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ENERGY AVAILABILITY STUDY FOR A REGENERATIVE HYDRAULICALLY ASSISTED TURBOCHARGER
ABSTRACT Engine downsizing and down-speeding are essential to meet
future US fuel economy mandates. While turbocharging has been
a critical enabler for downsizing, transient boost response
performance remains a concern even with variable geometry
turbochargers. This slow build-up of boost and hence torque is
commonly referred to as turbo-lag. Mitigation of turbo-lag has,
therefore, remained an important objective of turbocharger
performance enhancement research. A regenerative,
hydraulically assisted turbocharger is one such enhanced
turbocharging system that is able regulate the turbocharger
speed independent of the available engine exhaust energy. With
external power available on the turbocharger shaft, the engine
performance and emissions can be managed during both
transient and steady-state operations. The key to fully utilizing
the ability of such an assisted turbocharger depends on the
energy recovered from turbocharger shaft and/or vehicle
driveline. Energy available from the turbocharger shaft is
dependent on the engine exhaust gas energy. Energy recovered
from the driveline depends on vehicle braking energy. A
previously developed high-fidelity 1-D simulation of a diesel
engine with a regenerative-hydraulically assisted turbocharger
is used to investigate the energy availability for a medium duty
Dr. Tao Zeng is with DENSO International America. This work was completed when he was a Ph.D. student at Michigan State
University.
diesel engine over standard driving cycles. The study shows that
the energy recovery from turbocharger shaft is limited and
driveline energy recovery is necessary for achieving fuel
economy benefits on the order of 4%.
REGENERATIVE HYDRAULICALLY ASSISTED TURBOCHARGER (RHAT)
The hydraulically assisted turbocharger has been studied
since the early-1980's [6]. There have been publications and
patents on the use of a hydraulic turbine, driven by high-pressure
oil, on the turbocharger (TC) shaft to accelerate the turbocharger
[2-5]. The hydraulic turbines, in these systems, are compact
designs and are integrated into the turbocharger center housing
between the conventional compressor and turbine wheels. These
designs, however, relied on a standalone hydraulic pump to build
and maintain high pressure in the hydraulic accumulator thereby
incurring a high equivalent fuel economy (FE) penalty. This
may have been an important reason for the design not achieving
wide acceptability. In this study, an alternative design is adopted
where two energy recuperation devices are used. A hydraulic
pump is mounted on the TC shaft, just like the hydraulic turbine,
and is used to provide energy recovery through TC braking much
like the regenerative function of an electrically assisted TC. An
Tao Zeng Department of Mechanical Engineering
Michigan State University East Lansing, MI, USA
Yifan Men Department of Mechanical Engineering
Michigan State University East Lansing, MI, USA
Devesh Upadhyay Ford Motor Company Dearborn, MI, USA
Guoming Zhu Department of Mechanical Engineering
Michigan State University East Lansing, MI, USA
Proceedings of the ASME 2018Dynamic Systems and Control Conference
DSCC2018September 30-October 3, 2018, Atlanta, Georgia, USA
DSCC2018-9134
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additional driveline pump is used to recover vehicle braking
energy similar to an integrated starter generator (ISG) [1]. While
much of the recovered energy is from the driveline pump, the TC
shaft mounted hydraulic pump provides the important additional
ability to brake the TC actively. Hence the hydraulic turbine and
pump provide the ability for bi-directional speed control of the
TC. This allows benefits such as better surge control and the
ability to operate the compressor optimally. The high-speed
turbo-pump, like the hydraulic turbine, is a mature technology
developed within the aerospace industry [8]. NASA published a
report in 1974 indicating that a hydraulic turbo-pump, for rocket
applications, can achieve efficiencies on the order of about 73%
[8]. Other studies (including research at Ford Motor Company),
also indicated around 72% efficiency on larger sized hydraulic
turbines [1, 9, 10]. Two types of energy recovery are typically
possible. The first mode involves the TC shaft mounted
hydraulic turbo-pump. The most natural opportunity is during a
tip-out situation when the TC kinetic energy is converted to
hydraulic energy by braking the TC by engaging the hydraulic
turbo-pump. This form of energy recovery is considered as free
energy. Alternately, energy recovery is also possible during
engine firing mode by actively managing the speed of the TC by
braking via the pump and over-speeding via the VGT vane, such
