilized for differentprojects such as the performance simulation of marine diesel en-gines under extreme conditions 15,16.

3.2 Additional Models for Sequential Turbocharging.Sequential turbocharging is a rather new technique. Publicationsregarding this topic are extremely limited. Swain published someresults for the simulation of turbocharger switches with a specialfocus on the turbocharger performance which was modeled indetail 17. However, the engine model was extremely simplified.The study focused on sequential turbocharging system switchesassuming that the engine operation was quasi-constant.

The study presented in this paper includes an engine thermody-namic model to simulate significant variations of the engine speedand torque. In addition to the switch phases, the simulation of theturbocharger performance is based on two major criteria whichare the turbocharger speed evaluated with Newtons Second Lawand the pressure ratios derived from the First Law of Thermody-namics applied to the inlet and exhaust manifolds. Mass flows andisentropic efficiencies are obtained using a two-dimensional inter-polation of the compressor and turbine maps in matrix form.

The switch phases are triggered by a pre-determined turbo-charger speed threshold. The model includes the opening charac-teristics of the two valves mentioned in the previous section interms of timing and duration. An opening coefficient OV is

defined for each valve as the ratio between the current section andthe fully opened section, as follows:

OVS

Sf o. (2)

The turbine mass flow during the exhaust line valve valve Gopening and closing is obtained by multiplying the turbine massflow when the valve is fully opened and the opening coefficientOV as shown in Eq. 3.

dm tur

dtOVexh dm turdt

f o

. (3)

When the exhaust gas valve is open and the inlet air valve isclosed, the compressor operates with a certain speed and no mass

flow. As a result, the development of a new compressor modelwas necessary. The operating point is located in the surge area ofthe compressor map. Due to insufficient information in this regionof the compressor map, the power consumed by the compressor isderived from the following simplified equation:

WcompKcomp .NTC2 . (4)

The Kcomp coefficient is specific to the compressor being used.However, it may be extrapolated from a known value for a giventype of compressor since it seems to be proportional to the square

of the compressor wheel diameter according to our tests. Whenthe inlet air valve is being opened, the compressor power is cal-culated using an equation which combines the formula of thecompressor power under normal operating conditions given byEqs. 5 and 6 and the simplified compressor power law given byEq. 4.

Wcompdm comp

dt C poC p i

2 ToTi , (5)

where

ToTi

comp

comp1compi1/i. (6)

The compressor map is extrapolated to include negative massflows as shown in Fig. 4. The objective is to determine the effectof inadequate valve opening timings, including compressor surge18. The extrapolation is based on the modeling of the variouslosses occurring when the compressor operates beyond the adap-tation line. The relevant equations are presented in Appendix 1.

As shown in Fig. 5, the combination of the filling and emp-tying equations and those associated with the inertia of the fluidbetween the compressor and the inlet manifold allows the calcu-lation of surge loops 19,20. The mathematical model is pre-sented in Appendix 2. It provides the boost pressure and the com-pressor mass flow during surge as a function of time. Thecompressor power is evaluated using Eq. 5 when the compressormass flow is positive and Eq. 4 when the compressor mass flow is

null or negative.

4 Validation

The validation phase was performed for the SEMT PielstickPA6-STC engine for which measured data were available. Themain characteristics of the PA6-STC engine are shown in Table 1.

Figure 6 shows the measured and simulated boost pressure andturbocharger speeds for the 12PA6 STC 12 cylinders engineduring a switch from one turbocharger to two turbochargers atconstant engine speed 1010 rpm. The turbocharger switch isinduced by a slight load increase. For this specific test, the delaybetween the exhaust line valve and inlet line valve openings istwo seconds. This delay was adjusted to avoid compressor surgeby allowing a sufficient turbocharger speed at the inlet line valveopening as well as to sustain the engine air supply by maintaining

the boost pressure. Indeed, when the A valve is closed and the G

Fig. 4 Extrapolated compressor map

Fig. 5 Equivalent air charging system

Table 1 Main characteristics of the Pielstick PA6 STC engine

Number of cylinders: 12 or 16Type: Vee 60Bore: 280 mmStroke: 290 mmMCR Speed: 1050 rpmMCR Power: 325 kw/cylTurbocharging system: sequential axial turbine and

centrifugal compressor

564 Vol. 122, OCTOBER 2000 Transactions of the ASME



4/8

valve is open, the single turbocharger which supplies air to theengine no longer uses the full energy flow of the exhaust gas sincethe exhaust gas flow is shared by the two turbines. As a result, theboost pressure significantly decreases. It is thus critical to adjustthe timing of the G and A valve opening to avoid penalizing theengine performance when the engine loading process includes aturbocharger switch.

