Energy Availability Study for a Regenerative Hydraulically ... Articles/Ener… · Error! Not a valid bookmark self-reference.. Table 1. Vehicle information. Vehicle Weight 10000

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


Devesh Upadhyay Ford Motor Company Dearborn, MI, USA

Guoming Zhu Department of Mechanical Engineering


Proceedings of the ASME 2018Dynamic Systems and Control Conference

DSCC2018September 30-October 3, 2018, Atlanta, Georgia, USA

DSCC2018-9134

1 Copyright © 2018 ASME

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ichigan State University user on 04 Septem

ber 2019

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

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