Heavy Duty Truck Systems by Sean Bennett

40 Hybrid vehicles

Figure 1.21

Proportion of recoverable kinetic energy on front (a) and rear (b) axles depend-
ing on the braking intensity and maximum power of the electric system when
braking from 60 to 0 km/h.

The results shown on Figures 1.20 and 1.21 demonstrated that, for each power of electric
system installed, there is a deceleration value which maximizes the proportion of vehicle
kinetic energy that can be recovered. For decelerations less than this optimum deceleration,
the energy lost via vehicle losses increases. The extreme case, not shown on the figures, cor-
responds to freewheeling deceleration: in this case, the electric system recovers no energy
whatsoever and the initial kinetic energy is dissipated by the resistance to vehicle motion. For
decelerations greater than the optimum value, the electric system saturates at start of braking
and the mechanical brakes must be applied.

In view of the results shown on the previous figures, the question arises as to whether
it would be better to fit the electric machine on the front or rear axle to recover the braking
energy. The systems described in Chapter 5 show vehicles using these two configurations.

Figure 1.22 gives an idea of the recoverable energy difference depending on the position
of an electric machine of power limited to 20 kW. We observe that the recoverable energy
difference is greatest, in particular at low decelerations and moderate initial speeds, without
being decisive. For instance, a vehicle in urban use decelerating at 1 m/s2 from 60 km/h with
a machine placed on the front axle could recover nearly 49% of the total potential, while the
same machine placed on the rear axle would recover 37%.

As mentioned previously, distributing the braking force on the two axles according to
the equal adhesion parabola will guarantee vehicle safety during braking, irrespective of
the adhesion, and maximize the braking force before one of the wheels locks. For small

Chapter 1 · Vehicle use 41

decelerations (e.g. |γ| < 3 m/s2), we may nevertheless ask whether it would be possible to
brake using only the axle equipped with an electric machine. In this case, the risk of locking
this axle, estimated by calculating the adhesion μ necessary to avoid locking the axle, must
be managed.

Figure 1.22

Fraction of recoverable kinetic energy depending on the electrified axle, with
and without 20 kW power limitation when braking from 120 to 0 km/h (a) and
from 60 to 0 km/h (b).

Figure 1.23 gives the adhesion above which axle locking occurs when braking with a
single axle, for various decelerations. We therefore show that when braking with the rear
axle alone, much greater wheel adhesion is required than when braking with the front axle
alone. For instance, if for safety reasons we refuse that regenerative braking alone demands
an adhesion value greater than 0.5 for example on any wheel (0.5 is a low value that can be
reached on a wet road with slightly worn tires), there is no risk of locking before deceleration
of 5 m/s2 if braking is distributed according to the equal adhesion principle; in contrast, if
only the front axle brakes, locking may occur at 3 m/s2, and if it is the rear axle alone which
brakes, locking occurs at 2 m/s2. These results are independent of the initial speed at start of
braking since the highest braking forces are obtained at end of braking if the deceleration is
maintained constant all along the braking phase.

42 Hybrid vehicles

Figure 1.23

Minimum adhesion required to avoid locking the wheels depending on the
braking strategy and braking intensity.

1.6.4 Energy Recovery in Use
When the aim is to minimize the vehicle C 0 2 emissions (which is equivalent to its fuel con-
sumption), the proportion of kinetic energy that can be recovered during average use (e.g. the
urban Artemis cycle) must be examined.

The theoretical analysis conducted previously shows that there are limits to the amount
of energy that can be recovered during deceleration phases. In this paragraph, we will try to
clarify matters by answering the following questions:

- how should braking be managed (both in terms of distribution between axles and in
terms of the contribution of electrical braking compared with mechanical braking) to
try to maximize the energy recovery?

- what is the impact of the powertrain architecture definition (position of the electric
machine: front or rear axle, before or after the gearbox, power capacity of the electrical
components) on energy recovery?

1.6.4.1 Definition of Braking Management Strategies
The braking management strategies are expressed in terms of braking distribution:

- between axles, with in particular choice of position of the electric machine used for
energy recovery,

- between mechanical and electrical braking.

Chapter 1 · Vehicle use 43

A. Distribution Between Axles

Strategy 1 : always braking with the two axles

As explained in paragraph 1.6.2, the equal adhesion parabola indicates, for each brak-
ing level, the distribution between front and rear axles which maximizes braking efficiency,
irrespective of the road adhesion level. For each point on the parabola, both axles demand the
same adhesion level in view of mass transfer. All wheels will therefore lock at the same time
if the availiable tire-road adhesion drops below the adhesion demand (we assume in this case
that all the vehicle wheels are resting on a surface with the same grip, therefore excluding
the possibility of a patch of black ice under some wheels only). While optimum in terms of
braking efficiency, this distribution will nevertheless not maximize energy recovery if only
one axle is equipped with an electric machine.

Strategy 2: braking on a single axle up to a certain adhesion demand

In principle, the standard stipulates that the braking force must be applied on both axles
at the same time. This rule may nevertheless be ignored in case of gentle braking, which is
already the case for vehicles using the engine brake. The engine brake, due both to internal
friction of the engine and the transmission and to engine pumping losses (Chapter 2), is
equivalent to a braking force applied only on the driven axle. The braking torque measured
at the wheel during engine brake phase depends both on the speed of rotation of the engine
and the gear engaged. As an example, we can calculate the vehicle deceleration obtained
using the internal losses of a 1.6 L engine 16: in 1st gear, due to the high gear ratio, we obtain
a braking torque at the wheel of 300 N.m, which corresponds to a vehicle deceleration of
0.70 m/s2 on flat ground. The deceleration is just 0.38 m/s2 in 2 n d gear, dropping to 0.16 m/
s2 in 5th gear.

Along the same lines, on a hybrid vehicle with only one axle equipped with an electric
machine, gentle braking is possible by using only the axle equipped with the electric machine
(in this case, it is best to go into neutral and switch off the engine to cancel its braking force
which has become an unnecessary loss). Several criteria can be used to determine whether
braking is low enough to be carried out with a single axle:

- the braking ratio, which is an image of vehicle deceleration and can be deduced directly
from the driver's braking instruction,

- the adhesion demand on the single braking axle; this quantity is difficult to estimate
but provides a means of keeping a constant wheel locking margin (difference between
the adhesion demand and the tire-road adhesion coefficient).

This second criterion will be used in the remainder of the discussion, since it serves as a
good illustration.

For a vehicle equipped with an electric machine on the front axle, we will define the brak-
ing distribution strategy between the two axles as follows:

- braking by the front axle alone when the adhesion demand on this axle remains less
than a predefined threshold, noted z,

16. Calculation assumption: engine losses equivalent to 1.5 bar of friction losses (Chapter 2), i.e. a
resisting torque of 19.2 N.m at the crankshaft.

44 Hybrid vehicles
- on reaching this threshold, introduction of a rear braking component and modulation
of the front brake to move along a line of constant front axle adhesion demand, until
reaching the equal adhesion parabola. Then, follow the equal adhesion parabola.
This strategy is illustrated by the curve OABC on Figure 1.24 in the case where the aim

is to favor braking on the front axle.

Figure 1.24

Front-rear braking distribution management strategies for an electric machine
on the front axle.
In this figure, it is therefore represented by a segment of the x-axis OA which corresponds
to braking on the front axle only, then a portion of oblique line AB which is the region dur­
ing which the rear brake is progressively introduced while keeping the adhesion demand
(μ = 0.3) constant on the front axle, and lastly a portion BC of the equal adhesion parabola.
The same strategy can be applied to the rear axle rather than the front axle, as shown on
Figure 1.25.

Chapter 1 · Vehicle use 45

Figure 1.25

Front-rear braking distribution management strategies for an electric machine
on the rear axle.

It is technologically difficult to achieve a braking distribution which follows the equal
adhesion parabola: it is easier to create a proportionality between the rear braking force and
the front braking force. In this case, according to the standard, the straight line representative
of this setting remains underneath the equal adhesion parabola up to very high braking ratios,
thereby guaranteeing that the front axle will always lock before the rear axle in order to avoid
vehicle spinning.

The strategies described above still apply, apart from the fact that when braking on both
axles, we follow a straight line, instead of the equal adhesion parabola, as illustrated on Fig-
ure 1.26 for an electric machine on the front axle and Figure 1.27 for an electric machine on
the rear axle.

Note that some manufacturers, such as Toyota with the Lexus LS 600h, propose using
several front-rear distribution straight lines in order to maximize energy recovery during
braking. Since the vehicle is equipped with an electric machine on the rear axle, Toyota has
decided to use a distribution straight line which is above the equal adhesion parabola, but
with strict electronic monitoring of the vehicle behavior (ABS and ESP): vehicle dynamic
stability is analyzed permanently and fast enough to return to a more stable distribution
straight line if the vehicle is considered to be approaching instability conditions [Ueoka et al,
2007].

46 Hybrid vehicles

Figure 1.26
Front-rear braking distribution management strategies for an electric machine
on the front axle, following a straight line underneath the equal adhesion
parabola.

Figure 1.27
Front-rear braking distribution management strategies for an electric machine
on the rear axle, following a straight line underneath the equal adhesion
parabola.

Chapter 1 · Vehicle use 47

B. Distribution Between Electrical Braking and Mechanical Braking

In addition to braking distribution between the two axles discussed above, the development
of hybrid and electric vehicles has led to numerous studies on the distribution between regen-
erative braking with the electric machine and dissipative braking in the mechanical system
on a single axle and several systems have already been proposed by OEMs. The aim of
these developments is to maximize electrical energy recovery during braking phases while
guaranteeing stable vehicle behavior and offering good brake feel during these phases (linear
relation between vehicle deceleration and driver's command).

Unlike friction systems, the maximum resistive torque that can be supplied by an electric
machine is highly dependent on its speed of rotation. This feature must be taken into account
since, by definition, when braking, the vehicle speed decreases; the same applies to the speed
of rotation of the electric machine.

a. Engine Brake Type Recuperative Braking

The first models, in which the conventional mechanical braking system was not modified,
complied with the applicable standard and were not faced with any problems that could
appear with the highly dynamic control of the torque absorbed by the electric machine. In
contrast, recuperative braking is not progressive with the pressure applied on the brake pedal,
but occurs in stages:

- on releasing the accelerator pedal, a constant braking torque instruction Cj is applied
on the electric machine to best reproduce the effect of the engine brake existing on
internal combustion engine vehicles;

- when the driver presses the brake pedal, the torque instruction of the electric machine
is increased and reaches a new constant value C2 to improve recovery.

The instructions values C1 and C2 may be modified depending on the battery charge or
temperature (with in this case a risk of also modifying the driver feel).

In this type of system, the torque requested of the electric machine does not depend on the
actual demand from the driver. Braking progressiveness is provided by the mechanical brak-
ing system. Note also the presence of braking discontinuity for slight pressures on the brake
pedal (Figure 1.30): when the driver has released the accelerator pedal but has not yet pressed
the brake pedal, the torque applied is Cj. As soon as he touches the brake pedal (generally,
as soon as the brake lights come on), braking increases to torque C2. This discontinuity is
generally perceived badly by drivers, so such a point that this type of system can only be used
with low instruction torque values Cj and C2.

b. Simultaneous Electrical and Mechanical Braking

Another simple solution in which the hydraulic braking system is not modified (or only
slightly) consists in maintaining a fixed law between the brake pedal position and the
mechanical braking torque and adding an electrical braking torque proportional to the driv-
er's request. The total braking torque (mechanical + electrical) is then proportional to the
pressure on the brake pedal, which improves the driver feel, and the braking system remains
simple since there is no interaction between the hydraulic system and the electrical system.