Figure 6 shows that both turbochargers reach the same speedafter 1.8 s. The OVE and OVA vertical lines correspond tothe beginning of the exhaust and inlet line valve opening, respec-tively. The boost pressure decreases from 3.5 to 2.2 bar during the

turbocharger switch with an absolute minimum of 1.8 bar twoseconds after the exhaust line valve opening. The results show asatisfactory agreement between measured and simulated data. Theerror margin for the extreme values of the turbocharger speed andboost pressure is within 8 percent. The duration of each phase ispredicted with sufficient accuracy.

Figure 7 also shows a good agreement between measured andcalculated results for the engine. The exhaust line valve openinginitially results in a pressure drop in the exhaust manifold, whichtemporarily improves the engine performance. Consequently, theengine scavenging improves, resulting in a slight increase of the

Fig. 6 Engine and air charging system performance during a 1TC2TC switch at 1010 rpm for the 12PA6STCengine

Fig. 7 Engine performance during a 1TC2TC switch at 1010 rpm for the 12PA6STC engine

Fig. 8 Air charging system performance during a 1TC2TC switch at 900 rpm for the 16PA6STC engine




5/8

engine BMEP as well as the engine speed. The actual enginespeed exceeds the requested speed and the governor orders a de-crease of the fuel rack position. However, the boost pressure dropsafter approximately one second since the air is supplied by asingle turbocharger which receives only half of the exhaust gas

flow. As a result, the inlet air flow decreases as well as the enginespeed. The rack reaches its higher limit and the governor cannotcompensate the speed decrease. After the inlet air line valve open-ing, the boost pressure augmentation allows a slight increase ofthe engine speed. However, the engine BMEP does not increasesignificantly due to the higher limit of the fuel rack position. Afterapproximately five seconds since the beginning of the switchingprocess, the fuel rack limit associated with the use of one turbo-charger is replaced with the much higher fuel rack limit associatedwith the use of two turbochargers. As a result, the fuel rack posi-tion slightly increases for a few seconds allowing the engine tostabilize at the requested speed.

The validation process was continued with the 16PA6 STC 16cylinders engine, which is equipped with bigger turbochargersthan the 12PA6 STC engine. The simulation is again at constantspeed 900 rpm. The delay between the exhaust and inlet linevalve openings is 2.7 s for this engine instead of 2 s for theprevious engine due to higher turbocharger inertia. Figure 8 showsa satisfactory agreement between measured and simulated results.The error margin on the turbocharger speeds, boost pressure andduration of the various phases remains quite low.

5 Predicted Results

The SELENDIA code was utilized to simulate the 16PA6-STCengine performance in case of an anomaly. Figure 9 shows theperformance of the air charging system when the delay betweenthe exhaust and inlet line valve openings is reduced from 2.7 s to2.2 s. In such a case, the speed of the second turbocharger isinsufficient when the inlet line valve is opened. This results incompressor surge as shown by the turbocharger speed curve and

especially the boost pressure curve for which a sudden drop isobserved, indicating a negative compressor mass flow. Figure 9shows that the simulated results are satisfactory. The results showthat there is only one surge loop whose amplitude and durationare predicted with good accuracy. The trajectory of the compres-sor operating point on the compressor map contributes to a betterunderstanding of the air charging system performance. This tra-

jectory can be plotted using the simulated inlet manifold pressureand compressor mass flow. Figure 10 shows the operating point ofthe compressor that is being started during a switch from oneturbocharger to two turbochargers. In this figure, A is the com-pressor operating point at the inlet line valve opening. The corre-sponding compressor speed is 270 revolutions per sec rps. The

