Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile...

13
 http://pid.sagepub.com/ Engineering Engineers, Part D: Journal of Automobile Proceedings of the Institution of Mechanical  http://pid.sagepub.com/content/213/4/391 The online version of this article can be found at: DOI: 10.1243/0954407991526955 1999 213: 391 Proceedings of the Institution of Mechanical Engineers, Part D: Journal of Automobile Engineering Y. H. Zweiri, J. F. Whidborne and L. D. Seneviratne Dynamic simulation of a single-cylinder diesel engine including dynamometer modelling and friction Published by:  http://www.sagepublications.com On behalf of:  Institution of Mechanical Engineers can be found at: Proceedings of the Institution of Mechanical Engineers, Part D: Journal of Automobile Engineering Additional services and information for  http://pid.sagepub.com/cgi/alerts Email Alerts:   http://pid.sagepub.com/subscriptions Subscriptions: http://www.sagepub.com/journalsReprints.nav Reprints:   http://www.sagepub.com/journalsPermissions.nav Permissions:  http://pid.sagepub.com/content/213/4/391.refs.html Citations:   by guest on May 29, 2011 pid.sagepub.com Downloaded from 

Transcript of Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile...

Page 1: Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

8/4/2019 Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

http://slidepdf.com/reader/full/proceedings-of-the-institution-of-mechanical-engineers-part-d-journal-of-automobile 1/13

Page 2: Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

8/4/2019 Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

http://slidepdf.com/reader/full/proceedings-of-the-institution-of-mechanical-engineers-part-d-journal-of-automobile 2/13

Page 3: Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

8/4/2019 Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

http://slidepdf.com/reader/full/proceedings-of-the-institution-of-mechanical-engineers-part-d-journal-of-automobile 3/13

Y H ZWEIRI, J F WHIDBORNE AND L D SENEVIRATNE392

T r reciprocating inertia torque (N m)

torsional stiffness torque (N m)T Svelocity of the slider (m/s)U 

V d displacement volume (m3)

W  load on the slider (N/m)

x Cartesian coordinates in the x direction

piston displacement (m) y

h weighting coefficienti connecting rod angle (rad)

l piston pin offset (m)

Dt time step (s)

q crankshaft angular position (rad)

q1 dynamometer angular position (rad)

v dynamic viscosity of the oil (N s/m2)

x characteristic and inclination angle of the

upper of a tilted ring profile (rad)

oil density (kg/m3)z

� connecting rod angle when the piston is at

TDC (rad)

1 INTRODUCTION

The direct-injection (DI) diesel engine model has long

been established as an effective tool for studying en-

gine performance and contributing to evaluation and

new developments. Most of the work done in this area

has concentrated on steady state models for the pur-

pose of modifying engine design parameters in order

to minimize emissions and maximize power and fuel

economy of the engine. However, recent regulations

have imposed stringent emission and fuel economy

standards that cannot be addressed by a steady state

analysis of the engine. Simulation of transient engine

response is needed to predict performance and fuel

economy of diesel engines that frequently experience

rapid changes in speed and load. Hence, to contribute

towards solving this problem, the current research

work is conducted with the aim of developing a

non-linear dynamic model for direct-injection single-

cylinder diesel engines that can simulate the engine

performance under transient and steady state operat-

ing conditions.

Previous efforts in the area of engine dynamic mod-

elling can be grouped into two major categories:steady state non-linear and transient non-linear mod-

els. Examples of steady state non-linear models can be

found in references [1] t o [4], which simulate real

spark-ignition engines in order to estimate engine

torque and cylinder pressure.

Some examples of transient non-linear dynamic

models can be found in reference [5], where a model

composed of thermodynamic and dynamic constitutive

elements for a transient, multicylinder diesel engine

simulation is developed. This model utilizes a quasi-

steady thermodynamic process. A comparison of pre-

dicted and measured pressure traces during the

transient response was satisfactory overall, but also

indicated some limitations of the quasi-steady process

submodels, and so Filipi and Assanis [6] have ex-

tended the steady state diesel engine simulation to

include the prediction of instantaneous engine speed

and torque during transients.

Important aspects of engine dynamic operation are

the instantaneous torque and the cyclic nature of the

gas pressure force and the slider– crank kinematics.