48 Hybrid vehicles

The only additional complexity compared with the previous system is the use of a brake
pedal position sensor (or hydraulic circuit pressure sensor) instead of a discrete sensor.

This type of system calls for two remarks:

- firstly, from the lowest braking ratios, the mechanical brakes are activated, which rep-
resents a loss of recoverable energy;

- since the ability of the electrical system to generate a resistive torque may depend on
several parameters (speed of rotation of the electric machine, battery charge level,
etc.), with no possibility of compensating for this variability as the mechanical brakes
are not controlled, the electrical braking ratio must remain low to avoid disturbing the
driver in case of temporary inability to provide electrical braking.

c. System Using a Double-Stroke Pedal

In systems using a double-stroke brake pedal, the first part of the stroke only actuates electri-
cal recovery, while the rest of the stroke actuates the traditional mechanical braking system.
The electrical braking torque obtained at the end of the first brake pedal stroke can be:

- either a constant, in this case a value that can be reached irrespective of the speed of
rotation of the electric machine must be chosen;

- or a value calculated at start of braking depending on the speed of rotation of the elec-
tric machine observed at this moment (in the remainder of the braking, in principle the
vehicle speed should only decrease).

The maximum braking ratio possible in 100% electric depends on the performance of the
electric system (electric machine power and torque, storage power), but also on considera-
tions regarding the vehicle equilibrium during braking: if only one of the vehicle axles is
equipped with an electric machine, electric braking on this axle alone will always be less
efficient in terms of wheel adhesion use and may prove to be unstable if the electrically
braked axle is the rear axle.

In terms of driver feel, the force felt on the brake pedal must be proportional to the total
stroke, irrespective of this double stroke of which the driver is unaware.

d. Collaborative Braking System

To recover more braking energy, even closer collaboration is required between the elec-
trical and mechanical braking systems, which will involve significant modifications to the
mechanical braking system. In order to operate, management of the hydraulic pressure in the
mechanical braking system must be independent of the pressure on the brake pedal, in a by-
wire configuration. The system therefore determines the braking setpoint requested by the
driver by measuring the braking pedal stroke. It then calculates the maximum braking force
that the electric system can supply (depends on the battery state of charge and temperature,
the machine speed of rotation, etc.) and determines the braking force to be applied on the
mechanical system by calculating the difference between the braking force requested by the
driver and that which the electric system can supply.

Figure 1.28 illustrates the operation of this type of system when braking from 90 km/h:

- at the start of the braking phase (phase ©), only the electric machine is actuated to meet
the driver's request; as the driver continues to press on the brake pedal, the electric

Chapter 1 · Vehicle use 49

machine reaches the maximum torque that it can supply in view of its speed of rotation
(a limitation could also be due to the Energy Storage System);

the mechanical brakes are then progressively actuated to meet the driver's request
(phase ©);

during phase d), the driver's request remains constant but since the speed of rotation
of the electric machine decreases at the same time as the vehicle speed, the electrical
braking torque increases;

Figure 1.28

Operation of a collaborative braking system.
Source: [von Albrichsfeld et al, 2009]

- the mechanical braking torque is then controlled dynamically so that the total torque
remains constant and corresponds to the driver's request;

- during phase ®, the speed of rotation of the electric machine has dropped below its
basic speed; consequently, the braking torque that it can supply no longer depends on
its speed of rotation; the mechanical braking torque is now controlled accordingly;

- during phase (D, the speed of rotation of the electric machine becomes too low for the
inverter to inject current into the battery or to properly control the machine; the electric
braking decreases rapidly and the mechanical braking increases;

- lastly, in phase ©, the vehicle is stopped using mechanical braking only.

Toyota proposed this type of cooperative braking system in 2001 in its four-wheel drive
vehicle Estima. This system is also fitted on the Prius since the 2004 model. The braking

50 Hybrid vehicles
circuit diagram of the Prius3 (2009) is given in Appendix 4. Note the presence of a hydraulic
pump operating in both directions (while ABS systems only have a discharge pump), a pres-
sure accumulator and a stroke or brake pedal force simulator.

As part of the EVALVH (ADEME-INRETS-IFP Energies nouvelles) hybrid vehicle
evaluation program, the optimum sharing produced by the collaborative braking system on
the Toyota Prius 2 was demonstrated on a vehicle test. The procedure consists in reproduc-
ing a given fixed deceleration on a test bench while varying the initial battery state of charge.
Sensors measuring the braking circuit hydraulic pressure and the battery current were used
to determine the system behavior. Figure 1.29a shows that the well-charged battery at the
start accepts much less current during deceleration than the discharged battery, although the
driver's instruction on the brake pedal is exactly the same in the two tests. Figure 1.29b con-
firms that the collaborative braking management system reacted by significantly increasing
the hydraulic pressure due to the lower power absorbed by the electric system.

a) Battery current evolution

b) Evolution of the hydraulic pressure in the braking circuit

Figure 1.29
Example of operation of Toyota Prius collaborative braking on 2 identical
decelerations with different battery initial SOC [Vinot et al, 2006].

Chapter 1 · Vehicle use 51

Note that collaborative operation of the electrical and mechanical braking systems must
also be compatible with the various functions that may be fitted on the braking systems
(ABS, emergency brake assist, trajectory control, traction control). This may prove more
especially difficult since, in most cases, a single electric machine is used for the both wheels
on the same axle (coupling being obtained by a mechanical differential), while the above
functions often require individual wheel braking control. In this case, recuperative electrical
braking may have to be canceled as soon as one of these control functions is actuated [von
Albrichsfeld et at., 2009].

Lastly, special attention must be paid to management of downgraded modes: collabora-
tive braking systems can be classified as by-wire systems. The system must therefore be
designed (especially the rest position of solenoid valves) so that in case of any failure, espe-
cially electrical, normal hydraulic braking can be restored.

Analysis of Honda hybrid vehicles demonstrates the developments made in the braking
systems used, as shown on Figure 1.30: the first Honda hybrid vehicle (Insight, 1999) used
engine brake type recuperative braking with two levels, one on releasing the accelerator
pedal and the other on pressing the brake pedal. The same system was used on the 2005
Accord, but completed by a system using simultaneous electrical and mechanical braking,
up to a maximum electrical braking value (value independent of vehicle speed). In 2006, the
Civic inaugurated a collaborative braking system. The two engine brake type torque thresh-
olds remain, but if the driver's braking request increases, electrical braking is used in priority.
The hydraulic circuit is controlled dynamically to provide the additional torque required and
meet the driver's request.

52 Hybrid vehicles

Honda Insight 1999

Honda Accord Hybrid 2005

Honda Civic Hybrid 2006

Figure 1.30
Development of braking strategies on Honda hybrid vehicles.
Source'. [lijima et al, 2006]

Chapter 1 · Vehicle use 53

1.6.4.2 Estimation of Recoverable Energies under Real Use Conditions

The aim of this paragraph is to estimate the quantity of energy that we can expect to recover
depending on the braking management strategy (front-rear axle distribution, electrical-
mechanical distribution) and the choice of architecture (axle supporting the electric machine,
power of the electric machine). We decided to use urban conditions and the corresponding
Artemis cycle. These calculations are based on the assumption that the electric machine is
fitted after the gearbox. A machine positioned before could potentially benefit from the gear-
box to better adapt its operating point, but this is only worthwhile if the machine is used at its
optimum performance levels. For gearboxes with torque interruption when changing gear, it
nevertheless becomes impossible to change gear when braking, unless the mechanical brakes
are controlled to compensate for this. With gearboxes allowing gear change without torque
interruption, it becomes possible to change gear while braking provided that the electric
machine is precisely controlled since, in this case, it is the braking torque after the gearbox
which must continue to respect the driver's request, not the braking control supplied by the
electric machine. Whatever the case, this example corresponds to an additional degree of
complexity which is outside the scope of this study.

In the following calculations, we keep the same reference vehicle (mass 1,360 kg) and
consider that, when braking, the engine does not generate any drag (it is assumed to be
declutched).

The braking management strategies are identified as follows:

- strategy 1 : permanently follow the equal adhesion parabola,
- strategy 1 ' : permanently follow a straight line underneath the equal adhesion parabola,
- strategy 2 (resp. 2'): brake a single axle up to an adhesion demand limit (strategy

adjustment parameter noted z in Figure 1.24), then, follow the equal adhesion parabola
(resp. a straight line underneath the equal adhesion parabola.

Note that strategy 1 is the same as strategy 2, whose limiting adhesion from which brak-
ing is carried out on both axles instead of one would be equal to zero.

Note also that the strategies using a straight line underneath the equal adhesion parabola
rather than the parabola itself tend to favor braking on the front axle.

These strategies are completed by an index corresponding to management of the distribu-
tion between electrical and mechanical braking of the axle supporting the electric machine:

- a: use of the electric machine to the best of its capabilities (T < Tmax and P < Pmax),
which assumes dynamic management of mechanical braking and therefore a collabo-
rative braking system, since the electric machine braking torque changes during brak-
ing depending on the vehicle speed;

- b: use of the electric machine limited to the maximum resisting torque that it can
permanently supply (T < TNmax), which simplifies management of the mechanical
brakes.

Figure 1.31 shows the energy recovery potential on the urban Artemis cycle for various
braking strategies when using an electric machine fitted on the front axle.

54 Hybrid vehicles

When the braking force is permanently distributed between the two axles (strategy Γ
=^> ordinate at the origin of each curve), the recovery potential is already 61% with a 10 kW
machine on the front axle, provided that the full potential of the machine is used (strat­
egy l'a). In this case, by increasing the machine power we can reach 74% recovery with a
20 kW machine, which is close to the asymptote. If, on the contrary, we do not use the full
potential of the electric machine in order to maintain simple control over the mechanical
brakes (strategy l'b), the recovery potential is considerably reduced, dropping to 23% with
a 10 kW machine and 39% with a 20 kW machine. With this simple strategy, having a more
powerful machine (which increases the value of TNmax) offers a substantial advantage: with
a 30 kW machine, the recovery potential reaches 50%.

Figure 1.31

Potential for energy recovery by an electric machine fitted on the front axle,
depending on its power and the braking management strategy - urban Artemis
cycle.

If we change to a specific strategy so that only the front axle is braked during gentle brak­
ing (strategy 2'), we observe much better recovery with an electric machine used over its
entire operating range during braking (strategy 2'a), since we can expect to recover respec­
tively 70, 94 and 100% with an electric machine of 10, 20 and 30 kW. The maximum braking
adhesion on the front axle alone must then be set at 0.2, 0.3 and 0.4, which remains reason­
able. If the electric machine braking torque is limited to TNmax (strategy 2'b), however, the
gain is much less since the electric machine limits the energy recovery. In this case, there is
no point in considering braking on a single axle demanding more than 0.1 adhesion on this
axle.

Chapter 1 · Vehicle use 55

Figure 1.32 shows the same results as Figure 1.31, but for an electric machine fitted on
the rear axle.

