An approach to select an appropriate turbocharger for matching with an internal combustion engine

Abstract

The present paper proposes a novel methodology of turbocharging automotive engines to reach targeted performance. The actual method is tested and validated against simulation test results of two turbocharged diesel engines; engine I, three cylinders, 1.5 L, and engine II, six cylinders, 5.9 L. The present procedure is subdivided into four key parts; namely, database construction, selection procedure, turbocharger preliminary design, and engine modeling. Based on geometric dimensions and aerodynamic parameters provided by the preliminary design procedure, 3D geometries of the turbine and compressor are generated for each studied engine. After integrating previous data into a constructed turbocharger database, two turbochargers are selected for the engine I, while only one turbocharger for the engine II. The findings show that, at the engine speed of 4000 rpm, engine I matched with the adequate turbocharger reached a target power about 2.7%, compared to the original turbocharger equipping engine I. Furthermore, engine II reached a rated power of 299.3 kW at 2500 rpm which is slightly under the original one by 2.64 kW. The superimposition of the engine operating area on compressor and turbine maps provided satisfactory results in terms of turbocharger-engine output performance, fuel consumption, secure functioning and engine thermal strength. Finally, the main advantage of the developed methodology consists of its ability to be applied at both earlier and last stages of the engine turbocharging process or to find new adequate turbochargers to replace the original one for economic, mechanical or for safety reasons.

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Standard Article
International J of Engine Research
1–19
ÓIMechE 2022
Article reuse guidelines:
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DOI: 10.1177/14680874221108581
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An approach to select an appropriate
turbocharger for matching with an
internal combustion engine
Mohamed Amine El Hameur
1
, Youcef Sehili
1
, Mahfoudh Cerdoun
1
,
Lyes Tarabet
1
and Giovanni Ferrara
2
Abstract
The present paper proposes a novel methodology of turbocharging automotive engines to reach targeted performance.
The actual method is tested and validated against simulation test results of two turbocharged diesel engines; engine I,
three cylinders, 1.5L, and engine II, six cylinders, 5.9L. The present procedure is subdivided into four key parts; namely,
database construction, selection procedure, turbocharger preliminary design, and engine modeling. Based on geometric
dimensions and aerodynamic parameters provided by the preliminary design procedure, 3D geometries of the turbine
and compressor are generated for each studied engine. After integrating previous data into a constructed turbocharger
database, two turbochargers are selected for the engine I, while only one turbocharger for the engine II. The findings
show that, at the engine speed of 4000 rpm, engine I matched with the adequate turbocharger reached a target power
about 2.7%, compared to the original turbocharger equipping engine I. Furthermore, engine II reached a rated power of
299.3 kW at 2500 rpm which is slightly under the original one by 2.64 kW. The superimposition of the engine operating
area on compressor and turbine maps provided satisfactory results in terms of turbocharger-engine output performance,
fuel consumption, secure functioning and engine thermal strength. Finally, the main advantage of the developed metho-
dology consists of its ability to be applied at both earlier and last stages of the engine turbocharging process or to find
new adequate turbochargers to replace the original one for economic, mechanical or for safety reasons.
Keywords
Database, Diesel engine, engine modeling, selection, turbine performance map, turbocharger matching, preliminary design
Date received: 8 January 2022; accepted: 31 May 2022
Introduction
The increase of environmental regulations and eco-
nomic limitations have become progressively strict,
which obliged the automotive industry to develop a
new design procedure or redesign their products to meet
the better requirements. Turbocharging is still one of
the efficient ways to enhance the Internal Combustion
Engine (ICE) performance.
1,2
A turbocharged (TCed) engine has better fuel econ-
omy, cost efficiency, and power density than its equiva-
lent sized Naturally Aspirated (NA) engine.
3,4
However, matching Turbocharger (TC) to an ICE is
not an easy task to realize.
5,6
It is clear that a turboma-
chine is not ideally suited to operate in combination
with a reciprocating machine,
7,8
therefore the arrange-
ment of an ICE with TC needs to be planned carefully.
The main difficulties remain in the complicated balance
when matching the Radial Inflow Turbine (RIT) to
ICE, Centrifugal Compressor (CC) to ICE, and the
RIT with the CC. Furthermore, when economic, tech-
nical, and environmental aspects are taken into account
the matching between the aforementioned parts
becomes extremely difficult to achieve.
The automotive engine operates at various operating
conditions (speeds and loads), for this reason, TC must
deliver an adequate charge air and pressure to match
engine-breathing requirements at different operating
points. The principal objective of TC matching is to
1
Laboratory of Propulsion and Reactive Systems, Ecole Militaire
Polytechnique, Algiers, Algeria
2
Department of Industrial Engineering, University of Florence, Florence,
Italy
Corresponding author:
Mahfoudh Cerdoun, Laboratory of Propulsion and Reactive Systems,
Ecole Militaire Polytechnique, BP17 Bordj-el-Bahri, Algiers 16046, Algeria.
Email: cerdoun.mahfoudh@gmail.com

