24-10-2016, 03:01 PM
[attachment=73767]
ABSTRACT
The aim of this thesis work is to analyse the torsional vibration of a newly developed
heavy duty truck engine to decrease the engine speed both at the idle speed and the
cruising speed. New demands on fuel consumptions and CO2 emissions force the
manufacturers to develop engines operating at lower speeds. Down-speeding may
result in several problems such as torsional vibration, increased engine speed
fluctuation, difficulties to maintain boost pressure and high journal bearing loads.
Dynamic characteristics of the mounting system have been optimized to avoid
excessive deflection and bouncing of the engine. Stiffness and damping characteristics
of mountings assumed to be linear and the system is considered to be one degree of
freedom with harmonic excitation along the crankshaft axis. Moreover, the dynamic
characteristics of flywheel and input shaft have been modified to prevent excessive
torsional vibration and gear rattle in the powertrain system. Also conventional
flywheel has been compared with dual-mass flywheel in down-speeding. The results
showed that from torsional vibration point of view, downs-speeding was feasible.
Furthermore dual-mass flywheels can significantly reduce the powertrain`s deflection up to 43%.
Introduction
High efficient internal combustion engines can be reached by applying different
possible routes. Down-speeding is one of those routes that has been investigated and
presented in this thesis work.
Down-speeding means operating the engine at lower speeds by means of changes
through transmission and/or final drive ratio. It plays an important role in raising the
efficiency since the engine friction and relative heat transfer will be reduced and also
gas exchange losses will become smaller. Down-speeding is to achieve the same
power needed to run the vehicle while the engine speed is being lowered. This can be
achieved through two different ways:
First approach is to augment the boost power of the engine and second is through
changing the dynamic configuration of the engine.
The latter procedure is more complicated and demands more changes in the
configuration of the engine. Changes can be done to the slider-crank mechanism of
the engine, i.e. piston and crank. Enlarging the piston area to have a bigger surface for
combustible material to react, therefore obtain more pressure and consequently more
force acting on the piston. Eventually more torque acting on the crankshaft, also
extending crank to make the force at the crankpin have a bigger lever (moment arm)
thus making bigger torque applied to the crankshaft. The first method is accomplished
by rising up the engine torque by means of turbo chargers.
Conversely, down-speeding results in several problems such as torsional vibrations,
problems with the powertrain mounting system, high journal bearing load, and
increased engine speed fluctuation, difficulties to maintain boost pressure and NVH
(Noise, Vibration and Harshness) problems.
For a given vehicle, a reduced engine speed means that the engine is operating at
higher specific load; BMEP-Brake Mean Effective Pressure, which results in higher
efficiency and thus reduced fuel consumption. The reasons for the increased
efficiency are found in reduced engine friction, reduced relative heat transfer, less gas
exchange losses and faster combustion (in crank angle degrees).
In this thesis work vibrations as a consequence of down-speeding have been
investigated and verified to get the least noise and disturbances. Vibration abatement
and controlling the oscillations have been studied and applied through modification of
the engine mounting system characteristics and developing the transmission parts.
Methodology
2.1 Engine Dynamics
Internal combustion engines use two common combustion cycles; Clerk two-stroke
cycle and Otto four-stroke cycle. The second, which is the most common in trucks
and automobiles, takes four strokes of the piston to complete one cycle, that is, it
takes 720o of a crankshaft to complete one four-stroke cycle.
To study engine dynamics the kinematics of a typical slider-crank must first be
considered, because each piston and its cylinder along with the connecting rod and the
crank resemble a slider-crank mechanism except that the engine is “back-driven”
compared to the common slider-crank linkage which is usually driven through the
crank part by a rotary machine.
This slider-crank has motion aligned with the axis (X) which is perpendicular to the
crankshaft and passes through the main pin thus called “non-offset” slider-crank.
Total Engine Torque
The sum of the inertia torque and the gas torque constitutes the total engine torque.
Consider that the contribution of each of these depends on the speed of the engine, so
that the gas torque comparing to inertia torque is less sensitive to speed and is
dominant at low speed. Therefore at low engine speeds, the inertia torque, which is
the dominating portion of the total engine torque at high speeds, can be neglected.