This position is less favorable for energy recovery during braking, since the rear brakes of
a vehicle are used less than the front brakes (a situation which is even more pronounced when
following a distribution straight line underneath the equal adhesion parabola rather than the
parabola itself). In this case, a strategy designed to brake the rear axle only for gentle braking
should be implemented. If the vehicle is equipped with a collaborative braking system, the
recovery rate can be increased from 25% to 70%, 94% and 100%, as with an electric machine
fitted on the front axle. This means, however, that the adhesion demand is greater than on
the front.

If the electric machine cannot be used over its entire operating range, saturation of the
machine torque at TNmax then appears highly limiting, hence the need to install a more pow-
erful machine. More than for an electric machine fitted on the front axle, it is essential to
implement a braking strategy on a single axis in order to obtain the same recovery potential
as with an electrified front axle.

Figure 1.32

Potential for energy recovery by an electric machine fitted on the rear axle,
depending on its power and the braking management strategy - urban Artemis
cycle.

56 Hybrid vehicles

1.6.5 Conclusion

Energy recovery during decelerations and braking may contribute significantly to reducing
the energy consumption of electric vehicles. Energy recovery cannot be optimized on stand-
ard vehicles due to the way the brakes are managed: to guarantee vehicle equilibrium and
braking efficiency, conventional vehicles distribute the braking forces between the two axles,
taking mass transfers into account. On hybrid or electric vehicles, a strategy favoring braking
on the axle equipped with an electric machine can be implemented. This is only feasible for
gentle braking, when there is no risk of locking the wheels. In practice, this type of braking
is the most frequently used, resulting in high potential recovery rates, whether the electric
machine is placed on the rear axle or the front axle.

The most important limitation is in fact the ability to control electrical and mechanical
braking simultaneously ("collaborative" braking). A power limitation of the electric machine
corresponds to a torque which varies with the machine speed of rotation. This means that
the electric braking force varies with the vehicle speed. To provide a regular braking force
corresponding to constant pressure on the brake pedal, the mechanical brakes must provide
a variable braking force to compensate for the variation in braking force supplied by the
electric machine. This can be achieved using a by-wire system, which involves completely
redesigning the braking device, its reliability and its downgraded modes. To avoid radically
modifying the mechanical braking system, one solution would be to implement the electric
machine with a constant torque during braking, then provide additional braking with the
mechanical brakes. In this case, the electric machine is under-used, which explains that the
recovery rate is low for small machines and that an electric machine of more than 30 kW is
required in order to recover more than 50% of the potentially recoverable energy in real use.

1.7 CONCLUSION

The calculations presented in this chapter show that, starting from a simple model of the
vehicle, it is relatively easy to calculate the traction energy required for a given mission.
Apart from the values obtained, it is interesting to see how this energy is used and which are
the most important parameters to be varied in order to reduce it. While aerodynamic losses
remain dominant at high speed (therefore mainly on motorway trips), inertia effects consume
most energy in urban and periurban use in particular. In the latter cases, several solutions are
available: either reduce the vehicle mass (choose a vehicle size matching its use), limit the
number and amplitude of accelerations and decelerations per kilometer traveled (i.e. adopt a
smoother and more regular driving style), or consider recovering during deceleration some
of the energy expended during acceleration. This is one of the solutions made possible by
vehicle hybridization (5.1.3, 5.4.1). However, once this energy recovery principle has been
adopted, its optimization on vehicles imposes constraints in terms of vehicle architecture
(position of the electric machine, power of the electric components, etc.) and creates a certain
degree of complexity in terms of coordinated control of the various units.

Chapter 1 · Vehicle use 57

REFERENCES

Brothier JP (1991) Technologie du freinage ABS. Éditions ETAL

Iijima T, Honda R&D Co., Ltd (2006) Development of Hybrid System for 2006 Compact Sedan, SAE
2006-01-1503.

Nakamura E, Soga M, Sakai A, Otomo A and Kobayashi T, Toyota Motor Corporation (2002) Devel-
opment of Electronically Controlled Brake System for Hybrid Vehicle, SAE 2002-01-0300.

Pierre J (1982) Véhicules routiers. Cours de l'École Nationale Supérieure des Techniques Avancées.
Ueoka K, Mashiki Z, Maruyama K, Ito T, Ito M, Toyota Motor Corp. and Tomura S, Toyota Central

R&D Labs, Inc. (2007) Hybrid System Development of High Performance All Wheel Drive
Vehicle, SAE 2007-01-0296.

Vinot E, Badin F, Vidon R, Malaquin B, Perret P et Tassel P. INRETS LTE (2006) Projet EVALVH,
évaluation du véhicule hybride Toyota Prius 2004 et de ses composants, rapport final : Rapport

LTE 0626, novembre (INRETS/RR/06-530-FR).
Von Albrichsfeld C and Karner J, Continental, Division Chassis & Safety (2009) Brake System for

Hybrid and Electric Vehicles, SAE 2009-01-1217.

I Internal Combustion
Engines

Pierre Leduc |

The Internal Combustion Engine (ICE or engine), is the principal source of energy in a
hybrid vehicle. In an electric vehicle equipped with a range extender, a small engine provides
energy when the battery runs low. The engine is a critical component in a hybrid vehicle, and
continued research and development efforts in this field will contribute to reduced energy
consumption in vehicles for a long time to come.

The internal combustion engine has remained dominant in highway transportation even
though electric drive systems and external combustion systems, such as the steam engine,
have been used since the introduction of the automobile. In an internal combustion engine,
the combustion within the machine of a mixture of fuel (primarily of fossil origin) and air
releases gas under pressure that exerts force on a succession of mechanical parts, primarily
the pistons. This mechanical assembly ultimately drives the wheels of the vehicle.

Here, we will concentrate on the piston engine, which is widely used in the automo-
bile industry, even though other internal combustion engines are available, including rotary
engines and turbines. Well suited to automobile use, the piston engine has become, over time,
a highly efficient means of converting the energy from combustion in the form of heat into
energy in the form of work to drive the wheels. And through further technological improve-
ments, the internal combustion engine will undoubtedly see significant progress in the future.

In this chapter we provide background information (thermodynamic behavior, antipol-
lution requirements, fuel characteristics) and a description of gasoline and diesel engines,
and current development efforts. Part of this chapter is devoted to the use of onboard vehicle
engines and the consequences of hybrid use on the operation of the internal combustion
engine. Those seeking additional information should consult reference works such as those
by Heywood [Heywood, 1988] for engines, Guibet [Guibet, 1997] for fuels and engines, and
Kling [Kling, 1980] for thermodynamics.

60 Hybrid vehicles

2.1 INTERNAL COMBUSTION ENGINES - CHARACTERISTICS AND CONTEXT

2.1.1 Thermodynamic Principles of Internal Combustion Engines
2.1.1.1 Types of Engines
There are two types of piston engine: spark-ignition engines ("gasoline engines") and com-
pression ignition, or diesel, engines. Each of these two families is discussed in this chapter.
They have in common the use of combustion chambers, or "cylinders," fed by air and fuel,
most frequently by valves and injectors. In each cylinder, a piston travels in an alternating
linear movement, which moves it toward and away from the top portion of the combustion
chamber, known as the cylinder head (Figure 2.1). The two extreme positions of the piston
are known as top dead center (TDC) and bottom dead center (BDC). The sum of the volumes
swept by the pistons in the various combustion chambers constitutes the engine's "displace-
ment," which is a measure of its size. The pistons are connected to a crankshaft by connecting
rods, a system that ensures the transformation of the up-and-down motion of the pistons into
the rotational movement of the crankshaft. This rotational movement is then transmitted by a
succession of mechanical parts, known as a "transmission," to the wheels to drive the vehicle.

Figure 2.1
Schematic diagram of a piston engine.
Source: [Guibet, 1997]

The combustion chamber is initially supplied with air and, eventually (depending on the
type of engine), with fuel. The movement of the piston compresses the enclosed mass and
at the end of this compression phase fuel can be introduced. After combustion of the air and
fuel mixture in the chamber, the burned gases undergo expansion. During this period, the

Chapter 2 · Internal Combustion engines 61

energy released by the combustion process is transferred to the mechanical system by means
of the pressure forces exerted on the piston top. The burned gases are then evacuated into the
atmosphere through exhaust valves so a new cycle can begin - a new intake phase, followed
by compression and combustion, expansion, and exhaust.

Depending on whether the time allotted to the entire process (intake, compression, com-
bustion-expansion, exhaust) corresponds to a single rotation of the crankshaft, 360°, or two
rotations, the engine is said to be either "two-stroke" or "four-stroke" (in two-stroke engines
[2.4.2.3], the exhaust phase takes place at the end of expansion and intake at the start of com-
pression). Almost all automobile engines use a four-stroke cycle and the formulas used in this
chapter apply to four-stroke engines (Figure 2.2).

Figure 2.2

Operation of a four-stroke engine
Source: [Guibet, 1997]

62 Hybrid vehicles

One of the advantages of a two-stroke engine is its compactness, in other words, a favora-
ble power-to-volume ratio. These engines are frequently used in small, two-wheel vehicles
(motorbikes, scooters) and motorized three-wheelers, which are very common in Asia. In
some new applications, including those that extend the range of electric vehicles, this com-
pactness can be a determining factor. The same is true of rotary engines, like the Wankel,
which are also very compact, and have experienced renewed interest in recent years. These
two types of engines will be discussed at the end of the chapter, in the section on the impact
of hybridization on engines.

2.1.1.2 Simplified Chemistry of Combustion
The mass of fuel introduced into the chamber requires a certain quantity of oxygen (02)
from the atmosphere for its combustion. Nitrogen (N2), the principal component of air, is
also drawn into the engine and forms the majority of the mass enclosed within the cylinders,
but it is chemically inert and does not participate in combustion. For a given amount of fuel
introduced into the combustion chamber, the amount of air admitted is determined by the
flammability limits of the mixture. There is a single mixture of air and fuel that simultane-
ously allows the complete use of oxygen and the transformation, during combustion, of all
the fuel introduced into carbon dioxide (C02) and water (H20). This mixture is called the
"stoichiometry" or "equivalence ratio 1" of the mixture. If there is a less fuel, the mixture is
said to be "lean" (equivalence ratio less than 1); if there is more fuel, the mixture is said to
be "rich" (equivalence ratio greater than 1). Figure 2.3 shows the impact of the equivalence
ratio of the mixture on the constituents of exhaust gas.

Stoichiometric mixture: HC
Lean mixture (excess of air): HC
Rich mixture (excess of fuel): HC

Figure 2.3
Simplified chemical equation of the combustion reaction based on the fuel mix-
ture (HC, for hydrocarbons, represents the fuel) and its impact on the composi-
tion of the exhaust gases.

Chapter 2 · Internal Combustion engines 63

Whatever the equivalence ratio of the mixture, in reality, fuel combustion is never per­
fectly complete and we find, in all cases, traces of unburned fuel (written HC) and carbon
monoxide (CO) in the exhaust gases. During the combustion of a rich mixture, the production
of carbon monoxide becomes significant, for there isn't enough oxygen to oxidize all the car­
bon present in the fuel and transform it into C02. If nitrogen in the air entering the chamber
doesn't participate directly in the chemical reaction, the very high temperatures to which it is
subject during combustion - more than 2,000 K - result in certain molecules dissociating and
oxidizing, forming traces of nitrogen oxides (NO and N02), known as "ΝΟχ". Finally, due
to heterogeneous zones in the air and fuel mixture in the combustion chamber, combustion
zones that are locally rich in fuel - which does not prevent the mixture from being globally
stoichiometric, even lean, when we consider the entire combustion chamber - can lead to the
formation of soot (sometimes called particulates or smoke), especially in diesel engines. CO,
HC, ΝΟχ, and particulates are pollutants emitted by the engine and are subject to antipollu­
tion regulations limiting the amount that can be present in the vehicle's exhaust gas.