maintain the engine operates far from surge and choke
lines of the CC’s map. To reach these objectives, several
works tried to redesign or to propose a new design of
TC compressor to enhance the engine output power,
9,10
improve mechanical sustainability,
11,12
secure working
of the CC (surge margin),
13–16
or to optimize the design
procedure.
17,18
Some researches were focused on the
components of the TC or ICE including inlet and
exhaust manifolds,
19
intercooler,
20,21
the CC and RIT
casings,
22,23
variable turbine geometry,
24
exhaust gas
recirculation,
25,26
or twin-entry radial turbine.
27,28
Due to the TC’s technical complexity and cost, lim-
ited experimental studies were carried out in the match-
ing field.
29,30
Buchman
20
presented a method for
turbocharging single-cylinder ICE. The proposed solu-
tion consists to add an air capacitor, in the form of a
large volume intake manifold, between the TC com-
pressor and the engine intake to smooth out the flow
and to reduce the timing mismatch between the two
components.
Iyer et al.
31
boosted a two-cylinder Direct Injection
(DI) NA engine with waste gated TC to improve the
power and torque output. After selecting the adequate
TC, the engine was upgraded to provide about 80%
and 75% higher peak power and torque, respectively.
Wang et al.
32
studied the TC matching based on vehicle
performance requirements. Where, a simulation model
of a 1.5 L gasoline engine was built in 1D simulation
using GT-Power
Ò
software. From the findings, the
maximum torque and power of the NA engine were
risen from 135 N m and 78 kW to 215 N m and 115 kW,
respectively.
However, assessing the TC and ICE performance by
executing experimental tests is too expensive and should
only be done at the last stage of engine’s design or after
selecting an adequate TC. As alternative, numerous
studies attempted to match TC to ICE, using simula-
tion software tools or 0D and/or 1D programs assisted
by empirical correlations. The first matching methods
are common techniques namely, trial-and-error meth-
ods where several TCs are selected and tested on an
ICE.
33–35
Considering the second matching method which
consists of developing analytical approaches based on
0D and/or 1D codes supported by empirical correla-
tions.
36–38
The main advantage of these methods is their
rapidity and flexibility which allow several investiga-
tions impossible to carry out experimentally. For this
type of approach, Sanaye et al.
14,37
proposed a new
method where the compressor and turbine parts besides
the engine were analyzed and optimized, the optimum
selected TC had the minimum compressor and turbine
losses. Moreover, the operating points of the TCed
engine had sufficiently large distances from surge and
choke lines.
The trial-and-error method is still the most used
one, due to its simplicity and accuracy. The most disad-
vantage of this method is the unavailability of a TC
map. In addition, this method is time-consuming and
expensive due to the need of performing numerous
empirical tests. While the main disadvantage of the sec-
ond traditional TC matching technic is the dependency
on empirical correlations that may be usually effective
at specific functioning range. Nevertheless, limited
studies proposed an aerodynamic and thermodynamic
design of TC during the matching process
35,39
without
taking into account the matching process between the
TC components.
Indeed, TC manufacturers habitually provide 0D
online codes, which recommend the appropriate TC
among their existing products, based on user-specified
ICE characteristics and targeted performance.
Nevertheless, the results offered by these programs still
not sufficiently accurate and reliable to be used by
automotive constructors or professional individuals
during engine development or tuning steps. Moreover,
during the matching process of TC to ICE, the chosen
TC is one of the series products by the component sup-
plier, which are habitually designed and developed
according to the turbo-machinery experts, but not
designed for the particular ICE.
5,40
Furthermore, the
TCs proposed by TC manufacturers could be not as
suitable ones, and the TCs used have to be subjected to
experimental testing to validate them, which is time-
consuming and expensive procedure.
Also, the use of the internal combustion engine
extrapolated to other sectors rather the automobile
field such as the propulsion of aerial and maritime
drone, in which the turbocharging plays an important
role to achieve the goals. However, the available pro-
cess to select the appropriate turbochargers to be
matched with ICE does not covers these new sectors
which attracts multidisciplinary scientific audiences,
not systematically familiarized with design of the
turbo-machinery components.
To overcome the aforementioned drawbacks, the
present paper presents a new methodology based
mainly on the basic knowledge on the thermodynamics
of ICE and on the turbo-machinery laws. The present
investigation is a comprehensive end-to-end perfor-
mance study starting from the user targeted performance
to the final system integration (turbocharger-engine
matching), based on a series of algorithms that are
detailed within the paper, and standard software tools.
In this aspect, the present work will be interesting for
practicing engineers undertaking turbocharger matching
activities.
Hence, the main contribution of the present paper
can be listed as follows:
1. Regarding the literature review, we have reported
that the common techniques matching methods,
namely, trial-and-error methods, are largely used
in the matching process, in which several TCs are
selected and tested on an ICE. In this area, the
present methodology can at least reduce the num-
ber of the selected TCs to be simulated or
2International J of Engine Research 00(0)

performed experimentally, in order to reduce time
and cost.
2. The developed methdology can be used at both
earlier stage (boosting a naturally aspirated engine
using a turbocharger), and at the last stage (charg-
ing an engine already equipped with a turbocharger
with a new one for reasons of safety, better perfor-
mance, cost, maintenance...etc.) of TC-ICE match-
ing and/or during engine’s design assessment.
3. The present approach combines several in-house
codes, to estimate boundary conditions upstream
and downstream of the engine during the matching
process, an aero-thermodynamic design of suitable
turbocharger components (based mainly on the
Aungier 1D method. This later is validated against
hundreds of compressors and turbines, which can
be considered as a second validation of the overall
methodology).
4. A modified code to obtain the turbine performance
map is presented and validated.
5. The simultaneously design of the radial turbine
and centrifugal compressor is proposed consider-
ing a perfect matching.
The remainder of this paper gives consecutively a detail
description of the proposed methodology mainly the
database construction, selection procedure, turbochar-
ger preliminary design and engine modeling and appli-
cations considering two types of engines.
Methodology
This section is subdivided into four main parts. In the
first one, a detailed construction of Database (DB) is
explained. Next, the selection procedure of an adequate
TC from DB is clarified. Then, a TC preliminary design
based on 1D aerodynamic in-house code of the TC’s
rotors is implemented. Finally, two programs; namely,
Pre-engine model and Post-engine model are implemen-
ted and validated to estimate thermodynamic para-
meters at the ICE intake and exhaust, respectively.
Database construction
Contrarily to DBs created by Sanaye et al.
14
and Emara
et al.
41
the actual one contains around 500 different
types of TCs obtained from several manufacturers’
open sources
42–44
and covering a large mass flow and
pressure ratio ranges; from 0.041 to 1.46 kg, and from
0.26 to 3.74, respectively.
The main difficulties encountered during the DB
construction were due to the lack of information. Some
manufactures provide only the CC maps and not the
RIT maps, others supply only the geometrical para-
meters of turbine and/or compressor, while few of them
supply both geometric parameters and performance
maps. Figure 1 illustrates the mains parameters used to
build the DB such as TC wheel Trims, mass flow rates
_
mC,_
mT, pressure ratios PC,PT, geometric data, and
rotational speed N of the CC and RIT, respectively.
Using geometrical parameters of Figure 1(c) the Trim
of the CC and RIT are calculated as follows;
TrimC=D2
I
D2
E