2.6 Flywheel
Flywheels are parts which can store rotational energy. They display big inertia that is
resistance against any change in its state of motion. These rotational mechanical
devices are used mainly in reciprocating engines. Because the energy source is not
continuous in time, therefore utilizing these mechanical devices can significantly
reduce large oscillations in the torque-time function thereby provide continuous
energy. [3, 13]
Nowadays two types of flywheels are used in the industry:
· Conventional Flywheel
· Dual Mass Flywheel
A conventional flywheel (CF) usually consists of a single flat disk and a Dual-Mass
Flywheel (DMF) consists of two sections, primary mass and secondary mass. These
masses are connected to each other in a way so that they can move with respect to
other to a certain degree. Each type is bolted at one side to the crankshaft; in DMF the
primary mass, and the other side is bolted to the clutch; in DMF the secondary mass.
2.6.1 Advantages of DMF over CF
1. Having two different masses; two different sections, isolates vibration in the
crankshaft from passing on to the second mass to some extent. Thereby
preventing undesirable vibrations in the gear box, thereupon preventing gear
rattle
2. Immediate react to increased load amplitude
3. Due to the existence of a medium in DMF, between primary and secondary
mass; and configuration of throttle geometry DMF can be adapted to different
requirements better than CFs
Important effects of DMFs can be stated as:
1. Reduce booming noise and gear rattle by introducing torsion damper
2. Lower fuel consumption by allowing lower engine idle speed
3. Avoid engine and transmission vibration system`s resonance speed by
shifting it to below the idle speed
4. Protect transmission components by reducing engine irregularity and
reduction of primary mass compared to one integrated flywheel
Multi-cylinder Engine Designs
There are several configurations for multi-cylinder engines, namely “INLINE
Engines”, “VEE Engines”, “ROTARY Engines”, “OPPOSED Engines” and
“RADIAL Engines”. From all these arrangements “INLINE Engines” is the most
common and simplest; the engine under consideration in this thesis is also of this type.
In this arrangement as can be deduced from its name all cylinders are in a common
plane. This arrangement gives the simplest way to alleviate unbalancing and shaking
forces in the engine.
Analysis of powertrain mounting system
In the current approach on behalf of simplicity, the mounting system is considered as
a one degree of freedom rotational system, consisting of one torsional spring and an
inertia representing the mounting and the engine, respectively. However, the original
system is comprised of five mounts with three in front and two at the rear of the
engine, the mentioned approach can satisfy our goal with nearly the same precision as
the original system.
The objective which is laid out in this approach is to find appropriate stiffness for
mounts that can assure the following requirements:
· Avoid resonance at low speeds
· Stand engine weight
· Avoid engine bouncing
It is noteworthy that noise and vibration, which are adverse phenomena in the engine
system, enforce mounts to have conflicting characteristics for their best isolation
performance [1]. So that to isolate noise and forces transmitted by the engine to the
vehicle structure, lower stiffness and damping is desired; on the contrary at lower
speeds, typically idle speed, to isolate road induced vibrations, higher stiffness and
damping is required. Care should be taken that reducing the mounting stiffness below
a certain value will lead to rigid body motion in the model. This means that the
mounts cannot fulfil their very first task that is to carry the engine load and restrict the
relative motion of it, particularly bouncing [6].
In this thesis due to considering down-speeding effects on engine compartment, only
vibration is taken into account which is the dominant consequence of this approach.
The system under consideration consists of only one mount and the engine, which is
assumed that behaviour of the mount dynamic characteristics, stiffness and damping,
are linear; also presumed that the mount is massless.
Torsional vibrations analysis of powertrain
The purpose of analysing torsional vibrations of the powertrain is to get the least noise
transferred to the cabin, protect transmission components by reducing engine
irregularity and prevent gear rattle in the gear box.