2.1.1.3 Power, Load, and Output

The engine produces a torque whose value is primarily associated with the amount of fuel
burned in the combustion chambers. The power produced by the motor is the product of this
torque and the engine's rotational speed:

P = T.co (2.1)

where P is power in W, T is torque in N.m, and ω is the rotational speed in rad.s-1.

The "load," or "load level," is the amount of fuel introduced into the cylinders. The
engine is said to be "at wide open throttle" if it takes in the maximum quantity of fuel that
can be burned for a given engine rpm. This is the case when a driver presses the accelerator
pedal to the floor. Conversely, the engine is at "low load" if the amount of fuel introduced is
low and the engine produces little torque or power (for example, when slowing down or driv­
ing at low speed). The concepts of load and engine speed must be distinguished: an engine
can operate at "low load" even at a high rpm; it can also operate at low rpm even though
it is operating at full throttle (consider a vehicle climbing a steep incline). The load can be
quantified using mean effective pressure (MEP), the presumably constant fictive pressure
that, when multiplied by the corresponding volume of the engine's displacement, supplies the
same amount of work produced by the engine during a cycle. We can show that:

(2.2)

Vc: total engine displacement, in m3; MEP in Pa.
MEP is, therefore, proportional to the engine's specific torque, that is, torque divided by

the displacement (often expressed in N.m/L). RPM and MEP are two parameters frequently
used to characterize an engine's operating point.

The efficiency of an engine differs depending on the operating point under consideration.
It directly influences its specificfuel consumption (SFC), which corresponds to the consump­
tion of fuel compared to the power produced by the engine, and, therefore, measures the

64 Hybrid vehicles

efficiency with which fuel is used to produce power. The energy content of fuel is given by
its lower heating value (LHV). The efficiency η of the engine is calculated from the follow­
ing formula:

(2.3)

When analyzing these phenomena in greater detail, we find that the engine's overall effi­
ciency results in an "efficiency breakdown", each basic output characterizing a type of loss.
These elementary outputs can be broken down into:

- combustion efficiency r\c
- thermodynamic efficiency T|th
- cycle efficiency η
- mechanical efficiency r\m

Therefore, the overall efficiency can be written as:
(2.4)

Below, we provide orders of magnitude for the different elementary types of efficiency
(2.2.2.3 and 2.3.1.1).

Combustion efficiency can be used to determine, from LHV, the amount of fuel that is not
burned or only partly burned during the combustion stage. This amount corresponds to the
residual heating value of HC and CO present in the burned exhaust gases. Increased engine
efficiency, therefore, results from the most complete combustion possible.

Thermodynamic efficiency characterizes an ideal cycle that would consist of an adiaba-
tic compression phase followed by isochoric combustion, adiabatic expansion, and isobaric
exhaust and intake phases. Thermodynamic efficiency is associated with two parameters.
The first is the compression ratio, τ, which is a geometric characteristic of the engine relative
to the minimum volume of the combustion chamber; the second is a thermodynamic coef­
ficient, γ, which depends on the composition of the gases in the combustion chamber. The
higher the value of these two parameters, the better the efficiency. The more dilute the fuel
mixture, preferably with air (lean combustion mixture), the greater the value of γ. However,
dilution with burned gases is also possible (the recirculation of burned gases is referred to as
exhaust gas recirculation, or EGR). In an ideal cycle, we assume that the release of energy is
instantaneous and that it is produced when the volume of the chamber is minimal.

It can be shown that:

(2.5)

Cycle efficiency can be used to account for the various phenomena that distinguish the
real cycle from the ideal cycle (non-instantaneous combustion, heat losses at the walls,
energy losses associated with the transfer of gases during intake and exhaust phases, varia­
tion in the coefficient γ during the cycle, etc.). Among those losses, the loss associated with
mass transfer during the intake phase can be high, especially in a gasoline engine where the
load is controlled by means of the throttle, which generates high pressure drop (2.2.2.1). With
respect to heat losses at the walls, these decrease with lower combustion temperatures. Here
too, combustion of a lean mixture can improve efficiency.

Chapter 2 · Internal Combustion engines 65

The losses associated with the work expended to provide energy to the auxiliary equip­
ment essential to engine operation (a water pump to cool mechanical parts, an oil pump for
lubrication, an injection pump for diesel engines, an alternator to produce electricity in the
vehicle) and the losses from mechanical friction within the engine are considered in terms of
mechanical efficiency. It should be noted that frictional losses increase with engine speed.

Mechanical efficiency is expressed as the ratio of the available work supplied by the
engine at the end of the crankshaft and the work transmitted to the pistons by the forces gen­
erated by pressurized gas:

(2.6)

where p is the pressure of the gases in the cylinder in Pa and V is the cylinder volume in m3.

Since the origin of the engine, great progress has been made in terms of its energy effi­
ciency. Although the absolute values of the maximum efficiency of automobile engines may
appear modest, ranging as they do from 0.35 to 0.4, it's important to remember that these
values reflect the heating power of the burned fuel and that, before producing useful work
- needed to move the vehicle - the primary result of combustion is the release of heat. The
laws of thermodynamics show that the conversion of heat to work can never be complete,
far from it.

Improving the efficiency of piston engines now involves fine tuning each of the four basic
types of efficiency listed earlier. If progress is to continue, advanced research and develop­
ment will be needed in several fields, ranging from a detailed understanding of the com­
bustion process to the development of technologies for more precisely controlling energy
release.

2.1.2 Evolution of Pollution Standards and C02 Emissions

Vehicle exhaust gases are primarily composed of nitrogen (N2), carbon dioxide (C02), and
water vapor (H20). Also found are traces of carbon monoxide (CO), unburned fuel, and
products resulting from the partial oxidation of fuel (HC), nitrogen oxides (ΝΟχ), and par-
ticulates. CO, HC, ΝΟχ, and particulates are dangerous for humans and treated as regulated
pollutants. Pollution emissions from vehicles must comply with the limits set by laws, which
vary from place to place. The first such standards were enacted in the United States in the
mid-twentieth century. Measurements are made using predefined driving cycles, each region
of the world having its own cycle. In Europe, the NEDC cycle is the one currently in use,
consisting of an urban component (ECE cycle) and an ex-urban component (EUDC cycle).

Regulations are complex. In general, they take into account the engine type (gasoline or
diesel), the vehicle weight, and the number of passengers the vehicle can carry. Correction
factors are used to account for the age of antipollution systems, and so on [Degobert, 1992].
Throughout the world, these regulations have become stricter over time. In Europe, the first
regulations appeared in the 1960s in Germany and France. After 1972, common regulatory
limits were applied throughout the European Economic Community. The measurement of

66 Hybrid vehicles

pollutants was carried out using the ECE cycle alone. Initially, this involved measuring CO
and HC emissions; after 1976, ΝΟχ emissions were also taken into account. Until 1984 emis­
sions limits were based on vehicle weight, measured in steps of approximately 200 kg. After
1984, for passenger vehicles less than 2,500 kg, the weight categories were replaced by two
primary categories - gasoline vehicles and diesel vehicles - which were further broken down
into three subcategories based on engine displacement. Only in 1992, with the arrival of the
so-called Euro 1 standard, was reference to vehicle category abandoned (gasoline/LPG/CNG
or diesel) for all passenger vehicles weighing less than 2,500 kg, regardless of their displace­
ment. Also in 1992, the approval cycle began to include an ex-urban component in addition
to the urban cycle. With the exception of the elimination of an initial engine operating phase
in 2000, this cycle hasn't changed. Table 2.1 presents the evolution of the European standard
for light vehicles since 1992. Note that between 1992 and 2014, emissions of HC and NOx
were reduced by a factor of 5 and particulate emissions from diesel vehicles by a factor of 30,
at the cost of tremendous research and development efforts by the entire automobile industry
and various research labs.

To satisfy these ever more restrictive requirements and because engines alone could no
longer generate cleaner exhaust gas, the automobile industry gradually had recourse to after-
treatment systems, such as catalysts and filters, placed in the vehicle's exhaust system, as
discussed below.

Table 2.1. Change in European antipollution regulations since 1992, in g/km,
for passenger vehicles weighing less than 2,500 kg

Gasoline/LPG/CNG vehicles Diesel vehicles
CO HC + NOx HC ΝΟχ Particulates CO HC + NOx HC ΝΟχ Particulates

Euro 1 2.72 0.97 2.72 0.97 0.1400
(1992)

Euro 2 2.20 0.50 1.00 0.70 0.0800
(1996)

Euro 3 2.30 0.20 0.15 0.64 0.56 0.50 0.0500
(2000) 1.00 0.30 0.25 0.0250
1.00 0.10 0.08 0.0050 l 0.50 0.23 0.18 0.0050
Euro 4 1.00 0.10 0.06 0.0045 l 0.50 0.17 0.08 0.0045
(2005) 0.10 0.06 0.50

Euro 5
(2009)

Euro 6
(2014)

1. Only for direct injection gasoline-powered vehicles.

The European Commission is currently investigating ways to modify the certification
process so it is more representative of real-world driving conditions. Over the long term,
plans are also underway to develop a worldwide driving cycle that will be used around the
world, thereby replacing the multitude of certification cycles currently in use. Although
regional comparisons are difficult because the certifications themselves - procedures and

Chapter 2 · Internal Combustion engines 61

driving cycles - are very different, Table 2.2 provides a summary comparison of the regu-
lated emissions levels in Europe, Japan, and the United States for gasoline vehicles. It can
be seen from the table that China's standard is close to the Euro 4 level (Table 2.1). India
also bases its standards on Euro 3 or Euro 4 levels (depending on the region). In California,
the applicable standards are stricter than the American federal standards shown in Table 2.2.

It is important to keep in mind that the introduction of a new standard affects new vehi-
cles alone; it has no impact on vehicles already in circulation. Therefore, before a standard
becomes fully effective, the entire automobile pool must be completely replaced.

Table 2.2. Simplified comparison of pollution emission levels in Europe, Japan, and the United States
(gasoline and LPG vehicles)

Europe (Euro 5 - 2009) HC (Europe, Japan) NOx Particulates
CO NMOG/HCHO (United
Japan (2009) 0.060 0.005 3
United States1 States)2 0.050 0.005 3
1.00 0.100 0.031 0.006 4

1.15 0.050
2.11 0.047/0.009

1. The American standard is especially complex. Here, we consider the "bin 5" category at 50,000 miles (approxi-
mately 80,000 km).
2. Unlike Europe and Japan, which take into account unburned hydrocarbon emissions, the United States regulates
the emission of hydrocarbonated and oxygenated organic compounds, with the exception of methane (NMOG, or
non-methane organic gases) and formaldehyde (HCHO).
3. Only for direct injection gasoline vehicles.
4. At 120,000 miles (approximately 200,000 km).

Because of their harmful contribution to the greenhouse effect, C 0 2 emissions have
become an environmental concern, although much more recently than the regulated pollut-
ants shown above. C 0 2 emissions from individual vehicles taken in isolation are not directly
subject to limits as is the case with regulated pollutants. However, in some parts of the world,
especially Europe, governments are trying to limit C 0 2 emissions for all automobiles in use.
This involves tax incentives and regulations aimed at eventually achieving a mean C 0 2 level
per kilometer driven for all new vehicles in circulation. The emissions target is 120 g/km on
average in Europe, between 2012 and 2015, and 95 g/km by 2020. In comparison, the aver-
age emissions for new vehicles in circulation in Europe in 2009 was 146 gC02/km.