100 ð1Þ
TrimT=D2
E
D2
I

100 ð2Þ
Where, DIand DEare the inducer and the exducer dia-
meters, respectively, of both the CC and RIT wheel.
To overcome the problem of RIT map unavailabil-
ity, a more precise computer program is developed in
this paper to generate the appropriate maps and it is
based on works of ref.
45
The modified method consists
of estimating the stagnation pressure drop into the
RIT’s rotor at stationary and dynamic state basing on
the first principle of thermodynamic and turbomachin-
ery fundamentals. After combining the two states,
equation (3) is then applied to estimate the RIT total-
static expansion ratio.
PTts =exp 2gtK1V2
adim

ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi
1gtC1_
m2
adim
qð3Þ
Where, gt,C
1,K
1,V
adim and _
madim are gas specific
heat, the K1constant is equal to (r4
4=A2
4), C1is defined
as ½0:5ð1ðr5m=r4Þ2Þ, a speed ratio that equals
U4=ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi
gtRT04
p

and a dimensionless mass flow that
equals ( _
mgaz=½r04 r2
4ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi
gtRT04
p). The correlation (3) as it
is defined does not cover a large size of turbine; then, a
multiplication factor eis specially introduced to con-
verge the numerical results to the experimental data.
The proposed factor is defined by equation (4).
e=1
ffiffiffiffiffiffiffiffiffiffiffiffiffiffi
gtC1
ðÞ
p
!
=_
mgm ð4Þ
where, _
mgm is the maximum of gas flow rate traversing
the RIT’s rotor.
Contrarily to the previously mentioned work, a
RIT’s power balance is applied in this paper to estimate
RIT’s isentropic efficiency hTis , using CC thermody-
namic data provided by CC’s map, as it is shown in
equation (5). The principle advantage of equation (5), is
that there is no need to estimate the TC mechanical effi-
ciency hMech due to estimation of CC power directly
from the experimental data maps provided by TC’s
manufacturers.
hTts =hTis hMech =_
mairCpcT02 T01
ðÞ=
_
mgasCptT04 11
PCts

gt1
gt
!"#
ð5Þ
To validate this method, two different types of RIT
maps were picked from two TCs (namely, Honeywell
MGT1238 and Honeywell GT06Z). The findings show
El Hameur et al. 3

a good agreement between the results of the modified
method and experimental data, where the maximum
deviations in terms of expansion ratio were around
4.11% and 9.03% for the TC MGT1238 and TC GT06,
respectively as depicted in Figure 2.
Selection procedure
A flexible selection procedure of the suitable TCs from
DB is proposed in the present paper. Its starts by turn-
ing the TCed engine into NA one by removing its TC,
then, two main cases are proposed to the user as shown
Figure 1. Main parameters used to define the database: (a) CC performance map, (b) RIT performance map, (c) TC components,
and (d) design point of desired range.
Figure 2. Comparison between experimental and numerical results: (a) expansion ratio and (b) isentropic efficiency.
4International J of Engine Research 00(0)

in Figure 3. Choosing the right case depends on the CC
and RIT maps availability and/or on the accuracy
desired.
Starting from ICE specifications, air conditions, tar-
geted brake power and brake specific fuel consumption
(BSFC), a rapid but less accurate way is then applied in
case 1 to select a suitable TCs. However, to estimate
the required mass flow _
mCreq and the required pressure
ratio PCreq suitable for the engine to reach the targeted
performance, a Pre-engine model is then developed (see
section 2.4.1). After specifying the desired CC tolerance
TolC, equations (6) and (7) are then used to establish a
desired range using mass flow rate _
mCDP and pressure
ratio PCDP at the design point of each TC in DB, as is
illustrated in Figure 1(d). Several TC propositions can
be provided by the actual approach, in this aspect, a
minimal weighted function is calculated using equation
(8), where, a compromise of 50% between deviations of
mass flow rate and pressure ratio is considered.
Meanwhile, increasing TolCwill inevitably increase the
number of TCs proposed whereas the accuracy
decreases.
_
mCreq TolC_
mCreq

100 4_
mCDP 4_
mCreq +TolC_
mCreq

100
ð6Þ
PCreq TolCPCreq

100 4PCDP 4PCreq +TolCPCreq

100
ð7Þ
min f
fg=0:5
_
mCreq _
mCDP

_
mCreq
+0:5PCreq PCDP

PCreq
ð8Þ
Due to limited number of TC’s maps supplied by TCs
manufacturers, the case 1 rarely reach suitable solution.
To overcome this problem, a more precise approach is
also proposed (case 2). Nevertheless, a TC preliminary
design is applied after using the Post-engine model (see
section 2.4.2) to predict thermodynamic parameters at
the engine exhaust.
The case 2 compares the TC’s geometrical para-
meters provided from the preliminary design code and
the available geometries of TCs contained in DB (using
TrimCand TrimT) as shown in equations (9) and (10).
Therefore, two weighted functions f1and f2are calcu-
lated and used to estimate a global weighted function
fglobal, which expresses a compromise of 50% between
f1and f2as shown in equation (11).
f1=0:5p
4
D2
1DB D2
1design