Analysis of the powertrain is performed for two configurations:
1. Conventional Flywheel (CF)
2. Dual Mass Flywheel (DMF)
For the sake of simplicity the modelling system is considered as 2-DOF and 3-DOF,
respectively. However the complete powertrain model is more complicated than the
considered types; due to insignificant difference, it is acceptable to use the mentioned
models. The complete model can be shown as below:
Comparison
6.1 Conventional Flywheel
6.1.1 Inertia 1
Figure 4.5 demonstrates that configuration 3 has the lowest deflection among all
configurations. It also shows that configuration 2 has the highest deflection. By
studying the results one could deduce that the deflection at each configuration follow
the same pattern and decrease with increasing engine speed (frequency). This is
because the system stiffness and damping has been held constant during different
engine frequencies.
6.1.2 Inertia 2
The deflection for the second inertia shows a completely different manner in
comparison to the first inertia. Here configuration 3 has the highest deflection
configurations 1 and 2 almost follow the same decreasing pattern with increasing
engine frequency, but here configuration 2 has the lowest deflection till 18 Hz of the
engine frequency. From 22 Hz to 30 Hz all configurations nearly have the same
deflection and only differ by a few thousandth of a degree.
6.1.3 Torque
The steady state amplitude of the input shaft`s torque has the same pattern as figure
4.5. Again configuration 3 has the lowest torque and configuration 2 has the highest;
and also the torque decrease by increasing the engine frequency.
As mentioned because deflection of the second inertia (gear box) has the first priority
in designing the torsional system under consideration; the goal is to reduce
oscillations in the gear box and subsequently avoid gear rattle. Therefore
configurations 1 and 2 are the best choices but considering the torque amplitude
configuration 2 is the desired one overall.
6.2 Dual-Mass Flywheel
6.2.1 Inertia 1
The steady state amplitude of deflection at inertia 1 in DMF system has quite the same
pattern as inertia 1 in CF system. Here the second configuration has the highest
deflection till 14 Hz. In this system, configuration 1 has the lowest deflection till 10
Hz but from 14 Hz configuration 3 becomes the lowest.
6.2.2 Inertia 2
At the first 4 engine frequencies (7 to 10 Hz) configuration 3 is the one that have the
lowest deflection. At this time a sudden decrease in the slope can be seen from 10 Hz
to 14 Hz which is due to the existence of the damping.
6.2.3 Inertia 3
Similar to the second inertia configuration 3 has the lowest deflection but this time for
the first 5 engine frequencies. Again a sudden decrease is visible in the deflection
diagram at 10 Hz.
Torque
In the DMF system steady state torque amplitude at configuration 1 has the lowest
value and configuration 2 in the middle of configuration 1 and 3. Also a sudden
decrease in amplitude is obvious similar to the change in deflection part of inertia 2
and 3.
Overall configuration 3 is the one chosen. However torque amplitude of the third
configuration is not the lowest one but its amplitude of deflection is lower than the
other configurations at inertia 2 and 3 and as was mentioned previously decreasing
oscillations at the gear box is the utmost priority.
Notice that all of the sudden decreases in the slope of the indicated diagrams and also
irregularities in the decreasing pattern of the deflection and torque amplitude may be
due to the linear assumption of dynamic characteristics in the system; clearly it shows
that to get the best effect and higher efficiency non-linear characteristics should be
taken for stiffness and damping
Conclusion and further research
recommendations
This thesis investigates whether it is possible to lower the engine speed (downspeeding)
in order to reduce fuel consumption therefore less CO2 emissions. It is seen
that DMF shows better results and facilitate the motion of inertia. The comparison of
deflection at the gear box shows that there is a substantial improvement in DMFs,
which are about 43% lower deflection and consequently less noise and gear rattle.
The data is acquired by assuming the mounting system as 1-DOF system, which has
only one cushion instead of 5. Furthermore the torsional system has been analysed
only through the engine and the gear box and the rest of the powertrain has been
neglected. These assumptions give satisfying results for studying torsional system due
to the fact that practically the torque amplitude after gear box (prop-shaft) is zero.
Thus it will not affect the vibration of the system significantly. In order to get more
accurate results it may suggest using non-linear mounting characteristics and consider
the complete model of both the powertrain and the engine mounts.
As a final point it is concluded that the presented method gives satisfying results for
studying the vibration of the powertrain system.