compressor mass flow quickly increases to reach B. Then, theboost pressure increases at constant speed and the compressoroperating point reaches C. However, the compressor speed isinsufficient and the operating point goes to the surge area to even-tually reach D while the compressor mass flow significantly

decreases.Due to a reversed compressor flow, the inlet manifold pressure

decreases and the compressor operating point reaches A whichcorresponds to the minimum of the 270 rps speed line. Subse-quently, the compressor mass flow quickly increases to reachE. The A-B-C-D-E surge loop lasted approximately .5 s andthe compressor speed reached 280 rps. Since the compressor massflow is again positive, the inlet manifold pressure progressivelyincreases. The turbocharger speed is now sufficient and the oper-ating point is no longer in the surge area of the compressor map.The inlet manifold pressure stabilizes around F.

The SELENDIA code showed good predictive capabilities incase of anomalies. These predictive capabilities were utilized todetermine the effect of geometry modifications on the engine per-formance during a turbocharger switch. The effect of the inletmanifold volume was investigated. The inlet manifold volume ofthe test engine is 1.3 m3. Figure 11 shows the performance pre-diction of the SELENDIA code for an inlet manifold volume of .8m3, which was chosen arbitrarily.

The simulated results indicate the occurrence of two surgeloops before the stabilization of the engine operation. Figure 12shows the compressor operating point on the compressor map.

The initial operation of the compressor illustrated by the curvelabeled A corresponds to the previously described compressor

Fig. 9 Air charging system performance during a ITC2TC switch at 900 rpm in case of a reduced valve openingtiming16PA6STC engine

Fig. 10 Surge loop on the compressor map




6/8

performance when the inlet line valve is opened. When the com-pressor mass flow is again positive, the inlet manifold pressure isapproximately 1.75 bar. During the first surge loop labeled B,the compressor did not reach a sufficient speed that would preventsurge. As a result, another surge loop labeled C is observed.

The first loop corresponds to a compressor speed of 275 rps and

the second surge loop corresponds to a compressor speed of 278rps. The compressor speed after the second surge loop is 280 rps,which is sufficient to prevent a third surge loop. The compressoroperating point leaves the surge area of the compressor map andthe boost pressure stabilizes around D.

Figure 13 shows the engine and air charging system perfor-mance for a switch from one to two turbochargers with an inletmanifold volume of 2 m3. This figure shows that the turbochargerswitch does not induce surge when the inlet manifold volume is2 m3.

A larger exhaust manifold volume results in a larger inertia ofthe engine/turbocharger assembly. The amplitude of the inletmanifold pressure drop is thus smaller with a large manifold thanwith a small manifold. This may appear as a penalizing factor forthe engine performance since a higher boost pressure puts thecompressor operation closer to the surge area. However, a higherboost pressure increases the engine power. This power increaseprovides a better energy flow to the turbine which is being startedduring the one to two turbocharger switch. Consequently, the tur-bocharger speed increases faster and reaches a higher value when

the inlet line valve is opened. This significantly diminishes therisk of surge as predicted by the simulated results.

6 Conclusion

The steady state performance and transient response of highoutput marine diesel engines must be thoroughly investigated. As

Fig. 11 Engine and air charging system performance during a 1TC2TC switch at 1010 rpm for the 16PA6STC withan inlet manifold volume of .8 m3

Fig. 12 Performance of the compressor being started during a1TC2TC switch with an inlet manifold volume of .8 m3

Fig. 13 Engine and air charging system performance during a 1TC2TC switch at 1010 rpm for the 16PA6STC withan inlet manifold volume of 2 m3




7/8

shown in this paper, the SELENDIA simulation code allows acomprehensive investigation of the engine/turbocharger assemblyunder transient conditions in the case of sequentially turbochargedengines. The various results show a good agreement betweenmeasured and simulated data. As an added advantage, simulationcan be used to investigate the influence of specific engine param-eters. In parallel to test bed measurements, it may be useful for thedesign, development and optimization of turbocharged marinediesel engines for both healthy and faulty operations.