D2
1design
+0:5p
4
D2
2DB D2
2design

D2
2design
ð9Þ
f2=0:5p
4
D2
4DB D2
4design

D2
4design
+0:5p
4
D2
5DB D2
5design

D2
5design
ð10Þ
fglobal =0:5f1+0:5f2ð11Þ
where, D1DB ,D
4DB and D2DB ,D
5DB express the intake
and exhaust compressor and turbine diameters esti-
mated from DB, respectively. D1design,D
4design and
D2design,D
5design are the intake and exhaust compressor
and turbine diameters provided by the preliminary
design code.
Figure 3. Flowchart of the selection procedure used to boost an ICE.
El Hameur et al. 5

Turbocharger preliminary design
Limited works have been found in open literature that
treated a complete TC design,
46,47
due to the complex
balancing between the TC’s component. Therefore, a
1D aerodynamic in-house code is implemented in the
actual work to generate easily and quickly geometries
of TC’s rotors, from which further analysis can be
proceeded.
The input data of the TC’s preliminary design code
are picked from various traditional conception works
(see Table 1). Indeed, the design process of the RIT
and CC is based on a modified Aungier method.
48,49
The design performance are estimated during the anal-
ysis process using a 1D-gas flow and losses models.
Firstly, the RIT rotor is established using output data
of Post-engine model ( _
mT,P
4,T
4), turbine specific
speed ns, air ambient conditions (T1,P
1,dair,Cp
air)
and the CC targeted performance supplied by Pre-
engine model ( _
mCreq ,PCreq ,P
2,T
2,hCtt ). The RIT’s
expansion ratio was estimated applying energy balance,
knowing PCtt,T
02 and hMech (see equation (29))
instead of applying an iterative calculation, which
greatly reduces time calculation comparatively to the
proposed Aungier method.
48
The RIT’s rotor geometry
can be determined applying the assumptions made on
the velocity triangle (see Figure A1 on Appendix A),
the aero-thermodynamic equations, and key equations
(12)–(25) listed in Table 2. Contrarily to traditional
methods that balance between the CC and RIT using
only mass flow and velocity relations, hence, the devel-
oped code uses four balance equations to insure a per-
fect matching between the TC’s components applying
equations (26)–(29). Meanwhile, for more details about
the CC’s impeller design, see Chettibi et al.
46
and
Schiff.
50
Assumptions and design requirements used to
start the TC design process are summarized in Table 3.
Finally, the works of Bourabia e al.
18
and Dixon and
Hall
51
are judged sufficient to be used as TC losses
models in the present paper.
Engine modeling
In this subsection, two models are developed to esti-
mate the thermodynamic properties of air and gas at
the ICE intake and exhaust manifolds, respectively.
Pre-engine model. The main difficulties encountered dur-
ing matching TC to ICE consists of determining
required, air flow rate _
mCreq and charge pressure Preq
entering the engine’s intake manifold to reach the tar-
geted performance. Based on previous work presented
by Shamderakhshan and Kharazmi,
52
an analytical
model is implemented. The main steps follows the flow-
chart of Figure 4(a), hereafter, the detail of equations:
Table 1. Governing equations for turbocharger’s design equations.
TC component Eq. name Equation Eq. no.
RIT’s rotor Rotational speed w=nsDH0:75=ffiffiffiffi
_
v4
p(12)
Volume flow rate _
v4=_
mT
r04
(13)
Total to static velocity vs=0:73n0:2
s(14)
Total to static efficiency hTts =0:87 1:07 ns0:55ðÞ
20:5ns0:55ðÞ
2(15)
Rotor tip velocity U4=vsCis (16)
Isentropic speed
Cis =ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi
2CptT04 11
PCts

ggas1
ggas
!
v
u
u
t
(17)
Tangential tip velocity Cu4=U4hTts
2v2
s
(18)
Inlet absolute flow angle a4=10:8+14:2ns(19)
Rotor axial length Zr=1:5(rs5rh5) (20)
Number of blades NbT =12+0:03(33 a4) (21)
Slip factor s=1ffiffiffiffiffiffiffiffiffiffi
sin(b4)
p
N0:7
r
(22)
Inlet blade width b4=_
mT
2pr4r4Cm4(23)
Outlet hub radius rh5=0:185r4(24)
Meridional outlet velocity Cm5 =_
mT
2pr5r5b5(25)
CC & RIT
matching
equations
Mass flow balance _
mT=_
mC+_
mf(26)
Momentum balance r4
r2=1
U4=Cis (27)
Speed balance NT=NC(28)
Energy balance hMech =
_
WC
_
WT
=
_
mTCpgT4T5
ðÞ
_
mCCpaT2T1
ðÞ
=
_
mTCpgT41T5
T4

_
mCCpaT21T1
T2

=
_
mTCpgT41PT
1gg
gg
ts

_
mCCpaT21PC
1ga
ga
tt

(29)
6International J of Engine Research 00(0)

Determination of the required air mass flow for the
target power:
_
mCreq =PAF BSFC
3600 gr=sðÞ ð30Þ
Where, Pis the target power in (Watt), AF is air to fuel
ratio.
Calculation of the intake pressure to meet the required
power basing on the volumetric efficiency definition:
Preq =2_
mair RT3
hvNVdð31Þ
where, Preq is the required absolute pressure in the inlet
manifold in (Pa), T3is the intake manifold temperature
in (K), hvis the volumetric efficiency, Vdis the engine
displacement volume in (m
3
) and N is the engine speed
in (rps). R is the gas constant.
Calculation of the compressor outlet pressure:
P2c =Preq +DPloss ð32Þ
where, DPloss is the pressure loss between the compres-
sor outlet and engine intake manifold, which is the sum
of pressure loss through both intercooler and air ducts.
From the open data of TC constructors, the pressure
loss (between 6 and 27 kPa) can be estimated regarding
the existence or not of intercooler.
Estimation of the compressor inlet pressure;
P1c=Pamb DPfilter ð33Þ
where, Pamb and DPfilter are the ambient pressure and
pressure losses in the air filter in (Pa), respectively.
Consequently, the required pressure ratio is calcu-
lated as follows:
PCreq =P2c
P1cð34Þ
Estimation of compressor outlet temperature using
the total-to-total efficiency definition;
T2=T11+ 1
hCtt
P
g1
g
Creq 1

