proc imeche part a: turbocharging strategy among variable

Original Article

Turbocharging strategy among variablegeometry turbine, two-stage turbine,and asymmetric two-scroll turbine forenergy and emission in diesel engines

Dengting Zhu, Zhenzhong Sun and Xinqian Zheng

Abstract

Energy saving and emission reduction are very urgent for internal combustion engines. Turbocharging and exhaust gas

recirculation technologies are very significant for emissions and fuel economy of internal combustion engines. Various

after-treatment technologies in internal combustion engines have different requirements for exhaust gas recirculation

rates. However, it is not clear how to choose turbocharging technologies for different exhaust gas recirculation require-

ments. This work has indicated the direction to the turbocharging strategy among the variable geometry, two-stage, and

asymmetric twin-scroll turbocharging for different exhaust gas recirculation rates. In the paper, a test bench engine

experiment was presented to validate the numerical models of the three diesel engines employed with the asymmetric

twin-scroll turbine, two-stage turbine, and variable geometry turbine. On the basis of the numerical models, the

turbocharging routes among the three turbocharging approaches under different requirements for EGR rates are

studied, and the other significant performances of the three turbines were also discussed. The results show that

there is an inflection point in the relative advantages of asymmetric, variable geometry, and two-stage turbocharged

engines. At the full engine load, when the EGR rate is lower than 29%, the two-stage turbocharging technology has the

best performances. However, when the exhaust gas recirculation rate is higher than 29%, the asymmetric twin-scroll

turbocharging is the best choice and more appropriate for driving high exhaust gas recirculation rates. The work may

offer guidelines to choose the most suitable turbocharging technology for engine engineers and manufacturers to achieve

further improvements in engine energy and emissions.

Keywords

Turbocharging strategy, variable geometry turbine, two-stage turbine, asymmetric twin-scroll turbine, exhaust gas recir-

culation, diesel engine

Date received: 14 June 2019; accepted: 6 November 2019

Introduction

At present, advanced turbocharging technologies playa central role in meeting the increasingly stringentemission and fuel consumption legislation for internalcombustion engines (ICEs).1 In ICEs, more than30–40% of fuel energy wastes from the exhaust andjust 12–25% of the fuel energy convert to usefulwork.2 Conklin and Szybist3 found that the thermalenergy lost through exhaust gas was approximately27.7%, and Wang et al.4 determined that the totalfuel saving of the engine could be up to 34% fromthe exhaust under certain operating conditions.Turbocharger technology makes it possible forengine downsizing by decreasing pumping loss and,meanwhile, makes full use of exhaust energy.5

Hatami et al.6,7 investigated the exergy recovery

from a DI diesel engine by experimental research atvarious engine speeds, and the results showed thatby using recovered exergy, brake-specific fuel con-sumption (BSFC) decreased approximately 10%.Meanwhile, they optimized the finned-tube heatexchangers for diesel exhaust waste heat recoveryusing computational fluid dynamics (CFD) and cen-tral composite design techniques.8

Turbomachinery Laboratory, State Key Laboratory of Automotive

Safety and Energy, Tsinghua University, Beijing, China

Corresponding author:

Xinqian Zheng, State Key Laboratory of Automobile Safety and Energy,

Tsinghua University, Beijing 100084, China.

Email: [email protected]

Proc IMechE Part A:

J Power and Energy

0(0) 1–15

! IMechE 2019

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Exhaust gas recirculation (EGR) is a crucial tech-nology to meet more stringent nitrogen oxide (NOx)emission regulations for ICEs. By pumping someexhaust gases back into cylinders for a smaller pro-portion of O2, NOx emissions will be decreased. Guoet al.9 concluded that the exhaust temperature is acritical factor for NOx from diesel engines. EGR dir-ects the exhaust gas back to engine cylinders for sec-ondary combustion, and therefore, it decreases thecombustion temperature and oxygen concentrationfor lower NOx. The EGR rate, i.e. the general controlparameter for adjusting the open degree of the EGRvalve, is the percentage of the recirculation gas in thetotal inlet.10 Almeida et al.11 investigated the EGRinfluence laws on NOx and soot emissions, andincreasing the EGR rate will reduce NOx emissionsbut increase PM emissions. Meanwhile, the vastmajority of soot will be absorbed by the diesel par-ticulate filter in the after-treatment.