Therefore, the objective of this work is to develop a

non-linear single-cylinder diesel engine model with full

transient capability explaining the relationship between

the net engine torque and the angular speed of the

crankshaft. Another aspect that plays an important

role in engine transient modelling is the evaluation of 

frictional losses, especially piston assembly friction be-

cause it is a factor that strongly affects the economy,

performance and durability of the reciprocating inter-

nal combustion engines. The model takes this pointinto consideration in dealing with the detailed analysis

of engine friction components. Another advantage of 

this model is that full dynamic dynamometer mod-

elling with step loading (to avoid the chatter effect) is

included. In addition, the piston pin offset has been

taken into consideration during the transient analysis,

and motoring analysis capability could also be imple-

mented. This paper presents the salient features of the

developed model, along with a brief description of a

SIMULINK [7] implementation. The dynamic engine

operation is illustrated by simulation results, and the

predicted engine response is validated through com-parison with measured data from two different en-

gines.

The paper is arranged as follows. Firstly, the engine

and dynamometer model, which is composed of the

engine and dynamometer dynamic model, the instanta-

neous single-cylinder engine torque model and the fric-

tion torque model, is formulated on a crank angle

basis. Next is a description of the implementation,

followed by some simulation results to show the model

behaviour and validation. Finally, there is a discussion

and some conclusions are drawn.

2 ENGINE MODELLING

2.1 Engine dynamic model

Figure 1 shows a model of the engine coupled to a

dynamometer. The following two equations describe

the dynamic system:

T i− %5

k =1

T f k −T r=J q 8 +T S+T D (1)

Proc Instn Mech Engrs Vol 213 Part D D04698 © IMechE 1999

 by guest on May 29, 2011pid.sagepub.comDownloaded from 

Page 4: Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

8/4/2019 Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

http://slidepdf.com/reader/full/proceedings-of-the-institution-of-mechanical-engineers-part-d-journal-of-automobile 4/13

DYNAMIC SIMULATION OF A SINGLE-CYLINDER DIESEL ENGINE 393

T D+T S=J 1q 8 1+ %

n

 j =1

T L j  (2)

The above equations simply state Newton’s second law

for a rotational body. The variables used for these and

all the other equations are defined in the Notation. The

indicated engine torque, T i, is generated by the conver-

sion of chemical to thermal to mechanical energy dur-

ing the combustion process. The reciprocating torque,

T r, is produced by the motion of the piston assemblyand the small end of the connecting rod. The recipro-

cating torque and the friction torque terms, k =1

5 T f k ,

are subtracted from the instantaneous indicated torque

value to produce the brake torque at the shaft. The

resistance torque,  j =1

n T L j , which is the result of exter-

nal loading imposed on the engine by the dynamome-

ter, is in addition to the dynamometer inertia. Owing to

rapid changes in the cylinder pressure and consequent

changes in the forces acting on the crank during a cycle,

the instantaneous rotational speed of the crankshaft is

unsteady during any engine cycle, even if the mean

speed is constant. The torsional stiffness torque, T S,

and damping torque, T D, which depend on the stiffness

and damping in the coupling between the engine and

dynamometer, are given by the linear relationships

T S=S (q−q1) (3)

and

T D=D(q: −q: 1) (4)

2.2 Instantaneous engine torque model

Figure 2 shows the piston–crank mechanism with ap-

proximate kinetically equivalent point masses replacing

the connecting rod. The model includes the piston pinoffset. Important geometrical parameters are the

crankshaft angular position, q, the angle of the con-

necting rod, i, the crank radius, r, which is equal to

half of the stroke, the connecting rod length, L, the

piston pin offset, l, and the connecting rod angle when

the piston is at top dead centre (TDC), �.

The relationship between the indicated gas pressure,

Pi, and the indicated torque, T i, is deterministic and is

a function of engine geometry. This relationship is

expressed as

Fig. 2 Forces and acceleration of the piston–crank mecha-

nism

T i=( pi− patm)ArG (q) (5)

where

G (q)=sin(q+i)

cos i=sin q+'1−u

ucos q (6)

and

u=1−l+r sin(q−�)

L

n2(7)

From the piston–crank geometry, the piston displace-

ment, y, is given by

 y=(r+L)2−l2− [L cos i+r cos(q−�)] (8)

where angles � and i can be expressed as

�=sin−1l

r+Land

i=sin−1l+r sin(q−�)