Ever since C 0 2 became a cause of concern - the first official determination of emissions
from new vehicles sold in Europe was made in 1995 - automobiles have continued to make
progress in emissions reduction. Figure 2.4 presents the evolution of C 0 2 emissions from
new vehicles sold in Europe for the period 1995-2009. Because ultimately all the carbon
contained in fuel burned in the engine appears as C02, the reduction of C 0 2 emissions in new
vehicles sold demonstrates that engine efficiency has greatly improved during that period.

Following the oil crisis of 1973, the U.S. Congress took action intended to reduce vehi-
cle consumption through Corporate Average Fuel Economy (CAFE) legislation. Initiated in

68 Hybrid vehicles

1975, the program is still in force. Each year it establishes an average target fuel "consump-
tion" 1 per vehicle that each automobile manufacturer must achieve for all sales in the United
States. There is a financial penalty for manufacturers who fail to meet the standard. The far-
ther from the regulatory target, the higher the penalty. Even if the goals of the CAFE program
have not been met by all manufacturers, the standard has had positive effects and, in forty
years, average emissions for new vehicles sold in the United States have declined from 278
gC02/km in 1978 to 169 g/km in 2009. Figure 2.4 compares the change in C 0 2 emissions per
vehicle in the United States and Europe since 1995.

In Japan and South Korea, C 0 2 emissions for new vehicles in circulation have reached
levels that are very close - to within a few grams per kilometer - to the levels found in Europe.

Figure 2.4
Change in C02 emissions for new vehicles sold in Europe for the period 1995-
2009 [CE, 2009] and comparison with the United States [NHTSA, 2011]. Fig-
ures for 2010 are provisional.

2.1.3 Fuel
2.1.3.1 Traditional Fuels
Conventional petroleum-based fuels currently provide nearly all the energy used in moving
people and goods given that gasoline and diesel account for 93.8% of the energy used for
highway transport on a global scale. Biofuels account for only 2.9%, natural gas for 2%, and
LPG 1.3%.

Fossil-based liquid fuels provide, in addition to ease of storage and handling, consider-
able benefits for transportation. Their high mass energy density (approximately 43 MJ/kg
1. Contrary to European countries who express fuel consumption in L/100 km, the US consider it in
miles driven par gallon consumed, i.e. "mpg"

Chapter 2 · Internal Combustion engines 69

for gasoline and diesel) provides vehicles with considerable range, often more than 600 km.
Refining technology is proven and inexpensive. However, the exhaustion of fossil fuels as
well as increasingly severe environmental constraints have increased the need to find alterna­
tives to "historic" fuels.

Figure 2.5
Energy consumption associated with global highway transport in 2009 by fuel
type (KBC, IFPEN, NGVA).

2.1.3.2 Liquefied Petroleum Gas (LPG)

Historically, liquefied petroleum gas was the first true alternative fuel. It is a mixture of
butane and propane produced by petroleum refining (40% of the resource) and natural gas
processing (60% of the resource on a global scale).

LPG consumption accounted for 20.4 MTOE (million tons oil equivalent) in 2008 world­
wide, an increase of 12% over 2005, primarily from the introduction of LPG vehicles in
Eastern Europe. For example, the number of LPG-powered vehicles in Poland increased
from 470,000 in 2000 to nearly 2 million by the end of 2007.

However, the use of LPG vehicles in France is not widespread and, at this time, its devel­
opment appears to be stagnant. There were only 140,000 LPG-powered vehicles in France
compared to 200,000 in Germany and a million in Italy, although there are 2,000 fueling
stations distributed across 98% of the French highway system.

For those vehicles that use LPG, there are a number of advantages:
- C 0 2 emissions are reduced by 10% compared to a gasoline vehicle because of a more

favorable H/C (hydrogen-to-carbon) ratio, this ratio is much more favorable when
compared to diesel;
- Reduced ΝΟχ emissions compared to gasoline and diesel vehicles;
- No particulate emissions;
- High octane level2, which can result in an improvement in energy efficiency.

2. The octane rating of a fuel measures the fuel's resistance to self-ignition. The higher the octane
rating, the better the resistance to engine knocking (2.2.2.1).

70 Hybrid vehicles

Additionally, lower taxes have been established to launch LPG use, but success has been
relatively modest due to the non-negligible additional cost of vehicle purchase (15 to 20%
higher than an equivalent gasoline-powered model), a limited distribution network outside
major thoroughfares, and the space needed to accommodate the additional tank in the vehicle
(in France, LPG vehicles are primarily dual-fuel vehicles - gasoline and LPG - there are two
fuel tanks in the vehicle).

2.1.3.3 Compressed Natural Gas (CNG)

Compressed Natural Gas for Natural Gas Vehicles (NGV) has been the focus of increased
interest associated with the development of the natural gas market and the presence of more
durable resources than petroleum. Nonetheless, CNG consumption remains very limited. It
accounted for 37 Mtoe in 2010, 22.6 Mtoe for passenger vehicles and the remainder for
trucks. Italy was the market leader in Europe due to the growth of the CNG distribution net-
work, which began in the 1930s. Worldwide, Pakistan, Iran, Argentina, and Brazil have NGV
fleets that exceed a million units and are expanding rapidly.

In France, 2010 consumption barely reached 98 Ktoe with some ten thousand vehicles
in circulation. NGV have a number of advantages for the environment compared to gasoline
and diesel vehicles:

- C 0 2 emissions reduced from 20 to 24% compared to a gasoline vehicle due to the
highly favorable H/C ratio;

- Reduced NOx emissions compared to diesel vehicles;
- No particulate matter emitted and, in general, the toxicity of exhaust gases is sharply

reduced;
- Good resistance to knocking, which improves engine efficiency (2.2.2.1).

However, this does not mean that we should overlook the drawbacks of a relatively
untested approach:

- there is as yet no distribution network in France;
- on-board storage and range are limiting factors;
- the additional cost of the vehicle is still significant but the gradual elimination of dual-

fuel systems (gasoline/CNG) should help reduce it.

Technological developments are expected both in terms of improving dedicated CNG
drive systems and lowering pollution levels to comply with future standards.

2.1.3.4 Biofuels

At present there are two major types of so-called "first-generation" biofuels on the market:
- ethanol, which is produced from sugar-based plants such as cane sugar and beets, and
starchy plants such as wheat or corn, is used in gasoline engines;
- fatty (vegetable or even animal) acid methyl esters (FAME) are used in diesel engines.

Chapter 2 · Internal Combustion engines 71

Figure 2.6

First-generation biofuel pathways.

The principal biofuel in terms of volume consumption is bioethanol. Some 52 Mtoe were
produced globally in 2010, mostly in Brazil and the United states. At the same time, the
production of biodiesel reached 15 Mtoe. Germany, with production of 1.71 Mtoe, was the
leading worldwide producer, slightly ahead of Brazil and France.

The principal advantage of biofuels is that they can be mixed with gasoline and diesel,
thereby taking advantage of conventional fuel distribution networks. Once mixed, they do
not require any significant technological modification to vehicle engines.

Nonetheless, the use of biofuels must take into account resource sharing between food
and energy use. Worldwide management will have to be established in order to satisfy the
requirements of both uses. Additionally, while the C 0 2 balance of biofuels appears to be
positive and some impacts, primarily on the global scale (greenhouse effect, exhaustion of
fossil fuel resources), are well understood, others (the affect on water, soil, etc.) are poorly
understood, especially on the local level.

New pathways are being developed for more efficient use of biomass resources. Second-
generation biofuels (obtained from the chemical transformation of lignocellulosic biomass)
will enable us to avoid conflicts with the food sector, and the third generation should enable
us, over the long term, to transform algae into biodiesel, bioethanol, biogas, or hydrogen.

2.1.3.5 Hydrogen

Like electricity, hydrogen is an energy vector. Over the long term, hydrogen may come to be
seen as a fuel, used alone or mixed with natural gas (up to 20%) in an internal combustion
engine. Its use in the pure state, in a fuel cell and electric engine, can be viewed as an alterna-
tive to the direct storage of electricity in batteries.

72 Hybrid vehicles

At present, 99% of the hydrogen produced is used as an industrial gas. The refining sector
is the largest consumer (51%), followed by ammonia production (34%) and other specialty
chemicals (14%). Only 1% of global volume is used today for transport, primarily in the
space sector.

The most common sources for producing hydrogen are fossil energies. Steam reforming
of natural gas is the most commonly employed technology for producing large quantities at
low cost. Production from biomass or the high-temperature electrolysis of water are currently
the most promising alternatives from the environmental perspective.

It is important to note, however, that over the long term, hydrogen use will involve the
development of heavy infrastructure (pipeline transport, intermediary storage, in-vehicle
storage), which will lead to technical difficulties and significant additional costs. At present,
there are only about 40 service stations distributing hydrogen worldwide, spread relatively
uniformly among Europe, North America, and Japan.

2.1.3.6 Energy Comparison
Liquid hydrocarbons offer a definite advantage in terms of energy density per unit vol-

ume compared with gaseous fuels (hydrogen and CNG), as illustrated on Figure 2.7. For
gaseous fuels, high-pressure in-vehicle storage is necessary if we want to ensure acceptable
range, that is, sufficient energy density. Table 2.3 compares the principal characteristics of
some common fuels.

Figure 2.7
Comparison of energy densities for fuels. (To simplify the comparison with
data from the chapter on batteries, energy is expressed in kWh.)

Note that the values indicated on Figure 2.7 do not take into account the storage and
in particular the tank weight, which may be very high for gaseous fuels. A complete table
presenting the characteristics of the different forms of energy storage appears in Chapter 4
(Table 4.16).

Chapter 2 · Internal Combustion engines 73

Table 2.3. Comparative characteristics of fuels

H/C ratio O/C ratio Density at LHVby LHVby AFR1
(-) (-) volume mass (-)
standard (kJ/L)
temperature (kJ/kg)

and
pressure
(kg/m3)

Liquid fuels

Diesel 1.83 0 830 35,400 42,600 14.5

Biodiesel 1.89 0.111 882 33,100 37,500 12.4
(rapeseed ester)

Gasoline 1.82 0 750 32,200 42,900 14.6

E5 gasoline 1.89 0.016 752 31,700 42,100 14.3

E10 gasoline 1.93 0.033 754 31,100 41,200 14.0

E85 ethanol 2.75 0.392 787 22,900 29,100 9.8

Ethanol 3.00 0.5 789 21,100 26,800 9.0

Gas fuels

LPG (50% butane, 2.57 0 2.2 - 46,000 15.5
50% propane)

Methane 4.00 0 0.7 - 50,000 17.2

Hydrogen -- 0.1 - 120,000 34.2

1. AFR: Air-Fuel Ratio.
Even though all the diesel fuel distributed at the pump must comply with current specifications, diesel contains
several hundred different hydrocarbon molecules, which explains the variation among diesel fuels. Additionally,
diesel specifications vary significantlyfrom one region of the world to another. As with dieselfuel, gasoline contains
several hundred different hydrocarbon molecules, leading to variations in its physical-chemical characteristics.
Here as well, the characteristics of gasoline can vary significantlyfrom region to region. A gasoline labeled "Ex "
contains x% of ethanol. Ethanol is very hydrophilic and can contain up to 7% water. LPG fuel consists mostly of
propane (C3Hg) and butane (C4H10). But it also contains small quantities of other molecules, such as Cfl6, C4H8,
and C5H12. Its LHV, therefore, can vary as afunction of its composition. Mostly composed of methane (CH^, natu-
ral gas for vehicle use varies with supply and can contain significant amounts of inert gases, primarily nitrogen.