ABSTRACT
The aim of this thesis work is to analyse the torsional vibration of a newly developed
heavy duty truck engine to decrease the engine speed both at the idle speed and the
cruising speed. New demands on fuel consumptions and CO2 emissions force the
manufacturers to develop engines operating at lower speeds. Down-speeding may
result in several problems such as torsional vibration, increased engine speed
fluctuation, difficulties to maintain boost pressure and high journal bearing loads.
Dynamic characteristics of the mounting system have been optimized to avoid
excessive deflection and bouncing of the engine. Stiffness and damping characteristics
of mountings assumed to be linear and the system is considered to be one degree of
freedom with harmonic excitation along the crankshaft axis. Moreover, the dynamic
characteristics of flywheel and input shaft have been modified to prevent excessive
torsional vibration and gear rattle in the powertrain system. Also conventional
flywheel has been compared with dual-mass flywheel in down-speeding. The results
showed that from torsional vibration point of view, downs-speeding was feasible.
Furthermore dual-mass flywheels can significantly reduce the powertrain`s deflection up to 43%.
Introduction
High efficient internal combustion engines can be reached by applying different
possible routes. Down-speeding is one of those routes that has been investigated and
presented in this thesis work.
Down-speeding means operating the engine at lower speeds by means of changes
through transmission and/or final drive ratio. It plays an important role in raising the
efficiency since the engine friction and relative heat transfer will be reduced and also
gas exchange losses will become smaller. Down-speeding is to achieve the same
power needed to run the vehicle while the engine speed is being lowered. This can be
achieved through two different ways:
First approach is to augment the boost power of the engine and second is through
changing the dynamic configuration of the engine.
The latter procedure is more complicated and demands more changes in the
configuration of the engine. Changes can be done to the slider-crank mechanism of
the engine, i.e. piston and crank. Enlarging the piston area to have a bigger surface for
combustible material to react, therefore obtain more pressure and consequently more
force acting on the piston. Eventually more torque acting on the crankshaft, also
extending crank to make the force at the crankpin have a bigger lever (moment arm)
thus making bigger torque applied to the crankshaft. The first method is accomplished
by rising up the engine torque by means of turbo chargers.
Conversely, down-speeding results in several problems such as torsional vibrations,
problems with the powertrain mounting system, high journal bearing load, and
increased engine speed fluctuation, difficulties to maintain boost pressure and NVH
(Noise, Vibration and Harshness) problems.
For a given vehicle, a reduced engine speed means that the engine is operating at
higher specific load; BMEP-Brake Mean Effective Pressure, which results in higher
efficiency and thus reduced fuel consumption. The reasons for the increased
efficiency are found in reduced engine friction, reduced relative heat transfer, less gas
exchange losses and faster combustion (in crank angle degrees).
In this thesis work vibrations as a consequence of down-speeding have been
investigated and verified to get the least noise and disturbances. Vibration abatement
and controlling the oscillations have been studied and applied through modification of
the engine mounting system characteristics and developing the transmission parts.
Methodology
2.1 Engine Dynamics
Internal combustion engines use two common combustion cycles; Clerk two-stroke
cycle and Otto four-stroke cycle. The second, which is the most common in trucks
and automobiles, takes four strokes of the piston to complete one cycle, that is, it
takes 720o of a crankshaft to complete one four-stroke cycle.
To study engine dynamics the kinematics of a typical slider-crank must first be
considered, because each piston and its cylinder along with the connecting rod and the
crank resemble a slider-crank mechanism except that the engine is “back-driven”
compared to the common slider-crank linkage which is usually driven through the
crank part by a rotary machine.
This slider-crank has motion aligned with the axis (X) which is perpendicular to the
crankshaft and passes through the main pin thus called “non-offset” slider-crank.
Total Engine Torque
The sum of the inertia torque and the gas torque constitutes the total engine torque.
Consider that the contribution of each of these depends on the speed of the engine, so
that the gas torque comparing to inertia torque is less sensitive to speed and is
dominant at low speed. Therefore at low engine speeds, the inertia torque, which is
the dominating portion of the total engine torque at high speeds, can be neglected.