ð35Þ
Where, T1and hCtt are the ambient temperature and
compressor isentropic efficiency, respectively.
Then the intake manifold temperature is calculated
as follows;
T3=T21eIC
ðÞ+TIC eIC ð36Þ
where, eIC and TIC are the intercooler effectiveness and
the coolant temperature, respectively. The coolant tem-
perature is assumed equal to 35°C in the present work.
Post-engine model. Thermodynamic, heat transfer and
combustion equations have been used to model the ICE
exhaust gas properties based on works of Heywood
53
and Ferguson and Kirkpatrick.
54
This 0D single-zone
model is widely used in open literatures,
53–55
as it gives
a good description of in-cylinder heterogeneities
involved in the local combustion and pollutant produc-
tion terms, while it requires reasonable computer
resources. In addition, it allows to predict both beha-
vior and performance of engines of various displace-
ments at several operating conditions, without resorting
to expensive experimental test benches. The Post-engine
model outputs are used as input data to start the tur-
bine’s rotor design. However, as pollutants prediction is
not essential for the design and matching process as
depicted in the flowchart of Figure 4(b), theses terms
are not considered in the present code.
The thermodynamic single-zone model does not dif-
ferentiate between the burned gases and fresh charge
air so the in-cylinder mixture (air + fuel) is considered
uniform and homogeneous. The combustion process
can be simplified with a heating flux supplied from the
external environment. The differential form of the
energy conservation is then obtained:
dT
dt =
1
m
dQw
dt +dmfinj
dt LHV + Xdmin
dt hin Xdmout
dt hout udm
dt

P
m
dV
dt ∂u
∂f
∂f
∂t
 
∂u
∂t
ð37Þ
Table 2. Inputs data of the turbocharger design process.
Parameters Value Optimal variation range Reference
RIT specific speed 0ns00.59 [0.45–0.75] Aungier
48
CC radii ratio 0rh1=r200.15 [0.15–0.25] Aungier
49
CC maximum blade leading 0Blmax00.2 [0.1–0.9] Dixon and Hall
51
TC mechanical efficiency 0hMech00.95 ø0.9 Watson and Janota
5
CC isentropic efficiency 0hCtt 00.8 ø0.6 Dixon and Hall
51
CC inlet relative Mach number 0Mrel100.6 [0.1–0.9] Dixon and Hall
51
CC inlet absolute flow angle 0a100[0°–30°] Dixon and Hall
51
CC inlet incidence angle 0i100’0 Aungier
49
CC inlet relative flow angle 0b1060°[56°–60°] Dixon and Hall
51
RIT outlet absolute flow angle 0a500°[0°–45°] Bourabia et al.
46
Vane diffuser area ratio 0AR01.9 [1.4–2.4] Aungier
49
El Hameur et al. 7

where, h, Q, W, u and m are the specific enthalpy, heat
exchanged, work produced, specific internal energy and
total mass in the combustion chamber, respectively. In
addition, f,Q
w,m
finj , LHV are the air to fuel ratio
(AF), heat transferred through the cylinder wall, mass
of fuel injected and the lower heating value of the
injected fuel, respectively.
Applying the continuity equation, considering the
state equation and the differentiation to its logarithmic,
the total mass variation of the in-cylinder mixture is
found:
dm
dt =Xdm
dt

in Xdm
dt

out
+dmfinj
dt ð38Þ
Results and discussions
Two different displacement TCed Diesel engines;
namely, engine I, a three-cylinders 1.5 L and engine II,
a six-cylinders 5.9 L, are attempted to be boosted with
new TCs. In this aspect, the Ricardo Wave
Ò
simulation
software is used to validate the developed methodology
by comparing the simulation results of the proposed
TCs and the original ones after executing the developed
methodology. The technical characteristics of the stud-
ied ICEs and the targeted performance (power and
BSFC) are listed in Table 3 and their simulation models
are shown in Figure 5.
As presented in flowchart of Figure 3, the original
TCs mounted on the two studied engines are removed
to instigate the numerical simulation on NA ICEs only.
Indeed, the TC preliminary design code is extremely
sensitive to engine’s exhaust gases data (P4,T
4).
Therefore, the Poste-engine model is used to estimate
the exhaust parameters of the NA engine. The finding
are compared with the simulation results, as shown in
Figure 6(a) and (c) for engine I and Figure 6(b) and (d)
for engine II. The results show that the measured maxi-
mum deviations from the Post-engine model compared
to WaveÒsimulations, are for engine I, in term of P4
and T4around 2.22% and 2.66%, respectively, and
around 19.49% and 5.19% for engine II, respectively.
The deviation is better marked for the engine II, this
can be interpreted, first, by the use of the waste- gate
options in the case of Ricardo–wave simulations which
allows the six-cylinders 5.9 L to work in the optimum
operating conditions of the turbine and the compres-
sor, far from the instabilities zones, whereas, in the
developed simulations, the totality of the gases are used
in the thermodynamic cycles which may cause inaccu-
rate solutions. Second, the deviation is related to the
fact that the thermodynamic 0D-single zone model, is
slightly inaccurate to predict the thermodynamic para-
meters when the displacement of the engine becomes
large especially at high engine speeds due to inaccurate
quantification of heat transferred to ambient, released
gas flow, friction losses and fresh air and fuel
Figure 4. Flowcharts of: (a) pre-engine model and (b) post-engine model.
8International J of Engine Research 00(0)