2.1.3.7 Cost and Taxation of Fuel in France

French taxes on petroleum products consist of two elements: a tax that applies to petroleum
products themselves (a domestic tax on petroleum products, known as TIPP) and, on top of
that, a value-added tax, or VAT. Since April 1, 2000, the VAT has been 19.6%. TIPP con-
sists of a fixed amount per liter of fuel sold. For 2011, the amount of TIPP established by tax
legislation is shown in Table 2.4.

74 Hybrid vehicles

Table 2.4. TIPP for various products. Data on petroleum taxation (DGEC-DE, updated January 2011)

E85 ethanol Unit TIPP (euros)
Unleaded 95 RON E10 hL 17.29
Unleaded gasoline hL 60.69
Diesel hL 60.69
Heating oil hL 42.84
Aviation gasoline hL 5.66
Jet fuel (airplane use) hL 35.9
LPG hL 0
Heavy fuel oil hL 5.99
Natural gas t 18.5
100 m3 0

We see from the table that there is considerable disparity in terms of cost and taxation
among the fuels available in France. The least expensive fuel in terms of energy content is
CNG, with a price slightly below 15 euros/GJ. This offers considerable advantage when
compared to other fuels, which are heavily taxed. Figure 2.8 represents the prices and taxes
of each type of fuel. Electricity is included as well, for it is used, together with hydrocarbon
fuel, in rechargeable hybrid vehicles (see Chapters 5 and 7).

Figure 2.8

Cost, including taxes, of fuels in France in 2010.
Source : [MEEDDM, 2011]

Chapter 2 · Internal Combustion engines 75

2.2 SPARK-IGNITION ENGINES (GASOLINE ENGINES)

2.2.1 General

The most common engine type in the family of spark-ignition engines is the gasoline engine.
It employs a dedicated spark device to create a flame in the combustion chamber. The device
consists of a spark plug that generates a spark from an electric energy source at the appropri-
ate moment ("the spark timing"). Among the family of spark-ignition engines, there are gaso-
line engines, natural gas engines, LPG engines, and ethanol engines. "Flex-fuel" engines,
capable of operating on either gasoline or ethanol or a mixture of the two, are also examples
of spark-ignition engines. The piston engines that use hydrogen as fuel, currently the subject
of heavy research effort, are generally spark-ignition engines.

2.2.2 The Standard Situation

The standard situation, described below, is based on an engine equipped with a port fuel
injection system (that is, in the air-intake duct), operating stoichiometrically, and equipped
with a "three-way" catalytic converter for pollution prevention.

2.2.2.1 The Engine

A. Combustion

In a spark-ignition engine, the most common combustion method is homogenous combus-
tion of a gaseous mixture of fuel and atmospheric oxygen in stoichiometric proportions.
This means that, before the ignition spark appears, the fuel, if liquid, is vaporized, mixed
fully with air, and introduced uniformly into the combustion chamber. Known as "mixture
preparation," this phase is made more complex by the fact that it must occur very quickly.
For example, when the engine operates at high speed, the mixture must be prepared in
approximately 10 ms. After the spark is introduced, a so-called "pre-mix" flame develops
in the chamber. In conventional spark-ignition engines, the engine load is controlled by the
density of the fuel mixture that enters the combustion chamber. Most often, a valve is used
("throttle") to restrict the passage of air into the engine's intake ducts. This throttle creates
a pressure drop that directly controls the pressure and, therefore, the density of the mixture
introduced into the combustion chamber.

76 Hybrid vehicles

Additionally, the velocity of the laminar flame in spark-ignition engines is incapable of
ensuring complete combustion in the time allotted by engine rpm 3. Therefore, flame velocity
must be enhanced by ensuring that the aerodynamics in the combustion chamber are such that
they create turbulence in the fuel mixture. In such a case, rather than developing spherically,
the flame will bend under the effect of turbulence, thereby increasing the chemical combus-
tion reaction surface and leading to an apparent reaction rate - known as turbulent flame
velocity - that is greatly superior to the laminar flame velocity that occurs during the earliest
stages of combustion.

The aerodynamic movement of the gaseous mixture inside the combustion chamber is
consecutive with the introduction of gas at high velocity through the intake ducts and valves,
as well as the rapid motion of the piston in the chamber. The shape of the ducts and, some-
times, the profile of the piston head can strengthen the intensity of the aerodynamics within
the engine. The final intensity is a compromise. The flow of the gaseous mixture should be
as fast as possible to promote high turbulent flame velocity, but its velocity must be such that
energy losses at the walls, caused by convection movements that promote losses through heat
transfer, are not too high. Designing the profile of these elements is the result of a number of
optimization procedures using 3D-simulation tools and flow bench experiments to evaluate
internal flow within the cylinder head.

Figure 2.9
Flow velocity field during the intake phase. 3D calculations were made using
LES (Large Eddy Simulation), which is the most precise simulation technique
in current use. Note that the air velocity is at a maximum as it passes through
the intake duct.

3. Typically, laminar flame velocities are slightly less than 1 m/s and engine rpm varies between 650
and 8,000 rpm or higher in engines for competition vehicles. At 2,000 rpm, for example, and assuming
a flame velocity of 1 m/s, it would take 40 ms for the flame to travel the typical distance of half the bore
of a combustion chamber, that is, 80 mm in diameter - assuming the spark plug is in the center of the
cylindrical chamber - which is more than the time required by the engine to make a complete revolution
(30 ms at 2,000 rpm).

Chapter 2 · Internal Combustion engines 11

B. Knock

In certain situations, combustion can become abnormal, that is, combustion no longer occurs
from the development of a flame out of a spark generated by a spark plug. The most common
form of abnormal combustion, "knock," results from auto-ignition of an unburned part of the
air and fuel mixture due to the effect of temperature and pressure in the combustion cham-
ber. This spontaneous ignition creates a second flame front in addition to the one normally
propagated from the spark. The release of energy in the combustion chamber then becomes
very rapid and generates pressure waves that excite the walls, producing the characteristic
knocking sound. This phenomenon, unless it is rapidly discontinued, can lead to the destruc-
tion of the engine. Knock is a limitation that is taken into account during the earliest stages
of gasoline engine design. To limit the risk of knock, the compression ratio (2.1.1.3) of the
engine shouldn't be too high. However, the greater the compression ratio, the better the ther-
modynamic efficiency. The final compression ratio, therefore, is something of a compromise.

Recent "downsized" engines (which will be discussed below) make use of combustion
under very high loads. It turns out that, under certain conditions, which are still poorly under-
stood but are the subject of extensive research involving techniques of direct visualization of
the combustion chambers, these engines can be exposed to violent and sporadic combustion
that can sometimes exceed 200 bar. There is no need to point out that a single occurrence of
such combustion is capable of destroying a gasoline engine, generally designed to sustain
maximum pressures of 80 to 120 bar.

C. Principal Components and Control of Combustion

Two specific types of components are necessary for spark-ignition engines: actuators and sen-
sors. The principal actuators are the gas throttle, the ignition system (spark plugs and coils),
and the injectors. The principal sensors needed to control a spark-ignition engine are those
used to record air flow into the engine (measured directly with a flow meter or indirectly by
means of a pressure sensor downstream of the throttle), intake air temperature measurement,
a probe for determining the air/fuel ratio of the initial mixture based on a measurement of the
exhaust gases ("lambda sensor"), and an accelerometer for detecting knock.

Very often today the throttle is an electrically controlled actuator. The cable control
between the accelerator pedal and this throttle has disappeared. The pedal has become a poten-
tiometer and the driver's request for power is sent to a computer that determines the position of
the throttle. At low load, when the power requested by the driver is low, the throttle is almost
completely closed, which generates strong pressure drop in the air intake circuit. The admis-
sion of the mixture into the combustion chamber is then accompanied by a resistance (known
as "pumping loss") that the engine must overcome. These pumping losses are one of the
reasons for the energy losses found in a spark-ignition engine and partly explain the mediocre
efficiency of gasoline engines compared to diesel engines, especially at low load.

If combustion of the entire mixture were instantaneous, the combustible mixture would
have to be ignited at top dead center to maximize efficiency. Because combustion is spread out
over some tens of angular degrees of crankshaft rotation, the moment of ignition is advanced
with respect to the passage of the piston at top dead center to compensate for the duration of
combustion. The moment the spark at the spark plug's electrode is triggered (referred to as the

78 Hybrid vehicles

"spark advance") is today recorded in "maps," numerical tables stored in the on-board com­
puter that monitors the engine during operation. This moment is optimized for every operating
point in the engine, that is, it will vary with the rpm, load, temperature, and so on.

In the event of abnormal combustion, engine operation is quickly corrected to preserve its
integrity. An accelerometer mounted on the engine measures vibration during the combus­
tion process. If knock occurs, certain frequencies are excited and this information is sent to
the engine's computer. The spark advance is then slowed to reduce pressure and temperature
levels in the combustion chambers, stopping the knock. For the injectors - one per cylinder
- both the moment they open (which marks the beginning of the introduction of fuel into the
intake duct) and the duration they remain open are controlled. This second parameter deter­
mines the amount of fuel injected during each combustion phase.

2.2.2.2 Emissions Control and Complete Combustion
To comply with anti-pollution standards, spark-ignition engines are equipped with a catalytic
converter for processing exhaust gases (Figure 2.10). In the great majority of cases, the tech­
nology used involves a "three-way" catalytic converter. This converter has the advantage of
being able to simultaneously process unburned hydrocarbons (HC), carbon monoxide (CO),
and nitrogen oxides (ΝΟχ), providing that the different gases exist in a proportion that results
from stoichiometric combustion and that the catalyst has reached sufficient temperature,
typically 250 to 300 °C 4.

Figure 2.10
Engine and exhaust system. The digital image shows the engine (gray), the
transmission (blue), and the three-way catalytic converter with the beginning of
the exhaust pipe (orange).
4. In the three-way catalytic converter, oxidation reactions (for HC and CO), which assume the pres­
ence of oxygen, occur simultaneously with chemical reduction reactions (for the ΝΟχ), which become
impossible if the oxygen concentration is too high. Therefore, the operating range is very narrow from
the point of view of the fuel/air ratio of the mixture.

Chapter 2 · Internal Combustion engines 79

Most of the time, emissions control restrictions on exhaust gases require that spark-igni­
tion engines operate at stoichiometric combustion, which is a handicap in terms of efficiency.
In practice, it would be preferable to operate with a lean mixture, that is, at partial load
(2.1.1.3).

The lambda sensor, which analyzes the residual concentration of oxygen in exhaust gas
following combustion, is used to ensure that the engine operates with stoichiometric combus­
tion. If not, the engine's computer will correct the duration of injection so that the amount
of fuel injected is consistent with air flow into the engine. Therefore, knowing the air flow is
also necessary and this parameter is measured by a different sensor (air-flow meter or intake
pressure sensor).