that the desired TC speed is maintained. This form of energy
recovery is not free and has an associated FE penalty. Firing
mode energy recovery without fuel penalty is possible but is
limited to very high engine loads, as discussed in [7]. The second
method of energy recovery uses a vehicle driveline mounted
pump. The driveline pump recovers energy from the vehicle
driveline during vehicle brake events. This process is similar to
the regeneration mode of the hydraulic hybrid vehicle [11, 12,
13]. However, unlike the hydraulic hybrid vehicle, which must
launch the full vehicle, a much smaller hydraulic tank is needed
for the RHAT system. The RHAT system nevertheless must have
a dedicated hydraulic circuit that includes high and low-pressure
accumulators and associated fast acting valves. A schematic of
the RHAT system is shown in Figure 1, and design details are
included in the patent [1]. A brief description of the sequence of
events is described. Whenever the vehicle or engine has “free”
energy (e.g. during vehicle or engine deceleration, exhaust
braking, or during steady state when the intake throttle is used
for intake oxygen control, or when a wastegate is used, etc.), the
driveline pump will be engaged to recover vehicle kinetic energy,
while the hydraulic turbo-pump will recover energy from the TC
shaft. The power is collected by both the turbo-pump and the
driveline pump and will pressurize the fluid, and at the same time
slow down the TC to “synchronize” it with the decelerating
engine to avoid “tip-out” surge. The pressurized fluid is routed
through a check valve to a high-pressure hydraulic accumulator.
During engine acceleration, the high-pressure fluid from the
accumulator will be discharged to drive the hydraulic turbine,
which will then accelerate the turbocharger. When the TC turbine
wheel receives the external hydraulic energy, the VGT can be
opened wider for improved turbine efficiency. Thus, the enthalpy
drop across the turbine will be reduced (from reduced exhaust
manifold pressure), which decreases engine pumping loss and
increases net engine power output although with a potential for
reduced high-pressure exhaust gas recirculation (EGR) flow.
Figure 1. System layout of a diesel engine with regenerative
hydraulically assisted turbocharger (above) and integrated
system inside the turbo center housing (below)
RHAT (and other assisted TC’s) can convert exhaust gas energy
and driveline energy into mechanical energy more efficiently,
thus further improving the engine transient response and fuel
economy; the improvement in transient response is partly from
external hydraulic energy, and partly from the improved turbine
and compressor efficiency’s. During very aggressive tip-out
maneuvers, the hydraulic loading on the TC shaft (through the
turbo-pump) will slow down the TC to avoid tip-out surge while
recovering the aerodynamic and kinetic energy from the
turbocharger.
Managing the hydraulic pump and hydraulic turbine in this
manner provides a means to “synchronize” the turbocharger with
the engine operating condition and the boost demand while
ensuring that the compressor and turbine are working in a
narrower but more efficient region. This in turn allows for a more
efficient compressor and the exhaust turbine design for without
sacrificing operating range.
In this study, a 1-D simulation approach [Error! Reference
source not found.] is used to investigate energy recovery
opportunity and capability for regenerative hydraulic assisted
turbocharger system. The paper is organized as follows. First the
simulation environment is described. Then the energy recovery
opportunities are discussed via simulation studies. Finally, the
energy recovery capabilities and fuel benefit are investigated
through vehicle cycle simulations.
The main contribution of this paper is to examine the energy
recovery for regenerative hydraulically assisted turbocharger
system using a system level approach. These simulation results
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would also be a good reference for regenerative electrically
assisted turbocharger systems.
VEHICLE SIMULATION PLATFORM Vehicle level simulation is utilized to investigate the fuel
benefits and design trade-offs for hydraulically assisted
turbochargers. The model structure, as shown in Figure 2,
includes a vehicle model that includes the driveline and an
RHAT equipped engine. Several controllers are used in the
simulation, and they include a production version engine control
with an RHAT controller to replace the conventional VGT-based
air path controller, a 6-speed transmission controller, and a
torque converter controller. The simulation platform is tuned for
vehicles with GVW of up to 10000 lbs. Details can be found in
Error! Not a valid bookmark self-reference..
Table 1. Vehicle information.
Vehicle Weight 10000 lbs.