interaction, which become significant. Concretely, as
seen in Figure 6(b), (d), (f), and (h), the trend of the
curves obtained by the developed code follows perfectly
the Ricardo–wave simulation’s results which consoli-
dates the given explanations.
Anyway, applying the selection procedure of case 1,
three TCs were proposed for each studied engine,
whereas, for the case 2 and after applying the TC pre-
liminary design steps, two TCs were found for engine I
(namely, Master Power R343 and R444) and only one
TC over 500 TCs was found for engine II (namely,
Master Power R615), as shown in Table 4. As a result,
in the present paper only TCs found after executing the
second case will be matched and tested on NA engines.
Figure 6(e) and (f) depict the comparison between
the simulation and the Pre-engine model results for the
TCed engines, in terms of required mass flow rate
_
mairreq and required pressure ratio PCreq, respectively, to
achieve the targeted performance of engine I as stated
in Table 3. The same comparison is made for engine II
in Figure 6(g) and (h). The deviations of PCreq and
_
mairreq for engine I are around 3.23% and 9.91%,
respectively, while for engine II deviations are found
around 4.63% and 8.61% of the required data,
respectively.
The deviations found when comparing the results
of simulations and implemented Pre-engine model are
mainly due to unquantified pressure loss (primary and
secondary losses) through the intake pipes and due to
the assumption made in the subsection 2.4.1 relative
to pressure loss through the intercooler and air filter.
Table 3. Engines specifications and targeted performances.
Parameters Engine I Engine II
Engine type TCed, four stroke, Diesel engine TCed, four stroke, Diesel engine
Number of cylinders 03 06
Displacement [L] 1.5 5.9
Compression ratio 16.5 17
Bore [mm] 3Stroke [mm] 86 386 102 3120
Brake rated Power [kW] With TC : 129.92 @ 4000 rpm With TC : 200.86 @ 2500 rpm
Without TC : 49.57 @ 4500 rpm Without TC : 80.84 @ 2500 rpm
Brake rated Torque [N.m] With TC : 310.16 @ 4000 rpm With TC : 932.74 @ 1250 rpm
Without TC : 105.19 @ 4500 rpm Without TC : 399.13 @ 1250 rpm
BSFC [g/kW/h] With TC : 218.7 @ 4500 rpm With TC : 218 @ 2500 rpm
Without TC : 223.2 @ 4500 rpm Without TC : 541.6 @ 2500 rpm
Connecting rod length [mm] 128.6 200
Valve diameter [mm] 29.25 29
Valve width w [mm] 1.46 1.45
Valve geometric angle b[°]45 45
Targeted brake power [kW] Around 125 Around 200
Targeted BSFC [g/kW/h] Around 219 @ 4500 rpm Around 218 @ 2500 rpm
Figure 5. Engines simulation models obtained from Ricardo Wave
Ò
software: (a) engine I and (b) engine II.
El Hameur et al. 9

Figure 6. Comparison between the simulation results and the developed codes (Pre and Post-engine models): (a and b) engine
exhaust pressure, (c and d) engine exhaust temperature, (e and f) pressure ratio, (g and h) mass flow rate.
10 International J of Engine Research 00(0)

From the deviation findings, the implemented Pre
and Post-engine models are judged as accurate and
truthful models that can be used for further calcula-
tions during turbocharging process. The CC and RIT
designs are generated using design specifications listed
in Table 5 at full load and at engine’s speed of 4000
and 2500 rpm for engine I and engine II, respectively.
Table 6 summarizes the TCs geometric and aerody-
namic parameters, and performance data obtained after
the convergence of the TC preliminary design proce-
dure to reach the targeted performance of the two stud-
ied engines.
The estimated geometrical and aerodynamic para-
meters of adequate TCs for the two engines are then
used to establish the 3D geometry of TC’s rotors using
ANSYS BladeGen
Ò
software, where the flow path and
the blade profiles can be modified and modeled if it is
necessary. The CC and RIT 3D designs for each studied
engine are depicted in Figure 7(a) and (b). However,
neither the flow field analysis nor the matching
procedure of the generated rotors will be discussed in
the present paper.
Figure 8(a) shows results of engine’s brake power
obtained after comparing the simulation analysis of the
original engine I (as TCed and NA engine) and the two
new engines equipped with the selected TCs from the
DB. The maximum deviations observed between the
original TC and the TCs R444 and R343 equipping
engine I, are around 48.62% and 13.93%, respectively.
Furthermore, from Figure 8(c), engine I matched
with TC R444 expresses lower BSFC and develops
more output power than the original one, especially at
high speeds. Whereas, the TC R343 equipping engine I
developed higher BSFC compared to the original one
at both cases (as TCed and as naturally aspirated).
From Figure 8(b) the maximum deviation between the
original TC and the selected one equipping engine II in
terms of brake power is around 5.11% (2.64 kW higher
than the original one), which is judged acceptable after
satisfying the targeted performance specified in Table
Table 4. Number of TCs proposed by the selection procedure.
Engines Engine I Engine II
Case 1 Case 2 Case 1 Case 2
Number of TCs selected 3231
Table 5. Turbocharger design requirements.
Centrifugal compressor Radial inflow turbine
Parameter Engine I Engine II Parameter Engine I Engine II
pressure ratio 0PCtt 02.82 2.589 expansion ratio 0PTts 01.137 2.67
mass flow rate 0_
mair0[kg/s] 0.050 0.299 mass flow rate 0_
mgas0[kg/s] 0.05185 0.311
inlet pressure 0P010[bar] 1.013 inlet pressure 0P04 0[bar] 5.069 2.638
inlet temperature 0T010[K] 298.15 inlet temperature 0T04 0[K] 1094.17 908.89
CC tolerance 0TolC0(%) 10
specifique speed 0ns00.59
Table 6. Turbocharger geometric parameters and performances data of the designed CC and RIT of the studied engines.
Centrifugal compressor Radial inflow turbine
Parameter Engine I Engine II Parameter Engine I Engine II
tip radius 0r20[mm] 39.495 65.5 tip radius 0r40[mm] 38.814 58.189
shroud radius 0r1s0[mm] 32.4 39.2 shroud radius 0r5s 0[mm] 21.348 40.150
hub radius 0r1h0[mm] 10.4 15.3 hub radius 0r5h 0[mm] 4.076 4.015
number of blade 0NCb012 24 number of blade 0NTb010 10
speed 0NC0[rpm] 131,760 65,114 speed 0NT0[rpm] 131,760 65,114
isentropic efficiency 0hCtt 00.8348 0.7946 isentropic efficiency 0hTts 00.842 0.848
Inlet absolute flow angle 0a100 0 Inlet absolute flow angle 0a4072.092 72.4785
Inlet relative flow angle 0b1060 60 Inlet relative flow angle 0b40245 271.985
Exit absolute flow angle 0a20690,219 72.4447 Exit absolute flow angle 0a5000
Exit relative flow angle 0b20671,577 71.3637 Exit relative flow angle 0b5059.7435 250.4538
El Hameur et al. 11