2.6 Flywheel
Flywheels are parts which can store rotational energy. They display big inertia that is
resistance against any change in its state of motion. These rotational mechanical
devices are used mainly in reciprocating engines. Because the energy source is not
continuous in time, therefore utilizing these mechanical devices can significantly
reduce large oscillations in the torque-time function thereby provide continuous
energy. [3, 13]
Nowadays two types of flywheels are used in the industry:
· Conventional Flywheel
· Dual Mass Flywheel
A conventional flywheel (CF) usually consists of a single flat disk and a Dual-Mass
Flywheel (DMF) consists of two sections, primary mass and secondary mass. These
masses are connected to each other in a way so that they can move with respect to
other to a certain degree. Each type is bolted at one side to the crankshaft; in DMF the
primary mass, and the other side is bolted to the clutch; in DMF the secondary mass.
2.6.1 Advantages of DMF over CF
1. Having two different masses; two different sections, isolates vibration in the
crankshaft from passing on to the second mass to some extent. Thereby
preventing undesirable vibrations in the gear box, thereupon preventing gear
rattle
2. Immediate react to increased load amplitude
3. Due to the existence of a medium in DMF, between primary and secondary
mass; and configuration of throttle geometry DMF can be adapted to different
requirements better than CFs
Important effects of DMFs can be stated as:
1. Reduce booming noise and gear rattle by introducing torsion damper
2. Lower fuel consumption by allowing lower engine idle speed
3. Avoid engine and transmission vibration system`s resonance speed by
shifting it to below the idle speed
4. Protect transmission components by reducing engine irregularity and
reduction of primary mass compared to one integrated flywheel
Multi-cylinder Engine Designs
There are several configurations for multi-cylinder engines, namely “INLINE
Engines”, “VEE Engines”, “ROTARY Engines”, “OPPOSED Engines” and
“RADIAL Engines”. From all these arrangements “INLINE Engines” is the most
common and simplest; the engine under consideration in this thesis is also of this type.
In this arrangement as can be deduced from its name all cylinders are in a common
plane. This arrangement gives the simplest way to alleviate unbalancing and shaking
forces in the engine.
Analysis of powertrain mounting system
In the current approach on behalf of simplicity, the mounting system is considered as
a one degree of freedom rotational system, consisting of one torsional spring and an
inertia representing the mounting and the engine, respectively. However, the original
system is comprised of five mounts with three in front and two at the rear of the
engine, the mentioned approach can satisfy our goal with nearly the same precision as
the original system.
The objective which is laid out in this approach is to find appropriate stiffness for
mounts that can assure the following requirements:
· Avoid resonance at low speeds
· Stand engine weight
· Avoid engine bouncing
It is noteworthy that noise and vibration, which are adverse phenomena in the engine
system, enforce mounts to have conflicting characteristics for their best isolation
performance [1]. So that to isolate noise and forces transmitted by the engine to the
vehicle structure, lower stiffness and damping is desired; on the contrary at lower
speeds, typically idle speed, to isolate road induced vibrations, higher stiffness and
damping is required. Care should be taken that reducing the mounting stiffness below
a certain value will lead to rigid body motion in the model. This means that the
mounts cannot fulfil their very first task that is to carry the engine load and restrict the
relative motion of it, particularly bouncing [6].
In this thesis due to considering down-speeding effects on engine compartment, only
vibration is taken into account which is the dominant consequence of this approach.
The system under consideration consists of only one mount and the engine, which is
assumed that behaviour of the mount dynamic characteristics, stiffness and damping,
are linear; also presumed that the mount is massless.
Torsional vibrations analysis of powertrain
The purpose of analysing torsional vibrations of the powertrain is to get the least noise
transferred to the cabin, protect transmission components by reducing engine
irregularity and prevent gear rattle in the gear box.