L(9)

2 .2 .1 Reciprocating torque, T r

This term is the torque produced by the motion of the

engine reciprocating components and is given as

T r=MrG (q) y =MrG (q)[G 1(q)q: 2+G 2(q)q

 8 ] (10)

where geometrical functions G 1(q) and G 2(q) are

G 1(q)=r!

cos(q−�)

1+(r/L) cos(q−�)

u3/2

n

−'1−u

usin(q−�)

"(11)

G 2(q)=r

sin(q−�)+'1−u

ucos(q−�)

n(12)

Fig. 1 Engine and dynamometer model

D04698 © IMechE 1999 Proc Instn Mech Engrs Vol 213 Part D

 by guest on May 29, 2011pid.sagepub.comDownloaded from 

Page 5: Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

8/4/2019 Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

http://slidepdf.com/reader/full/proceedings-of-the-institution-of-mechanical-engineers-part-d-journal-of-automobile 5/13

Y H ZWEIRI, J F WHIDBORNE AND L D SENEVIRATNE394

where M  is the mass of the piston, rings, pin and small

end of the connecting rod, and y  is the acceleration of 

the reciprocating components. The connecting rod is

treated as an equivalent mass system, the first concentric

mass is assumed to be connected to the crankpin as a big

end while the second concentric mass is attached to the

piston assembly as a small end. The forces acting on the

connecting rod, the inertia forces and the bearing forces

act at the ends of the rod. It is assumed that the big endof the connecting rod may be placed at the crankpin

centre rather than at the correct point. Thus, there are

no transverse components of the force between the ends

of the rod to bend or shear the link, and therefore the

member is in axial tension or compression.

Implementation of the instantaneous torque model

obviously requires accurate masses of the reciprocating

components in addition to the detailed engine geometry.

The piston pin is slightly offset in order to reduce engine

noise and wear during the change in direction of the

normal force on the piston at the end of compression.

2.3 Friction torque model

2 .3 .1 Piston ring assembly friction torque, T  f  1

The literature [8–10] suggests that the piston ring assem-

bly may be responsible for 50–75 per cent of the entire

engine friction. The components that contribute to

friction are: compression rings, oil control ring, piston

skirt and piston pin. The forces acting on the piston

assembly include static ring tension, the gas pressure

force and the inertia force. The piston assembly friction

is dominated by the ring friction components [11]. This

model takes into account only the hydrodynamic lubri-cation, since the friction torque is identically zero at the

top and the bottom dead centre position. The piston

assembly friction torque is expressed as

T f1=F f1rG (q) (13)

where

F f1=sgn( y;  )% F f 

RLi +F f 

SL

n(14)

where sgn( y;  ) is the signum function (i.e. the sign of 

friction force is the same as the sign of piston velocity)

defined as

sgn( y;  )=>1, y; \0

0, y; =0 (15)

−1, y; B0

The present approach is based on calculating the piston

assembly friction using a simplified model [12, 13] that

is based on hydrodynamic lubrication. The lubrication is

considered to be one-dimensional as both ring and bore

are assumed to be perfectly circular with the same centre,

in which case the clearance in the circumferential direc-

tion is constant, the ring is considered to be infinitely

long and there is no gap effect. In this case the Reynoldsequation becomes

(

(x

h3

v

( p

(x

=−6U 

(h

(x+12

(h

(t(16)

The load equation is

W =& B 

0

 p dx (17)

and the friction force is

F f =& B 

0

h

2

( p

(x+vU 

h

dx (18)

By integrating the Reynolds equation twice with

boundary conditions x=0, p= p1(t) and x=B , p=

 p2(t), the oil-film pressure is expressed as

 p=6{U −2(hl −hm)/[tan x(Dt)]}vB hm

2 K 

× 1

h2

−K +1

h2

2(K +2)−

1

K +2

n+ p1

+( p1− p2)(K +1)2(h2

2−1)

[(K +1)2−1]h2

2(19)

where

h2=h/hm and K =B  tan x

hm

(20)

Finally, from equation (18) the friction force per circum-ferential unit length is

F f 

l =

hm

2[ p1− p2(K +1)]+

1

2

tan x

+vUB 

hmK ln(K +1) (21)

where W  is obtained from equation (17).