3. Furthermore, as seen in Figure 8(d), the BSFC of
engine II matched with TC R615 is under the original
one especially at mid-high engine speed.
From Figure 8(a) it appears that TC R444 equipping
engine I provides a good performance with a maximum
brake power over the original one around 11.26kW at
4000 rpm. Moreover, the TC R343 offers relatively
good results overall the engine’s speed range, but it
develops a rated power at 4000 rpm of 126.41 kW,
which is under the rated power of the original one by
3.51 kW. The TC R615 equipping engine II presents a
good agreement with the performance of the original
TC. The new TCed engine offers a rated power of
299.3 kW, which is slightly under the original one by
2.64 kW. In addition, measured CC and RIT isentropic
efficiencies equal 75.24% and 65.19%, respectively.
Consequently, the new proposed TCs responded posi-
tively to the targeted performance of engine I and II.
Comparing only the output power and BSFC deliv-
ered by the engines after matching them with proposed
TCs, seems not sufficient. For this reason, a superim-
position of engine airflow on a compressor and turbine
maps at different engine speeds is then essential to
assess the TC-ICE operating stability.
After launching engines’ simulations using CC per-
formance maps extracted from DB and RIT perfor-
mance maps generated by the in-house code, a
comparison is then established between the original
TCs (see Figures 9(a) and (b) for engine I and Figures
10(a) and (b) for engine II) and the proposed TCs as it
is shown in Figure 9(c) to (f) for engine I and Figures
10(c) and (d) for engine II.
From Figures 9(a), (b), 10(a), and (b) the surge mar-
gin, the CC and RIT isentropic efficiencies at engine
rated power and 4000 rpm are 49.35%, 63.35%, and
61.80% for the engine I and 48.99%, 76.68%, and
70.47% for engine II, respectively.
From the Figure 9(c) and (d), the operation area of
the first engine equipped with TC R343 is far from the
surge line at all its speed range with a surge margin of
44.17% at rated power. Furthermore, the engine’s oper-
ation area at high speeds approaches isentropic effi-
ciency of CC and RIT around 59.48% and 39.18%,
respectively. Engine I matched with TC R444 has an
operating area that approaches the surge line at low
and medium speed with a surge margin that equals
21.46% at 4000 rpm. It has also a CC and RIT isentro-
pic efficiencies of 65.27% and 50.87%, respectively (see
Figure 9(e) and (f)). For the TC R615 matched with
engine II, a surge margin of 26.19% is estimated (see
Figure 10(c)) at 2500 rpm.
Finally, to ensure that the developed methodology
proposes a satisfactory TCs in terms of engine and tur-
bine thermal stress. A more detailed investigation is
performed this time considering the in-cylinder pressure
Pcyl and turbine inlet temperature (TIT). The Pcyl and
TIT calculated by the Post-engine model after matching
TCs R444 and R343 with engine I are compared to Pcyl
and TIT produced by the original engine at full load
and rated power versus engine’s speed range. The find-
ings depicted in Figure 8(e) shows that TC R444 devel-
ops 1078.6 K of TIT, which is the lowest one compared
to the TITs generated by the original TCed engine
and the engine matched with TC R343. However, the
Figure 7. Proposed CC and RIT 3D designs for each engine: (a) Engine I and (b) Engine II.
12 International J of Engine Research 00(0)

Figure 8. Comparison of the engines’ performance and thermodynamic limitations between the original and new matched TCs:
(a) brake power, (c) BSFC, (e) TIT, and (g) Pcyl of engine I; (b) brake power, (d) BSFC, (f) TIT, and (h) Pcyl of engine II.
El Hameur et al. 13

TIT measured with R615 is slightly higher than the
original one especially at low and mid-engine speeds
(see Figure 8(f)).
The TIT deviations measured in both engines while
using R343, R444 and R615 are not so high and present
no risks on the RIT, knowing that nowadays materials
used in RIT’s rotors can support a TIT around 1300
K.
51
The same comparison is performed for engine I
and II in terms of Pcyl as shown in Figure 8(g) and (h).
Findings show that all proposed TCs generate a Pcyl
under the original ones at all engine speed ranges, which
ensure a good engine’s strength against thermal stress.
Figure 9. Turbocharger joint operating maps at full load of the engine I: (a and b) CC and RIT original maps, respectively, (c and d)
CC and RIT maps of the TC Master Power R343, respectively, (e and f) CC and RIT maps of the TC Master Power R444,
respectively.
14 International J of Engine Research 00(0)