2.2.2.3 Efficiency Breakdown Applied to the Gasoline Engine

The efficiency breakdown (2.1.1.3) applied to the gasoline engine reveals:

- combustion efficiency typically between 0.95 and 0.98, with HC and CO emissions
from the engine (before after-treatment in the catalytic converter) representing a few
g/kWh and a few dozen g/kWh, respectively, that is, between 2 and 5% of the energy
content of the fuel introduced into the engine;

- thermodynamic efficiency on the order of approximately 0.5, by taking into account
compression ratios between 9 and 12, and stoichiometric combustion (the order of
magnitude of the thermodynamic coefficient, γ, is 1.3);

- the values for cycle efficiency vary considerably; measured experimental values range
from 0.1 (low load) to 0.9 (high load);

- mechanical efficiency also varies considerably with load - it is zero, for example, at
idling speed, when combustion is used primarily to compensate for friction and the
consumption of auxiliary components, but exceeds 0.9 at full load.

Figure 2.11 shows an example of an efficiency map for a conventional gasoline engine.
With a value of 0.35, the maximum efficiency of this engine corresponds to a specific fuel
consumption of approximately 240 g/kWh. For operating points that produce on the order
of 3-5 kW at around 2,000 rpm, typical of operation in city driving at constant speed, the
efficiency drops to approximately 0.2. At idling speed, engine consumption rises to approxi­
mately 0.5 L/h.

80 Hybrid vehicles

Figure 2.11
Example of an efficiency map for a conventional gasoline engine.
Displacement 1.3 L. The map shows iso-efficiency curves (black) and iso-
power curves (brown). At the top of the map, the engine experiences combus-
tion; at the bottom of the map (negative torque), combustion has stopped and
engine braking occurs, which corresponds to a situation where the vehicle is
decelerating or traveling down a slope.

2.2.3 Technological Advances
In wide use and capable of further progress, the engine described above corresponds to
the simplest case generally found in common gasoline-powered vehicles. Technological
advances have helped improve spark-ignition engines by increasing performance and effi-
ciency. The technologies presented below are already employed in some modern vehicles.

Three principal approaches are currently in use:
- downsizing, associated with supercharging;
- variable valve train;
- charge-diluted combustion and, especially, stratified combustion.

2.2.3.1 Downsizing
Downsizing (corresponding to the reduction of the engine's displacement), by eliminating
the need for the spark-ignition engine to operate at low load, is a good way to increase effi-
ciency given that, at low load, specific consumption is high (Figure 2.12).

Chapter 2 · Internal Combustion engines 81

Figure 2.12

Illustration of downsizing to improve efficiency. Theoretical comparison of
1.6 L and 1.1 L engines in the same vehicle.
The figure compares the engine 's operating points (red circles) for the same
vehicle, originally equipped with a 1.6 L displacement engine (a), which has
been replaced with a 1.1 L engine1 (b). For the same suppliedpower of 5 kW at
2,000 rpm in the example, the 1.1 L engine consumes approximately 16% less
fuel at this operatingpoint than the 1.6 L engine. Its efficiency is 0.25 compared
to 0.21 for the larger engine.

1. Assuming that both engines have the same efficiency map as a function of
MEP, which, to a first approximation, is valid. Here, we have arbitrarily used
the efficiency map shown previously in Figure 2.11 for a 1.3 L engine. In real-
ity, the reduction in displacement is often accompanied by a (slight) loss of
efficiency for a given MEP.

82 Hybrid vehicles

At a given level of power delivered by the engine in response to the driver's action, the
smaller engine will have to work under a greater load, which is inversely proportional to the
decrease in displacement5. Figure 2.12 shows that the engine then operates in those power
zones that provide a better efficiency. This technique is currently known as downsizing, with
reference to the reduction in engine size. It is an effective means of reducing consumption in
urban areas, where the power demand is relatively low. (See Table 1.1. In the example given
in Chapter 1, the average power requested throughout the Artemis urban cycle is 6.4 kW.)

Most of the time, this technique is accompanied by supercharging air fed to the engine to
compensate for the loss of torque and power that the reduction in displacement would other-
wise lead to. By increasing pressure in the engine's air intake manifold whenever necessary
- that is, at high load - supercharging increases the amount of air the engine is capable of
enclosing in the combustion chambers. This leads to the combustion (and, thus, the introduc-
tion) of a greater amount of fuel and, ultimately, greater power. The maximum MEP that can
be supplied by the engine is greater and compensates for the reduced size.

Compressing air requires a significant amount of energy, sometimes as much as 10% of
the energy supplied by the engine. That is why, in most cases, supercharging is handled by
a turbocharger (a centrifugal compressor and a centripetal turbine - Figure 2.13), which has
the advantage of taking the energy necessary for its operation from the engine's exhaust sys-
tem, where the turbine is located. This energy is, therefore, relatively "free." However, when
there is a sudden demand for power, a short amount of time is needed for the exhaust gases to
transfer sufficient energy to the compressor through the turbine. This turbo lag can become
a handicap. Similarly, the presence of the turbine increases the pressure in the exhaust gas
system and can disturb gas flow, in some cases resulting in acoustic interactions among the
cylinders and an increase in the amount of residual burned gases in the combustion cham-
bers, which can harm engine efficiency. Nonetheless, the turbocharger is a highly effective
technology, and its drawbacks in no way diminish the gains it provides when downsizing
the engine. Considerable research and development effort today is directed at keeping these
disadvantages to a minimum.

Other, less common, mechanisms exist, such as mechanically driven compressors. These
operate by drawing energy from the engine's crankshaft, which affects its final efficiency and
is their principal handicap. However, their reaction time is very short. Combining the advan-
tages of a turbocompressor and a mechanically driven compressor, some recent engines use
them together. Although complex, they result in significant reduction in engine displacement
together with greatly reduced fuel consumption.

The maximum MEP of strongly downsized engines can reach high values of as much
as 25 bar, for a specific torque of close to 200 N.m/L (2.1.1.3). It goes without saying that
control of combustion is especially difficult here to the extent that the engine tends to race.
Consequently, the resistance of the combustion chambers to knock is especially important.
Generally, this involves optimization of the engine's exhaust phases so as to minimize the
proportion of residual burned gas at high load. So-called "scavenging" techniques in the

5. Assuming that the reduced-displacement engine operates at the same rpm as the original engine. In
reality, we may want to choose "shorter" transmission ratios for the "smaller" engine, which results in
slightly higher engine rpm.

Chapter 2 · Internal Combustion engines 83

combustion chambers are sometimes also used. Using compressed air from the supercharger
to eliminate burned gases from the previous combustion step, they provide the ability to
nearly completely empty the dead volumes in the combustion chambers in extremely short
time, on the order of a few milliseconds. They require the use of variable valve timing and,
frequently, the direct injection of gasoline.

Figure 2.13

Twin-scroll turbocharger. The twin-scroll limits interactions of the cylinders
through their exhaust, which improves engine performance. In light gray, on
the right, the compressor stage casing; in the center, the central casing with
lubrication and cooling circuits. In dark gray, on the left, the turbine casing with
the two scrolls side by side.
Source: Honeywell, Wastegate Twinscroll Turbo. Photo, Alain Ernoult

With the increase in drivetrain electrification, there has been renewed interest in turbo-
chargers whose rotors incorporate a small reversible electric machine. Research efforts are
directed at developing such systems, which could be used to recover any surplus energy
in the engine's exhaust gases that hasn't been consumed by the centrifugal compressor
and, conversely, to assist the turbocharger during periods of rapid acceleration to limit its
response time. The difficulties lie in designing an electric machine that can sustain a particu-
larly severe thermal environment and high rpm (typically, a turbocharger rotates at 150,000
to 200,000 rpm). Therefore, this electric machine must be combined with the appropriate
electronics and the vehicle's electrical architecture must include a system for storing and
supplying the electrical energy produced or consumed by the electric machine.

84 Hybrid vehicles

2.2.3.2 Variable Valve Train System
A. Variable Valve Train Systems
The train system is responsible for controlling the opening and closing of the valves. These
systems have undergone extensive development over the years, including:

- camshaft phasers that can angularly offset the valve's lift by a few dozen degrees with
respect to the crankshaft, thereby controlling when the valves open and close

- variable valve lift systems that can be used with relatively wide valve openings
- a combination of these two systems
Cam phasers have become fairly commonplace. Current developments show that, in time,
we will have mechanisms without camshafts {camless engine), where control of valve open-
ing and closing and valve lift will be extremely flexible. Valve control will then be electro-
hydraulic, or even electromagnetic. Here as well, the electrical aspect becomes significant
and the energy consumption of these systems must be considered when evaluating any even-
tual efficiency gains.

Figure 2.14

Camless system from Valeo. In this valve train system, the valves are con-
trolled electromagnetically.
Source: Valeo

Variable timing of valve opening and closure, provided by a "basic" cam phaser, allows
for better use of the acoustic waves present in the engine's intake system. When the wave
arrives at the appropriate moment during the engine's intake phase, it improves the supply of
air to the cylinders. We then speak of acoustic tuning. Without a camshaft phaser, acoustic
tuning is effective for only a small number of engine speeds. With a cam phaser, acoustic
tuning can match a wide range of rpm. The engine's maximum torque curve is then "fuller."

Chapter 2 · Internal Combustion engines 85

The performance gain obtained with a cam phaser can be converted into a slight decrease in
engine displacement or a reduction of transmission gear ratios (in order to decrease engine
speed), leading in both cases to improved fuel consumption in the vehicle.

Fine control of valve timing can also help control the flow of burned gases from the
exhaust pipes to the combustion chambers, or even the reuse of burned gases in the intake
manifold through the combustion chambers. We then have charge-diluted combustion, which
promotes increased efficiency, as described below.

In a conventional engine, the height and duration of valve lift is fixed, being determined
by the profile of the cam. For the intake valves, the height is designed to provide the required
air flow to the engine when operating at maximum power. Variation in intake valve lift by
means of a suitable system (for example, multicam system) can help modify valve opening
for the air flow required at a specific operating point. In this way, at low load, the intake
valves can be opened very little, allowing only a trickle of air to pass through, which avoids
the need to close the throttle to control reduced feed with high pressure drop. Cycle effi-
ciency is improved and pumping losses are reduced (2.1.1.3 and Figure 2.15).

Figure 2.15

Control of intake air through variable valve train.
1. Intake valve is opened.
2. Intake valve is closed.
The diagram provides an overview of the change in pressure in the cylinder
during the air intake phase. In the variable valve actuation engine, the intake
valve is closed when the piston is midway through its travel; the end of the pis-
ton's downward stroke occurs when the valves are closed. In this way, pumping
losses are reduced.

Here too, the problems of speed and energy consumption are key and must be taken into
account when developing new systems. Valves move in a very short period of time. Consid-
erable power is needed to actuate the valves when they open and the technology ultimately
used must not counteract the anticipated gains in efficiency. Considerable research efforts are
underway to develop systems that can control such limitations.