2 .3 .2  Bearings friction torque, T  f  2 

Bearings friction contributions come from the journal

bearings and their associated seals. Journal bearings areusually designed to provide a minimum film thickness of 

about 2 mm. The journal bearings operate under hydro-

dynamic lubrication, which means a large load can be

carried by the journal bearing with low energy losses

under normal operating conditions. Following work

done by Rezeka and Henein [14], the friction torque, T f2,

in the bearing is expressed as

T f2=hADb

2( pi− patm)

cos q q: 

(22)

Proc Instn Mech Engrs Vol 213 Part D D04698 © IMechE 1999

 by guest on May 29, 2011pid.sagepub.comDownloaded from 

Page 6: Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

8/4/2019 Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

http://slidepdf.com/reader/full/proceedings-of-the-institution-of-mechanical-engineers-part-d-journal-of-automobile 6/13

DYNAMIC SIMULATION OF A SINGLE-CYLINDER DIESEL ENGINE 395

Fig. 3 Transient diesel engine model representation using SIMULINK

2 .3 .3  Val 6e train friction torque, T  f  3 

The valve train carries high loads over the entire speed

range of the engine. Loads acting on the valve train at

lower speeds are due primarily to the spring forces,

while at higher speeds the inertia forces of the compo-

nent masses dominate. From reference [15] (p. 738), the

friction torque, T 3, in the valve train is expressed as

T 3=169.8(1−0.00127q: )nivD iv

1.75V d

2d 2r

n(23)

2 .3 .4  Pumping losses torque, T  f  4 

The pumping work is the integral of the product of the

pressure and the volume over inlet and exhaust strokes.

The work measures two effects: the first is the restric-

tions outside the cylinder, in the inlet and exhaust

systems—the air filter and intake manifold on the inlet

side and the exhaust manifold, muffler and tail pipe on

the exhaust side; the second effect is the valve flow,

which corresponds mainly to pressure losses in the inlet

and exhaust valves. The pumping losses torque is the

summation of the two effects. From reference [15] (p.

728), it is expressed as

T 4=1.0618 V d

2.28

(nivncD iv

2 )1.28nq: 1.7 (24)

2 .3 .5  Pumps friction torque, T  f  5 

The pumps are employed to circulate the oil, water and

fuel. From reference [16] (p. 246), the pumps friction

torque is expressed as

T 5=6.79×10−6zV d(2r)2.5

'D

v

nq: 2.5 (25)

The formulation obtained by the three models pre-

sented in this section leads to a set of non-linear

differential equations. These can be numerically inte-

grated to obtain the simulated engine performance.

3 MODEL IMPLEMENTATION

The simulation is created using MATLAB/SIMULINK

[7, 17]. Figure 3 shows the structure of the single-cylin-

der diesel engine SIMULINK model. The main advan-

tage of SIMULINK is its capability to represent the

entire engine model by an assemblage of interconnected

blocks. Also, it has eight variable-step solvers and six

fixed-step solvers for the integration of differential

equations, and hence the most suitable integration

method can be chosen. Input design parameters are

passed on to the blocks from an input file, but all of theoperating parameters come from the block (functions)

for the other components of the system.

4 MODEL BEHAVIOUR AND VALIDATION

In order to validate the behaviour of the engine dy-

namic model with experimental results, simulations

have been performed for two single-cylinder diesel en-

gines labelled A and B. Some geometrical specifications

for engine A are shown in Table 1.

Table 1 Engine A geometrical specifications

130 mmBore, d 80 mmCrank radius, r

Connecting rod length, L 269.3 mm15Compression ratio, c

Piston pin offset, l 1.69 mmEngine moment of inertia, J  1.4 kg m2

Dynamometer moment of inertia, J 1 0.37 kg m2

D04698 © IMechE 1999 Proc Instn Mech Engrs Vol 213 Part D

 by guest on May 29, 2011pid.sagepub.comDownloaded from 

Page 7: Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

8/4/2019 Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

http://slidepdf.com/reader/full/proceedings-of-the-institution-of-mechanical-engineers-part-d-journal-of-automobile 7/13

Y H ZWEIRI, J F WHIDBORNE AND L D SENEVIRATNE396

Fig. 4 Comparison between the predicted (—) and measured (· · ·) instantaneous speed of engine A with

dynamometer load (– – –)

Fig. 5 Fluctuation in the crankshaft angular acceleration of engine A during transient response

A comparison between predicted and measured val-

ues of the crankshaft instantaneous angular velocity

during engine transient behaviour is illustrated in Fig.