Finally, considering the results observed in Figures 8
and 9, the two proposed TCs for engine I present a good
performance data and security range from the choke
and surge lines. However, if only one TC would be
selected between these two TCs, the TC Master Power
R444 is appropriate if the user desires higher CC and
RIT efficiencies and lower BSFC. However, if the secu-
rity is desired during all engine operations far from the
surge line, in this case, the TC Master Power R343 is
preferred due to its higher surge margin and lower Pcyl .
The TC R615 offers a small surge margin; neverthe-
less, its CC and RIT efficiencies are located at the map’s
highest efficiency islands and offers acceptable TIT and
BSFC at mid-high engine speeds.
Conclusion
In the present paper, a turbocharging methodology
based on TC preliminary design and selection is devel-
oped to reach targeted performance. The methodology
is tested and validated with simulation test results on
two different displacement engines. The procedure
consists of four main parts. The developed selection
procedure proposes two cases depending on compres-
sor’s map availability and on the accuracy required by
the user. Consequently, a TC preliminary design is gen-
erated when the second accurate proposed case is fol-
lowed. Based on the latter geometrical parameters, a
TC’s 3D geometries were proposed for each studied
engine. Integrating these data into the constructed TCs
database and applying the developed RIT’s perfor-
mance map code. Only two kinds of TCs over 500 were
selected to turbocharge the first engine, while only one
TC was selected for the second ICE. The proposed TCs
were matched on each respective engine, and analyzed
under full-load conditions and over engine’s speed
range. A good agreement was found between the results
of the original TCed engines and the matched ones in
terms of performance and consumption. In addition,
the superimposing of the engines’ operating areas on
CC and RIT maps was carried out to measure their
isentropic efficiencies and the TC-ICE matching stabi-
lity. Furthermore, to ensure that the developed metho-
dology proposes a satisfactory TCs, an investigation
Figure 10. Turbocharger joint operating maps at full load of the engine II: (a and b) RIT and CC original maps, respectively; (c and
d) RIT and CC maps of the TC Master Power R615, respectively.
El Hameur et al. 15

was established to measure the Pcyl and TIT to ensure a
proper TC-ICE thermal strength.
Finally, the main advantage of the developed
approach is, its ability to be applied at both earlier and
last stages of the engine turbocharging process.
Consequently, it can be used to boost a NA ICEs or to
reach a better performance and/or consumption of
TCed engines.
Declaration of conflicting interests
The author(s) declared no potential conflicts of interest with
respect to the research, authorship, and/or publication of this
article.
Funding
The author(s) received no financial support for the research,
authorship, and/or publication of this article.
ORCID iDs
Mohamed Amine El Hameur https://orcid.org/0000-0001-
9119-7205
Mahfoudh Cerdoun https://orcid.org/0000-0002-1079-
0996
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El Hameur et al. 17

Appendix A
Notation
Symbols
ACross-section area (m
2
)
AF Air to fuel ratio
bBlade width (m)
CAbsolute velocity (m/s)
CpSpecific heat at constant pressure (kj/kg
K)
CvSpecific heat at constant volume (kj/kg K)
dhHydraulic diameter (m)
DDiameter (mm)
hSpecific enthalpy (kj/kg)
LB Blade mean camberline length
MMach number
_
mMass flow rate (kg/s)
NRotational speed (rpm), number of blade
nsSpecific speed
oThroat (m)
PPressure (Pa)
PPower (kw)
DPPressure loss (kPa)
_
QHeat flow rate (W)
rRadius (m)
RRadius (m), gas constant (J/kg/K)
ReReynolds number
sSpecific entropy (kj/kg K)
TTemperature (K)
UPeripheral velocity (m/s)
VVolume (m
3
), volumetric
_
vVolume flow rate (m
3
/s)
wAngular velocity (rad/s)
WRelative velocity (m/s)
ZNumber of blades
zVolute flow passage depth (m)
ZrRotor axial length (m)
Abbreviation and acronyms
AR Area to radius ratio (m)
BMEP Brake mean effective pressure (bar)
BSFC Brake specific fuel consumption (g/kw/h)
CC Centrifugal compressor
CCA Combustion chamber
CFD Computational fluid dynamic
CI Compressor ignition
CCM Combustion chamber
DB Database
DC Design conditions
DI Direct injection
ICE Internal combustion engine
NA Naturally aspirated
RIT Radial inflow turbine
SI Spark ignition
TC Turbocharger
TDM thermodynamic models
TIT Turbine inlet temperature
IVO, C Intake valve opening, closing
EVO, C Intake valve opening, closing
Greek letters
aAbsolute flow angle (°)
bRelative flow angle (°)
gAdiabatic coefficient
dHeat exchanger effectiveness
eEffectiveness, correction factor
hEfficiency
uAzemutal angle (°), tangential direction
lFriction factor
PPressure ratio
mDynamic viscosity (kg/s)
rDensity (kg/m
3
)
sImpeller slip factor
tDiesel ignition delay (s)
Figure A1. Velocity triangles applied to establish turbocharger’s components design.
18 International J of Engine Research 00(0)

fFlow coefficient, air to fuel ratio
cStage loading coefficient
vAngular speed (rad/s)
Subscripts
0 Total condition, impeller inlet
1, 2, 3,
4.etc.
States points
xAxial direction
OVR Overall
th Thermal
gGas
a Air
Dev Deviation
b Blade
ch Choking
T Turbine
C Compressor
is Isentropic
h Hub
tt Total total
ts Total static
m Meridional direction
req Required
r Radial direction
s Shroud, blade interspace (m)
t Thickness (mm), tangential direction, tip
x Axial direction
I Inducer
E Exducer
D Design
Mech Mechanical
IC Intercooler
El Hameur et al. 19