AcknowledgmentsWe would like to thank the Center dEtude des Machines Ther-

miques of SEMT Pielstick and especially M. G. Grosshans for hiscontribution.

Nomenclature

N rotation speed rpmn number of cylinders

BMEP Brake Mean Effective Pressure barW power WP cylinder pressure bar

diff difference between requested & actual speed -m mass kgS area m2

Cp specific heat at constant pressure J/kg/K efficiency - specific heat ratio -V volume m3I inertia kgm2t time s

rack fuel rack position mmK coefficients - compression rate -

OV opening coefficient -T temperature K

Subscripts

eng enginecyl cylinder

comp compressortot total

exh exhaustTC turbochargertur turbine

max maximumn nominali inlet

o outletfo full opening

Appendix I: Elements for the Extrapolation of the Con-

stant Speed Lines of a Compressor

A single equation cannot characterize the compressor pressure-flow diagram across its complete range. Figure 14 identifies thefollowing 4 distinct sections:

Section 1 is obtained by the experiment and usually providedby the manufacturer

Sections 2 and 4 may be obtained by the experiment 18 Section 3 does not seem to be accessible by the experiment

The parametric expressions of the compressors characteristicspresented in this appendix are derived from a combination of theclassical equations and those related to the main losses experi-enced in sections 1, 2, 3 and 4. The following assumptions weremade:

A compressor characteristic is unique, independently of thesteady or dynamic during surge nature of the compressoroperation.

A compressor characteristic is represented by a function thatis continuous and derivable across the complete mass flowrange.

The minimum of a compressor characteristic occurs when themass flow is equal to zero.

Section 4 can be assimilated as the parabolic curve of a pres-sure loss.

Using the various notations defined in the general nomenclatureas well as the appendix nomenclature, the compressor character-istics are modeled as follows:

Section 2: DDp2

The following equations are based on energy conservation, Eu-lers theorem and experimental data. They include the effect ofslip factor , friction and incidence losses as well as the previ-ously listed assumptions.

siDcDad a1a2 .DcKf .Dc2a 3 1DadDc 2

/1

siDc

Dad

a1

a2 .Dc

Kf .Dc2

a 3 1

Dc

Dad2

/1

where a 11(42r2

2)/ (C p.T1*)N2

a2.N

C p .T1 .1 p .l 2 .tg 2

a 322 r1

2r2

2

C p.T1N2et

Dada3 .Dp

p1/

a 1a3a2 .Dp

2

Kfa 3

1

Dad .Dp

1

Dp2

a2

2.Dp

Section 3: 0DcDp2

The formulation of this section was arbitrarily selected to linksection 2 and 4. K is fitted to ensure continuity and derivability.

K.Dc20 where:

K

1 a2

DpK f

2a3

Dad 1

Dp

1

2Dad

a1 a 2 .Dp2 Kf .Dp

2

4a3 1 Dp2Dad

2

1/1

Fig. 14 Schematic of compressor operating map




8/8

0 a1 a 2 .Dp2 Kf .Dp

2

4a3 1 Dp2Dad

2/1

K.Dp2

2

Section 4: Dc0

K.Dc20 where K p0/Dp

2

Appendix II: Equations Resulting from the Air Charg-

ing System Model Shown in Figure 5

Valve: Barre de Saint Venants equation for subsonic flows

Pipe: Newtons First LawdDc

dt

Ac

LcPcPr

Manifold: Mass and energy conservation

dPr

dt

r..Tr

rDc Tc

TrD

vdTr

dt

r.Tr2

Pr.rDc TcTrDvDcDv

Compressor:

TcT0 PcP01/

Dcf PcP0Appendix Nomenclature

Ac Pipe cross sectionDc Compressor mass flowDad Compressor mass flow at adaptationDp Compressor mass flow during surgeD

v Valve mass flow

l2 Width of the wheel outletLc Pipe lengthN Compressor speedPr Manifold pressurePc Compressor outlet pressure

P0 Compressor inlet pressurer1 Radius of the wheel inletr2 Radius of the wheel outlet

r Mayers constantTr Manifold temperatureTc Compressor outlet temperatureT0 Compressor inlet temperature

t Time2 Angle of the wheel outlet

Polytropic coefficient Compressor pressure ratio

p Pressure ratio at surge occurrence0 Pressure ratio at compressor zero mass flow1 Air density at the compressor inlet Slip factor

r Manifold volume

References

1 Chesse, P., Hetet, J. F., Tauzia, X., and Frayret, J. P., 1995, Influence of theAlteration of the Compressor Surge Line on the Operation Limit of a Marine

Turbocharged Diesel Engine, Internal Combustion Engine Division of theASMEFall Technical Conference 2427 Sept., Milwaukee.