4. The measured values are taken from reference [6].

Almost no external load is imposed by the dynamome-

ter for the first two seconds, so the engine accelerates

from low idle speed and passes through the entire speed

range until it is at high idle speed. The acceleration

during the transient is shown in Fig. 5; the engine

accelerates because the net torque value is positive. As

shown in Fig. 4, between 2.4 and 3 s, the dynamometer

increases the external load in order to keep the engine

Proc Instn Mech Engrs Vol 213 Part D D04698 © IMechE 1999

 by guest on May 29, 2011pid.sagepub.comDownloaded from 

Page 8: Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

8/4/2019 Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

http://slidepdf.com/reader/full/proceedings-of-the-institution-of-mechanical-engineers-part-d-journal-of-automobile 8/13

DYNAMIC SIMULATION OF A SINGLE-CYLINDER DIESEL ENGINE 397

speed constant, hence acting as a cut-off of the fuel

pump. Finally, the external load is increased signifi-

cantly after 3 s in order to reduce the engine speed

while indicated torque remains at the same value. The

resulting net torque value is negative, so the engine

decelerates as shown in Figs 4 and 6.

The overall agreement between the measured and

predicted traces is excellent. The very small discrepan-

cies at a speed of 1900–2180 r/min are linked to using

dynamometer step loading rather than fuel pump cut-

off to avoid engine over-running, inaccuracies in the

values of the engine model parameters, changes in

Fig. 6 Fluctuation in the crankshaft angular deceleration of engine A during applied external loads

Fig. 7 Fluctuations in the indicated torque on the crankshaft of engine A during steady state low idle speed

over one engine cycle

D04698 © IMechE 1999 Proc Instn Mech Engrs Vol 213 Part D

 by guest on May 29, 2011pid.sagepub.comDownloaded from 

Page 9: Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

8/4/2019 Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

http://slidepdf.com/reader/full/proceedings-of-the-institution-of-mechanical-engineers-part-d-journal-of-automobile 9/13

Y H ZWEIRI, J F WHIDBORNE AND L D SENEVIRATNE398

Fig. 8 Fluctuations in the instantaneous crankshaft angular speed of engine A during steady state low idle

speed (from t1 to t2 represents one engine cycle)

Fig. 9 Steady state high idle crankshaft speed of engine A

friction and oil viscosity during transient process and to

the fact that the engine angular velocity fluctuations are

subject to the effect not only of engine torque but also

of the reactive forces from the engine and dynamometer

mounting in a sharp transient operation.

The instantaneous torque produced by the engine at

low idle speed over one engine cycle (two revolutions) is

shown in Fig. 7. The maximum torque value represents

the maximum pressure in the cylinder during the combus-

tion stroke. As a consequence of the huge fluctuations

Proc Instn Mech Engrs Vol 213 Part D D04698 © IMechE 1999

 by guest on May 29, 2011pid.sagepub.comDownloaded from 

Page 10: Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

8/4/2019 Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

http://slidepdf.com/reader/full/proceedings-of-the-institution-of-mechanical-engineers-part-d-journal-of-automobile 10/13

DYNAMIC SIMULATION OF A SINGLE-CYLINDER DIESEL ENGINE 399

Fig. 10 Fluctuation in the indicated torque on the crankshaft of engine A during transient response

in engine torque during the cycle, the variations in the

instantaneous crankshaft rotational speed are obvious

between t1 and t2 in Fig. 8; the sudden drop in the speed

and its subsequent increase can be linked with the

negative and positive peaks of the engine torque. The

amplitudes of the cyclic speed fluctuations tend to

increase as the mean engine speed decreases owing to the

fact that at low engine speed the cycle time is long and

the engine deceleration at the end of the compression

stroke is dominant and vice versa, as shown in Figs 8 and

9. This is a very important criterion in making compro-

mises between the flywheel size, the engine speed of 

response and the engine low idle speed limit. Owing to

the harmonic motion of the reciprocating assembly, the

relation between the phase of indicated torque and the

instantaneous engine acceleration is virtually identical, as

shown in Figs 10 and 5.