86 Hybrid vehicles

B. Valve Timing and the Miller Cycle

Above we saw that by delaying the moment when the intake valves close with the piston at
bottom dead center, cylinder filling can be modified (Figure 2.15). Another consequence is
the reduction of the engine's actual compression ratio during the compression phase. From
the point of view of the air admitted into the combustion chamber, the compression ratio is
then no longer equal to the engine's "geometric" compression ratio 6. The actual compres-
sion ratio becomes equal to the ratio between the cylinder volume when the intake valves
close and the cylinder volume at top dead center. This effect can be obtained:

- by closing the intake valves before bottom dead center; this is known as the Miller cycle
- or by closing the intake valves late; this is known as the Atkinson cycle

The laws of thermodynamics show that the efficiency (Eq. 2.5) is barely affected by
this drop in compression ratio, the remaining expansion ratio being equal to the geometric
compression ratio. On the other hand, the risk of knock is reduced because the combustion
mixture is less compressed and not as hot during the compression phase. In a gasoline engine,
this type of cycle allows for better control of the trade off between high efficiency and the
risk of knock.

The use of early or delayed closure of the intake valves reduces cylinder air-feeding. If
the valves close early, the time allotted to air intake is shortened; if they close late, some
of the air that had been admitted into the cylinder is pushed into the intake ducts when the
piston begins to move up while the intake valves are still open. When the engine is at full
load, this results in a loss of torque and power. The presence in the valve timing system of
a camshaft phaser can help to improve performance (the cam phaser can be adjusted so that
the intake valves close at bottom dead center) or efficiency (by transitioning to the Miller or
Atkinson cycle). The lack of air at full load can also be compensated for by the addition of a
supercharger. In the Prius, Toyota made use of the Atkinson cycle and the engine is assisted
by a hybrid power system (Chapter 5).

2.2.3.3 Charge-Diluted Combustion

Diluting the air and gasoline charge with inert gas can help improve engine efficiency at
partial load, primarily by reducing pumping losses and thermal loss in the combustion cham-
bers. At high loads, when a large amount of fuel must be burned to liberate a great deal of
energy, the technique is not used, for it would reduce engine torque and power. Charge-
diluted engines, however, provide two modes of operation: they employ charge dilution at
low loads and return to conventional combustion - without dilution - under high loads. Sev-
eral techniques are used to dilute the charge: lean mixtures, exhaust gas recirculation (EGR),
or charge stratification.

6. Recall that the compression ratio is the ratio between the maximum volume of the cylinder (at
bottom dead center) and the minimum volume of the cylinder (at top dead center) (2.1.1.3).

Chapter 2 · Internal Combustion engines 87

A. Lean Mixtures and EGR

The charge can be diluted either by supplying excess air 7 or by using burned gas from the
previous combustion step, sometimes a combination of the two. If the diluent is atmospheric
air, the engine is said to be a lean burn engine; if burned gas from the previous combustion
step is used, the engine is said to operate with exhaust gas recirculation (EGR). Direct injec­
tion stratified-charge engines often use both excess air and EGR.

At low load, for a given power level request from the driver, the partial filling of the
cylinders with gas that cannot participate in chemical reactions during combustion, forces
the throttle to open, thereby admitting an amount of air - and, therefore, fuel - that will pro­
vide the stoichiometric quantities found in ordinary combustion. This additional opening of
the throttle causes a drop in pumping losses and an improvement in cycle efficiency. At the
same time, dilution reduces the temperature of the gases in the cylinders during combustion
(the inert load serves as a heat sink). The temperature drop reduces thermal loss at the walls,
which also promotes the improvement of cycle efficiency.

Another positive effect of dilution is the improvement of the thermodynamic coefficient,
γ, of the gases (2.1.1.3). It can be shown that the presence, in the mixture of gases, of small
molecules, mostly diatomic (air, for example, which consists of N2 and 02) or triatomic mol­
ecules (recirculated burned gases, consisting of N2 and C 0 2 and H20), increases the value of
γ and, thus, thermodynamic efficiency.

Lean burn engines that use air dilution are in a better position from this point of view than
those that incorporate EGR technology. However, EGR provides the advantage of dilution
without the need to modify the density ratio between gasoline and air. Stoichiometry can then
be maintained, in spite of the dilution, which is not the case with an excess of air. This means
that engines equipped with EGR can use an emissions control system employing a conven­
tional three-way catalytic converter (2.2.2.2). Lean burn engines, however, cannot control
ΝΟχ emissions with a three-way catalyst because exhaust gases contain air in excess of the
stoichiometric balance. Consequently, they must also incorporate a dedicated after-treatment
device for nitrogen oxides.

Up till now we've been describing what is known as homogeneous dilution - the diluent
and the stoichiometric air and gasoline charge are uniformly mixed. Nonetheless, dilution is
possible only to a limited extent. If the air-gasoline mixture is too dilute, it is no longer within
the flammability limits and combustion is impossible. For lean burn engines, it is technically
difficult to go below mixtures on the order of 0.65 equivalence ratio. When recirculating
burned gases, the introduction of EGR ratios greater than 30% (by volume) is extremely
difficult.

The use of dilute charges in gasoline engines is the subject of considerable research.
The goal of this work is to develop engines that can operate with increasingly higher dilu­
tion ratios while maintaining stable combustion, that is, repeatable from cycle to cycle,
which is the sign of good drivability for the vehicle and good control of pollution emissions.
This work is primarily directed at creating aerodynamic movement within the combustion
chambers favorable to the development of the flame front in dilute mixtures, using ignition

7. In other words, "air in excess of the stoichiometric balance."

88 Hybrid vehicles

technologies that can extend the limits of flammability, or at extending the operating range
for this type of combustion. The use of recirculated burned gas is still possible at high loads
in downsized engines, in which we can compensate for the volume occupied by inert gas in
the cylinders with increased pressure charging.

B. Stratified Combustion
In extending dilution limits, it's possible to use what's known as stratified combustion. This
technique can be used to produce a flammable mixture near the spark plug and surround the
flammable charge with a highly dilute mixture in the remainder of the combustion chamber.
We then have what's known as a heterogeneous mixture. It then becomes possible to obtain
combustions with an average equivalence ratio in the chamber as low as 0.1. Here, too,
stratified engines can operate in different modes. Stratified combustion is used at low load.
At mid-load, the engine generally swings toward homogeneous combustion. This can be
achieved with a lean-mix for intermediate loads, a stoichiometric mixture for higher loads,
sometimes even a rich mixture at very high load (Figure 2.16).

Figure 2.16
Map of operating modes of a stratified charge engine. The engine is a mass-
production 1.8 L Mitsubishi GDI direct injection gasoline model using strati-
fied combustion. At low load and low rpm, there is an operating region for the
stratified-charge engine (green). If the rpm increases, the engine then moves to
a homogeneous lean mixture combustion mode (yellow). If the load increases,
stoichiometric combustion takes place (pink). At high load, an enriched mix-
ture of fuel is used to produce the maximum torque and power and to control
the temperature of the exhaust gases (beige). Finally, at low rpm and full load,
a dual-injection technique is used to reduce the engine's tendency to knock
(blue).

Chapter 2 · Internal Combustion engines Next Page
89

Engines that make use of stratified combustion use direct injection systems (gasoline
direct injection) (Figure 2.17). The injectors feed into the combustion chambers and the fuel
is introduced directly into the cylinder in the form of a cloud of very fine droplets. For strati­
fied combustion, the fuel is injected belatedly, that is, shortly before the spark appears. The
droplets do not have time to disperse throughout the entire combustion chamber and the
equivalence ratio of the mixture is greater near the spark plug. Conversely, for higher loads,
with homogeneous charge throughout the chamber, injection occurs early, as soon as the
exhaust valves close. The droplets then have time to disperse throughout the entire chamber
and mix uniformly with the air.

Figure 2.17

Cutaway of a Mercedes-Benz 3.5 L CGI direct injection engine with stratified
combustion equipped with piezoelectric injectors.
Source: Daimler-Benz

Although direct injection engines have been used in industry since the mid-90s - the response
of the Asian automobile industry to the widespread introduction of common rail injection tech­
nologies in diesel engines in Europe - research in this especially complicated field is ongo­
ing. The focus today is on ways of creating increasingly efficient and stable stratification in a
combustion chamber, where the air motion is violent and turbulent, and, therefore, random. A
number of questions also remain concerning the genesis of pollutants in heterogeneous mixture
combustion. A compromise will need to be made between rapid combustion, before the droplets
have been dispersed throughout the combustion chamber, to the detriment of stratification, and
the time needed for the droplets to vaporize. If the flame encounters a liquid droplet, combustion
will be locally very rich, and unburned products or even particulates will be produced.

The anticipated strengthening of antipollution standards could also necessitate the use of
particle filters on gasoline direct injection engines, as is already the case with diesel engines.
The exhaust system on stratified engines would then become extremely complex, with a
three-way catalytic converter, particle filter, and ΝΟχ trap. After-treatment of NOx cannot
be handled by a three-way catalytic converter when operating with lean mixtures, and the use
of ΝΟχ trap systems then becomes necessary.

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2.2.3.4 Other Refinements

Even without taking advantage of the benefits of stratified combustion to improve efficiency,
the use of gasoline direct injection can be seen as an advantage or a necessity. Direct injec-
tion is required to implement the "scavenging" combustion found in some highly downsized
engines (2.2.3.1). It's true that some designs could eliminate this in the future, but they are
only in the development stage [Anselmi et al., 2007]. In terms of benefits, the vaporization
of gasoline droplets is accompanied by cooling of the surrounding environment, because the
phenomenon is endothermic. In the case of direct injection, because vaporization takes place
within the combustion chamber, it is the air that enters the chamber that is cooled. It then
becomes denser and, for a given displacement, the engine's air capacity will be increased.
This can raise performance levels - measured on the maximum torque and power curve - by
5 and, sometimes, 10%.

Cylinder deactivation is another technique that is sometimes used in large engines to
improve overall efficiency. The technique is even more effective when the vehicle's speed is
low. There are V8 and VI2 engines in which a cylinder bank (that is, every other cylinder)
is off at low load. All the pistons continue their alternating motions but only half of them
"experience" combustion. All the cylinders ignite whenever there is a power demand from
the driver. Recently, Honda has even applied this principle to a V6 engine, capable of oper-
ating on 3, 4, or 6 cylinders [Fujiwara et al, 2008]. Research is ongoing and, one day, we
should be able to apply deactivation to 4-cylinder engines [Ternel et al, 2009]. At low speed,
the vehicle would operate on only two cylinders.

Going forward, different mechanical systems should enable us to develop engines with
variable compression ratios. These mechanisms will enable us to modify the compression
ratio even when the engine is running. In this way, we will be able to increase the rate of com-
pression during engine operation at low load to promote good thermodynamic efficiency. At
high load, when spark advance would normally have to be adjusted to prevent knock, the
compression ratio would be reduced, which will avoid the problem without penalizing cycle
efficiency.

2.2.3.5 The Future: Controlled Auto-Ignition Combustion?

In two-stroke engines (2.4.2.3), intake and exhaust phases are simultaneous. There is an over-
lap phase, a period during which the exhaust ports are not yet closed, while the intake ports
are still open. Under these conditions, the pressure of the intake manifold must be greater than
that of the exhaust manifold so that intake air enters the cylinder and pushes burned gases into
the exhaust system, and so that exhaust gases from the combustion chamber are prevented
from entering the intake system. In practice, this condition is met by generating overpressure
in the intake manifold, either by means of an external compressor or by means of a crankcase
compression. However, even with such a system, operation at low load is problematic because
the intake throttling used to adjust the load tends to eliminate this overpressure on the intake
side. The engine then operates with air-fuel mixtures that have been highly diluted by burned
gases produced during the previous cycle and poorly evacuated toward the exhaust. In general,
this high dilution leads to slow combustion and a lack of engine stability.