2 Chesse, P., Tauzia, X., Hetet, J. F., Inozu, B., and Roy, P. 1996, Study bySimulation of the Behavior of a Two Stage Turbocharged System During

Surge, ASME Spring Technical Conference, Engine Simulation Cession,

Proceedings, 27-4, pp. 107113, October 2023, Fairborn OH USA.3 Herrmann, R., 1989, Sequential Turbocharging for PA6 Engine, IMechE

seminar, Sprint Rated Engines, London Nov. 28 1989.

4 Herrmann, R., 1990, Sequential Turbocharging for PA 6 Engines, ISMEKobe.

5 Grosshans, G., 1995, The Marine Version of SEMT Pielsticks New Genera-tion of Medium Speed Engines, Fall Technical Conference, Milwaukee,ASME 95, I.C.E. Vol. 25-1.

6 Hetet, J. F., Chesse, P., and Inozu, B., 1994, An ACSL Simulation for Op-timum Operation of Turbocharged Marine Diesel Engines, ASME Paper 94-

ICE-7.7 Heywood, J. B., 1988, Internal Combustion Engine Fundamentals, Mc Graw

Hill, New York.8 Keenan, J. H., and Kayes, J., 1948, Gas Tables, John Willey, New York.

9 Wiebe, 1967, Halbempirische Formel fur die Verbre nungsggeschwind-igkeit, Verlag der Akademie der Wissenchaft der UdssR Moscow.

10 Gaudart, L., 1987, Incidence des configurations de fonctionnement denavires militaires sur la suralimentation a deux etages de leurs moteurs diesel

de propulsion PA6-BTC, These de doctorat ENSM.11 Woschni, G., 1967, An Universally Applicable Equation for the Tinstanta-

neous Heat Ransfert Coefficient in the Internal Combustion Engine, SAE

Paper 670931.

12 Chen, S. K., and Flynn P. F., 1965, Development of a Single Cylinder Com-pression Ignition Research Engine, SAE Paper 650773.

13 Watson, N., and Janota, M. S., 1982, Turbocharging the Internal CombustionEngine, Mac Millan, New York.

14 Mitchell, G., A.C.S.L.: Advanced Continuous Simulation Language, Con-cord, MA, USA.

15 Inozu, B., Gervaise, H., Roy, P., and Hetet, J. F., 1995, Performance Simu-lation of Marine Diesel Engines Under Extreme Conditions. ASME Fall

Technical Conference, Milwaukee, WI.16 Inozu, B., 1996, Performance Simulation of Marine Propulsion Systems Un-

der Extreme Conditions, final report, Gulf Coast Region Maritime Technol-

ogy Center GCRMTC research project no. AMTC95-020A.17 Swain, E., 1993, Diesel Engine Transient Performance Prediction During

Sequential Turbocharging Operations, Imech 93, C465/012/93 pp. 123131.18 Chesse, P., 1995, Determination des limites dexploitation des Diesel de forte

puissance. Incidence des circuits de liaison moteur-turbocompresseur. Instabi-

lites et pompage des compresseurs, Ph.D. thesis, Universite de Nantes and

Ecole Centrale de Nantes.19 Greitzer, E. M., 1976, Surge and Rotating Stall in Axial Flow Compressors:

ITheorical Compression Model, ASME J. Eng. Power, 98, pp. 190198.20 Yano, T., and Nagata, B. I., 1971, A Study on Surging Phenomena in Diesel

Engine Air Charging System, J.S.M.E., 14, No. 70.


Documents

Performance_Simulation_of_Sequentially_Turbocharged_Marine_Diesel_Engines_With_Applications