To test the dynamic model behaviour under steady

state with exerted external load, a simulation has been

performed for a low-speed single-cylinder diesel engine;

some geometrical specifications for engine B are shownin Table 2.

Figure 11 shows the predicted instantaneous angular

velocity during the starting process until steady state

angular velocity at rating torque. The starter-off speed

was about 40 rad/s. After the starter torque is turned off,

the angular velocity is accelerated by increasing input

torque to achieve rating indicated torque. The external

load torque (from the dynamometer) was increased

between 4 and 9 s in three steps (to avoid the chatter

effect) until it reached the rating load torque (maximum

engine output torque) and the steady state angular enginespeed was achieved. Figure 12 shows the comparison

between predicted and measured steady state angularengine velocity at rating torque and it is in excellentagreement.

Figure 13 shows the average values of the engine

friction components at steady state. The piston assemblyfriction torque is about 60 per cent, the crankshaft and

camshaft bearing friction torque is about 15 per cent andthe valves and pumping friction torque are nearly equal

at steady state and are about 10 per cent each. Finally,the pump losses are about 5 per cent. Figure 14 illustratesthe mean engine acceleration (averaged over each enginecycle) to describe the engine step acceleration decreases

during dynamometer loading steps and the zero averageacceleration at steady state.

5 DISCUSSION AND CONCLUSIONS

This paper presents a dynamic model for a single-

cylinder diesel engine that can simulate engine per-formance under both transient and steady state operating

Table 2 Engine B geometrical specifications

100 mmBore, d 62.5 mmCrank radius, r218.8 mmConnecting rod length, L18Compression ratio, c

Piston pin offset, l 1.75 mm1.7 kg m2Engine moment of inertia, J 

Dynamometer moment of inertia, J 1 0.37 kg m2

D04698 © IMechE 1999 Proc Instn Mech Engrs Vol 213 Part D

 by guest on May 29, 2011pid.sagepub.comDownloaded from 

Page 11: Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

8/4/2019 Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

http://slidepdf.com/reader/full/proceedings-of-the-institution-of-mechanical-engineers-part-d-journal-of-automobile 11/13

Y H ZWEIRI, J F WHIDBORNE AND L D SENEVIRATNE400

Fig. 11 Instantaneous speed of engine B from starting (—) with dynamometer load (– – –)

Fig. 12 Comparison between predicted (— ) and measured (– – –) steady state fluctuation speed of engine

B at rating torque

conditions. The model has been implemented in SIM-

ULINK. Validation has been performed for two types

of diesel engine, one for transient response and the

other for steady state. Predicted profiles of the instanta-

neous engine speeds through the transient and steady

state are in excellent agreement with measurements.

The model includes all the engine friction components,

namely the piston assembly, the crankshaft bearings,

the valve train, the pumping losses and the pumps.

Figures 4 and 12 illustrate the importance of this

Proc Instn Mech Engrs Vol 213 Part D D04698 © IMechE 1999

 by guest on May 29, 2011pid.sagepub.comDownloaded from 

Page 12: Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

8/4/2019 Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

http://slidepdf.com/reader/full/proceedings-of-the-institution-of-mechanical-engineers-part-d-journal-of-automobile 12/13

DYNAMIC SIMULATION OF A SINGLE-CYLINDER DIESEL ENGINE 401

friction during both transient response and at steady

state. The model also includes consideration of the

piston pin offset. A dynamic dynamometer model is

also included, which enables a variety of engine tests to

be carried out.

The model has been developed with the aim of 

investigating different strategies for transient fuel con-

trol. The work presented here does not include any

modelling of the thermodynamic processes within the

engine. Work is ongoing in developing such models. In

Fig. 13 Average friction torque components for engine B at steady state: (a) piston assembly; (b)

crankshaft bearings (– · –·), pumping (—), valve train (······) and pumps (– – –)

Fig. 14 Mean acceleration of engine B from starting until steady state

D04698 © IMechE 1999 Proc Instn Mech Engrs Vol 213 Part D

 by guest on May 29, 2011pid.sagepub.comDownloaded from 

Page 13: Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

8/4/2019 Proceedings of the Institution of Mechanical Engineers, Part D Journal of Automobile Engineering-1999-Zweiri-391-402

http://slidepdf.com/reader/full/proceedings-of-the-institution-of-mechanical-engineers-part-d-journal-of-automobile 13/13