Engine V8, 6.7 L, Diesel
Turbocharger Variable geometry
turbocharger
Transmission 6-speed
Turbocharger VGT
The base transmission shift strategy, boost pressure set-
point, fuel injection set-point and EGR fraction set-point are
duplicated from the stock engine controller. As mentioned
earlier, RHAT integrated into the air handling control is used to
regulate the boost pressure and EGR rate set-points.
Figure 2. Simulation platform and control structure
Preliminary meanline analyses of the hydraulic turbine and
turbo-pump were conducted. Regular engine oil at 100 °C was
assumed in the meanline analysis. From a friction-loss
perspective, engine oil may not be the optimal choice, due to its
high viscosity. However, the problems associated with sealing
different fluids at different pressures, for isolation purposes,
makes engine oil the best choice at this point. The hydraulic
turbine efficiency supplier maps show that for hydraulic-turbine
powers above 10 kW, a substantial operating area of the
hydraulic turbine can have efficiencies above 70%. Careful
management of flow rate and pressure ratio, allows efficiencies
above 60% at other power ratings. The turbo-pump can achieve
peak efficiencies in the range of 70%, as long as the pressure in
the hydraulic energy storage is managed to match the oil flow
rate at the operating TC speed. The hydraulic turbine, hydraulic
turbo-pump and hydraulic driveline pump were integrated with
a GT-Power vehicle model. The oil temperature was maintained
at 100 °C throughout the test (FTP-75) cycle. Some other
assumptions relating to the RHAT system, as used in the
simulations, include:
1. Hydraulic fluid pressures are 100-150 bar in the high-
pressure accumulator tank, and 10-20 bar in the low-pressure
accumulator tank. The pressure range for the low-pressure tank
was set to avoid cavitation in the hydraulic turbo-pump or
turbine.
2. Hydraulic tank volume, driveline pump and hydraulic
turbine were varied for design trade-off investigation.
Figure 3. Vehicle model validation using FTP-75 drive cycle.
For cycle simulations, a driver model is used to generate the
acceleration and brake pedal positions to allow vehicle speed
tracking. The target engine brake power is based on a calibrated
map as a function of engine speed and gas pedal position. The
fuel demand is a function of engine speed and demanded engine
torque which is in turn a function of the pedal position. Fuel
injection is controlled by both feedback and feedforward loops
to achieve the target torque. The desired boost pressure is
achieved via closed loop control of VGT vane position and
hydraulic assist power. The transmission control is based on a
shift schedule map as a function of engine speed and gas pedal
position. Both high-pressure EGR (HP-EGR) and low-pressure
EGR (LP-EGR) are used to achieve target EGR mass flow rate.
Both VGT vane and EGR valve controllers are map-based gain-
scheduled proportional-integral-derivative (PID) controllers.
Detailed engine and turbocharger model validation results can be
found in [14]. The performance of the vehicle system simulation
200 400 600 800 1000 12000
20
40
60
80
100
Time [s]
Veh
icle
spe
ed
[km
/h]
vehicle model validation
Dyno test
GT simulation
200 400 600 800 1000 12001
2
3
4
5
6
Time [s]
Tra
nsm
issi
on
ge
ar
nu
mb
er
Dyno test
GT simulation
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is shown through the vehicle speed tracking performance of the
target trajectory in Figure 3.
ENERGY RECOVERY OPPORTUNITIES Different energy recovery opportunities, with or without
fuel cost, exist during a vehicle’s normal operation. The
hydraulic turbo-pump on the turbocharger shaft can be taken as
turbocharger shaft load or loss during turbocharger operation. It
is clear from the turbocharger shaft rotor dynamics in Equation
(1), that any hydraulic load on the TC shaft will lead to lower
turbocharger shaft speed and therefore to lower compressor
power.
𝐽𝜔�̇� = 𝑓�̇�𝑇
( 𝑢𝑣𝑔𝑡,𝑃3
𝑃4
, 𝜔, 𝑇3) − 𝑓�̇�𝑐
( 𝑃2
𝑃1
, 𝜔) − 𝑓�̇�𝐿𝑜𝑠𝑠
− �̇�𝑟ℎ𝑎𝑡 (1)
�̇�𝑟ℎ𝑎𝑡 = 𝑓�̇�𝑇
( 𝑢𝑣𝑔𝑡,𝑃3
𝑃4
, 𝜔, 𝑇3) − 𝑓�̇�𝑐
( 𝑃2
𝑃1
, 𝜔) − 𝑓�̇�𝐿𝑜𝑠𝑠
− 𝐽𝜔�̇� (2)
where �̇�𝑟ℎ𝑎𝑡 is the power loss by hydraulic pump; 𝑓�̇�𝑇
is
turbine power; 𝑓�̇�𝑐
is compressor power; and 𝑓�̇�𝐿𝑜𝑠𝑠
is power
loss due to friction.