Analysis of the powertrain is performed for two configurations:
1. Conventional Flywheel (CF)
2. Dual Mass Flywheel (DMF)
For the sake of simplicity the modelling system is considered as 2-DOF and 3-DOF,
respectively. However the complete powertrain model is more complicated than the
considered types; due to insignificant difference, it is acceptable to use the mentioned
models. The complete model can be shown as below:
Comparison
6.1 Conventional Flywheel
6.1.1 Inertia 1
Figure 4.5 demonstrates that configuration 3 has the lowest deflection among all
configurations. It also shows that configuration 2 has the highest deflection. By
studying the results one could deduce that the deflection at each configuration follow
the same pattern and decrease with increasing engine speed (frequency). This is
because the system stiffness and damping has been held constant during different
engine frequencies.
6.1.2 Inertia 2
The deflection for the second inertia shows a completely different manner in
comparison to the first inertia. Here configuration 3 has the highest deflection
configurations 1 and 2 almost follow the same decreasing pattern with increasing
engine frequency, but here configuration 2 has the lowest deflection till 18 Hz of the
engine frequency. From 22 Hz to 30 Hz all configurations nearly have the same
deflection and only differ by a few thousandth of a degree.
6.1.3 Torque
The steady state amplitude of the input shaft`s torque has the same pattern as figure
4.5. Again configuration 3 has the lowest torque and configuration 2 has the highest;
and also the torque decrease by increasing the engine frequency.
As mentioned because deflection of the second inertia (gear box) has the first priority
in designing the torsional system under consideration; the goal is to reduce
oscillations in the gear box and subsequently avoid gear rattle. Therefore
configurations 1 and 2 are the best choices but considering the torque amplitude
configuration 2 is the desired one overall.
6.2 Dual-Mass Flywheel
6.2.1 Inertia 1
The steady state amplitude of deflection at inertia 1 in DMF system has quite the same
pattern as inertia 1 in CF system. Here the second configuration has the highest
deflection till 14 Hz. In this system, configuration 1 has the lowest deflection till 10
Hz but from 14 Hz configuration 3 becomes the lowest.
6.2.2 Inertia 2
At the first 4 engine frequencies (7 to 10 Hz) configuration 3 is the one that have the
lowest deflection. At this time a sudden decrease in the slope can be seen from 10 Hz
to 14 Hz which is due to the existence of the damping.
6.2.3 Inertia 3
Similar to the second inertia configuration 3 has the lowest deflection but this time for
the first 5 engine frequencies. Again a sudden decrease is visible in the deflection
diagram at 10 Hz.
Torque
In the DMF system steady state torque amplitude at configuration 1 has the lowest
value and configuration 2 in the middle of configuration 1 and 3. Also a sudden
decrease in amplitude is obvious similar to the change in deflection part of inertia 2
and 3.
Overall configuration 3 is the one chosen. However torque amplitude of the third
configuration is not the lowest one but its amplitude of deflection is lower than the
other configurations at inertia 2 and 3 and as was mentioned previously decreasing
oscillations at the gear box is the utmost priority.
Notice that all of the sudden decreases in the slope of the indicated diagrams and also
irregularities in the decreasing pattern of the deflection and torque amplitude may be
due to the linear assumption of dynamic characteristics in the system; clearly it shows
that to get the best effect and higher efficiency non-linear characteristics should be
taken for stiffness and damping
Conclusion and further research
recommendations
This thesis investigates whether it is possible to lower the engine speed (downspeeding)
in order to reduce fuel consumption therefore less CO2 emissions. It is seen
that DMF shows better results and facilitate the motion of inertia. The comparison of
deflection at the gear box shows that there is a substantial improvement in DMFs,
which are about 43% lower deflection and consequently less noise and gear rattle.
The data is acquired by assuming the mounting system as 1-DOF system, which has
only one cushion instead of 5. Furthermore the torsional system has been analysed
only through the engine and the gear box and the rest of the powertrain has been
neglected. These assumptions give satisfying results for studying torsional system due
to the fact that practically the torque amplitude after gear box (prop-shaft) is zero.
Thus it will not affect the vibration of the system significantly. In order to get more
accurate results it may suggest using non-linear mounting characteristics and consider
the complete model of both the powertrain and the engine mounts.
As a final point it is concluded that the presented method gives satisfying results for
studying the vibration of the powertrain system.