A shown in Figure 4, it is possible to manipulate the
compressor efficiency and hence the energy demands by
adjusting the TC speed. For example, at light loads, by
increasing (assist) the turbocharger speed from C to D the
compressor efficiency will increase. On the other hand, at high
load conditions with high turbocharger speeds, when the TC
speed is reduced (via regen), the compressor efficiency can be
improved using RHAT, as shown for the load transition from A
to B. Improved compressor efficiency also reduces the
compressor power demand to maintain the same boost response.
This allows adjusting the VGT position to support a lower
turbine work output and contributes to lower pumping losses
hence improved fuel economy.
Figure 4. Compressor efficiency vs. TC shaft load condition.
Diesel fuel shutoff (DFSO) is an effective way to improve
fuel economy during deceleration; see Figure 5. During the
vehicle deceleration, engine fuel is shut off to minimize fuel
consumption. The turbocharger slows down due to reduced
exhaust energy. During free deceleration, the VGT can be used
for exhaust braking by fully closing the VGT vane position.
Alternately the TC can be decelerated via the Turbo pump for
energy harvesting without fuel cost. There are two types of
energy recovery during the diesel fuel shutoff event. One is at
the beginning of acceleration pedal release and the other at the
beginning of brake pedal engagement. The initial engine
operating condition, for these two scenarios, is quite different,
leading to different energy recovery capabilities. When the driver
releases the acceleration pedal, the engine is still operating at
high speed, leading to higher energy recovery level. When the
brake is engaged, the engine speed is typically lower leading to
relatively low levels of energy recovery. Hence engaging the
recovery action at the start of tip-out is preferred.
Figure 5. Energy recovery from diesel fuel shutoff.
Driveline-based hydraulic energy recovery, during vehicle
braking, shares the same principle as electrical energy recovery
in hybrid electric vehicles, where a hydraulic pump is used,
instead of an electric motor, to recover the brake energy.
The energy recovery opportunities and corresponding
energy levels for RHAT system are summarized in Table 2.
Table 2. Regenerative hydraulically assisted turbocharger
energy recovery opportunities.
Event Opportunities Energy level
Energy
recovery
with fuel
cost
Energy recovery
with fuel cost High High
Energy
recovery
without
fuel cost
Steady State Energy
recovery Low Low
Diesel fuel shut-off Low Low
Energy recovery
from exhaust brake Low High
Energy recovery
from driveline brake
energy
High Low
SIMULATION RESULTS AND DISCUSSION
Energy Recovery from Diesel Fuel Shut-Off (DFSO)
Turbocharger energy recovery during DFSO largely
depends on the engine operating condition prior to fuel shut-off
as well as the duration of the fuel shut-off. The calculation of
energy recovery during DFSO is shown in (3).
𝑊𝑟ℎ𝑎𝑡 = ∫ [𝑓�̇�𝑇( 𝑢𝑣𝑔𝑡,
𝑃3
𝑃4, 𝜔, 𝑇3) (𝑡) − 𝐽𝜔�̇� − 𝑓�̇�𝑐
( 𝑃2
𝑃1, 𝜔) (𝑡) − 𝑓�̇�𝐿𝑜𝑠𝑠
(𝑡)]𝑡2
𝑡1
(3)
subject to
Gas Pedal
Release
120 kRPM
115 kRPM
100 kRPM
85 kRPM
60 kRPM
40 kRPM
D
C
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𝜔 > 𝜔𝑚𝑖𝑛
𝑁𝑒𝑛𝑔 > 𝑁𝑒𝑛𝑔,𝑚𝑖𝑛
𝑃3 < 𝑃3,𝑚𝑎𝑥
The constraints on the final energy recovery include terminal
engine and turbocharger speeds, and the allowable maximum
exhaust manifold pressure. The final or minimal engine speed 𝑁𝑒𝑛𝑔,𝑚𝑖𝑛 and turbocharger speed 𝜔𝑚𝑖𝑛 should be high enough
to support the next tip-in without the need for excessive assist
energy. The maximum exhaust manifold pressure 𝑃3,𝑚𝑎𝑥 is
necessary for engine exhaust gasket protection. Different cases
are investigated based on different engine initial speeds (1000-
3000 rpm) and different initial turbocharger speeds (20k-100k
rpm); see Figure 6. All studies are based on the same final
condition (Engine speed = 800 rpm, Turbocharger speed = 15k
rpm). Different hydraulic loading levels are used for energy
recovery. The loading torque applied as hydraulic pump load on
the TC shaft is as high as 3 N-m. Energy recovery surfaces show
the maximum turbocharger kinetic energy recovery for this
engine is around 13 kJ for a single event. With respect to the
terminal constraints, the optimal hydraulic turbo-pump can be
sized based on these simulation results.
Figure 6. Energy surface for different initial conditions.
Figure 7. Constraints under different initial conditions
Figure 7 shows the allowed loading torque subject to terminal
engine speed and turbo speed constraints, respectively, under
different initial engine and turbocharger speeds. With high initial
TC speed, larger loading torque is allowed for energy recovery.
Similar conclusion can be drawn from initial engine speed
results. To meet the constraints of both terminal engine speed and
TC speed, the upper right region formed by the intersection of
engine speed and TC speed constraints under the same initial
engine speed provides the feasible values for loading torque.
Figure 8. TC speed and engine speed distribution for FTP-75.
Figure 9. TC speed and engine speed distribution for US-06.
From the above energy analysis results, two driving cycles (FTP-
75 and US-06) are considered, the distribution for engine speed
and turbocharger speed are shown in Figure 8 and Figure 9. Both
initial conditions are given for driver acceleration and brake
pedal release. FTP-75 has more tip-ins and tip-outs compared to
US-06, leading to more opportunities for energy recovery.
Higher TC kinetic energy recovery is possible during tip-out
with higher engine speed and TC shaft speed. Energy recovery
from turbocharger shaft can be estimated based on the initial tip-
Engine speed=2500 RPM (turbo speed constraint)
Engine speed=2000 RPM (turbo speed constraint) Engine speed=1500 RPM (turbo speed constraint) Engine speed=1000 RPM (engine speed constraint)
Engine speed=1500 RPM (engine speed constraint) Engine speed=2000 RPM (engine speed constraint) Engine speed=2500 RPM (engine speed constraint)
Engine speed=3000 RPM (engine speed constraint)
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out conditions and energy recovery surfaces in Figure 6. The
estimated energy recovery can then be used for sizing the
hydraulic turbine and hydraulic pump as well as the driveline
pump for a given vehicle. The estimated energy recovery from
turbocharger shaft for both FTP-75 and US-06 are shown in
Figure 10.
Figure 10. Tank energy balance for energy recovery from
turbocharger shaft.
Driveline Energy Recovery and Premiliary Drilveline Pump
Sizing
The driveline brake energy available over FTP-75 and US-
06 drive cycles and the corresponding braking durations are
shown in Figure 11. The brake power in the top plots are
calculated from the highest power available during each brake
event. Only brake pedal engagement durations longer than 4
seconds are considered. The energy distribution for driveline
recovery shows the energy recovery for each individule tip-out.
The maximum braking power can be as high as 180 kW (FTP-
75) and the minimum power is 48 kW (US-06). Driveline energy
recovery opportunties from FTP-75 are more than that from US-
06 cycle.
Figure 11. Brake power distribution over drive cycles, where x-
axis shows the number of brake event and the top plots show
the power of driveline recovery.
In order to properly size the hydraulic driveline pump and
hydraulic accumulator, a sweep study for fixed VGT position
(equivalent to a fixed geometry turbocharger (FGT)) and
driveline pump power is investigated; see Figure 12. Each tip-in
is followed by a tip-out. In order to have a small tank size, the
energy deficit from hydraulic turbine free energy recovery, must
be compensated by energy recovery by the driveline pump, in
order to maintain a balanced hydraulic energy state of charge
(SOC). The hydraulic turbine must provide different levels of
assist power during tip-in based on the VGT vane position
setting. Under these considerations a properly selected hydraulic
driveline pump size is used in cycle simulations to validate the
system component sizing and the associated fuel benefits.
Figure 12. Driveline pump sizing
Figure 13. Energy recovery comparison
The energy recovered during diesel fuel shut-offs and energy
recovery from driveline pump are compared in in Figure 13. It is
0 10 200
50
100
150
200
Barke event
Pow
er
[kW
]
FTP 75
1 2 3 4 5 6 7 80
50
100
150
200
Barke event
Pow
er
[kW
]
US 06
0 10 200
5
10
15
Barke event
Tin
e [
s]
FTP 75
1 2 3 4 5 6 7 80
5
10
15
Barke event
Tim
e [
s]
US 06
Tim
e [s]
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clear that the total energy recovered purely during DSFO events
from the Turbo pump is much smaller than the energy recovery
from the driveline pump. For better fuel economy improvement,
the driveline pump is necessory.
Cycle Energy Balanced Analysis
In this full-cycle simulation study, various fixed VGT vane
positions are used to size an equivalent fixed geometry turbine
design as well as determine the tank energy balance based on
vehicle cycle simulations. Hydraulic driveline pump is sized at
25 kW and is controlled by the driveline pump valve, and a tank
is size of 20 L is considered.
Figure 14. Cycle simulation for different VGT position.
Simulation results are shown in Figure 14. The top plot
shows the cycle vehicle speed trace from 0 to 1850 seconds, the
middle plot shows the tank energy variation with driveline
energy recovery and the bottom plot shows the tank energy
variation without driveline energy recovery. Results clearly show
that for the wider open VGT positions, higher hydraulic assist
energy is consumed in order to meet the boost pressure target,
which is consistent with the results in [16]. The largest energy
drop over a given drive cycle can be used for tank sizing. From
Figure 14 (middle plot), it is clear that the tank energy SOC is
fully balanced at the end of the cycle only for the cases VGT =
0.5 and VGT = 0.65. While smaller VGT positions reduce the
assist energy needed and are therefore more suited for a balanced
SOC, they do not allow high levels of fuel economy benefit as
was also reported in [16,17].
Based on the design validation through FTP-75 driving
cycle simulations shown in Figure 15, energy recovery capability
decreases at the wider open VGT position, leading to larger tank
energy variation during the driving cycle. This also results in the
trade-off between the tank size and fuel economy improvement
for the studied vehicle under FTP-75 cycle; see Figure 16. The
feasible tank size, designed by considering the feasible
packaging size, could lead to around 4% fuel benefit for the
proposed regenerative hydraulically assisted turbocharger.
Figure 15. Energy recovery.
Figure 16. Trade-offs between fuel saving and tank size.
CONCLUSIONS Regenerative hydraulically assisted turbocharger system
with driveline energy recovery is a feasible technology for fuel
economy improvement. By properly sizing the hydraulic turbine,
VGT=0.5+RHAT VGT=0.65+RHAT VGT=0.75+RHAT VGT=1+RHAT0
0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
0.9
1
Norm
alize
d e
ne
rgy
Hydraulic turbine energy used
Hydraulic TC pump recovered energy
Driveline pump recovered energy
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hydraulic pump, and driveline pump, a 4% fuel economy
improvement can be achieved.
The major concern for this technology is energy availability
for the hydraulic turbine. This study shows that the energy
recovered from turbocharger alone is not enough for significant
fuel economy improvement due to insufficient assist energy
available at turbocharger shaft. With the help of driveline pump
energy recovery, the tank energy can be balanced allowing for
aggressive hydraulic assist and imporved fuel economy benefits.
NOMENCLATURE
EGR Exhaust gas recirculation
𝑓�̇� Power function (kW)
FE Fuel economy
J Rotational inertia
N Rotational speed (r/min)
P Pressure (kPa)
t Time (s)
T Temperature (K) TC Turbocharger
u Control input VGT Variable geometry turbocharger
W Energy (kJ) 𝜔 Angular speed (rad/s)
Subscripts:
1 Compressor inlet
2 Compressor outlet
3 Turbine inlet
4 Turbine outlet
C Compressor
Loss Energy loss
T Turbine
REFERENCES
1. Sun, H.H., Hanna, D.R., Levin, M., Curtis, E.W. and Shaikh,
F.Z., Ford Global Technologies, LLC, 2014. Regenerative
assisted turbocharger system. U.S. Patent 8,915,082.
2. Kapich, D.. "Turbo-hydraulic engine exhaust power
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