Euro VI NOx limit of 0.4 g/kWh has been set forworld harmonized steady-state cycle and world har-monized transient cycle, which is quite challenging tomeet. Different after-treatment technology routes inICEs have different requirements for EGR rates.12

Cloudt et al.13 summarized three technical routesfor reducing NOx emissions: selective catalytic reduc-tion (SCR)-only, SCRþEGR, and EGR-only. For a370 kW Euro VI engine, the three ways aboverequire EGR rates of 0, 15–30%, and 40–60%,respectively. At present, some advanced turbochar-ging technologies, such as asymmetric twin-scrollturbine (ATST), two-stage turbine (2ST), and vari-able geometry turbine (VGT), are increasingly beingused by automotive manufacturers. The three tur-bines can provide a higher exhaust back-pressurefor driving EGR.

VGTs are increasingly mounted on engines com-bined with EGR. VGTs can change the turbinethroat area by adjusting the position of vanes.14

Improved engine performance considering bothpower and emission can be achieved.15 Notably,they will dramatically increase the low-speed torqueand transient response.16,17 Ahmed et al.18 provided areliable method to simulate the influences of exhaustgas flow through the variable turbine geometry sec-tion on the flow control mechanism. Meanwhile,Galindo et al.19 developed and validated a one-dimen-sional radial turbine model for transient pulsatingflow applications, and Hatami et al.20 achieved anoptimized design for the vanes geometry with76.31% efficiency (averagely in all pressures). Allsimulators above showed proper compliance withexperimental and CFD results. Using VGT to driveEGR has clear advantages. In EGR–VGT dieselengines, Glenn et al.21 addressed some issues asso-ciated with real-time control. Zamboni andCapobianco22 experimentally studied the impacts ofcontrol strategies combining VGT with low-pressure(LP) and high-pressure (HP) EGR and achieved

reductions of 2.1and 50% of fuel consumptionand NOx. Moreover, Llamas and Eriksson23 andGelso and Dahl24 optimized the EGR–VGT controlof heavy-duty diesel engines during transients.Between the EGR and VGT, an essential aspect ofthe gas exchange regulation issue is a complex inter-action. They aim to optimize fuel consumption byminimizing pumping losses. Compared with a fixedgeometry turbine (FGT), the VGT has an improve-ment of about 10% on the full-load torque,25 andZamboni et al.26 achieved a fuel consumption thatwas down by 1.5–9.5% at low speed/load combiningVGT and EGR.

The 2ST is another advanced turbocharging tech-nology to decrease fuel consumption and pollutantemissions and simultaneously increasing engineoutput torque and power.27 The 2ST system has twoturbines with different throat areas, and it can achieveenough boost-pressure and decrease pumping lossthrough exhaust flow distribution.28 The HP stageand the LP stage are connected sequentially and regu-lated via a by-pass valve. Galindo et al.29 investigatedthe performance of the two-stage turbochargingapproach, and the results presented that the overallperformance of the 2ST system for the entire opera-tive range and characterized the optimum control ofthe components for various operating conditions.Avola et al.30 investigated the accuracy of perform-ance measurement in two-stage turbocharging sys-tems in the context of aerothermal interstagephenomena and concluded that the turbine net effi-ciency was reduced by about 10% at elevated cor-rected mass flow operations. The 2ST is alsoadopted at high altitudes or in flight, as it can offera high compression pressure ratio. Yang et al.31 stu-died the performances of a regulated 2ST system athigh altitudes and its potential for exhaust energyrecovery of engines. Zhao et al.32 focused on the char-acteristics of the 2ST under steady and pulsed admis-sions. The averaged cycle expansion ratio of HPTreduced while that of LPT increased as the pulse amp-litude increased. Moreover, VGT and 2ST can be usedtogether. Zheng et al.33 conducted a theoretical andexperimental study using an HP VGT and a LP FGTfor a higher boost-pressure and EGR. The optimalfuel efficiency could be achieved by compromisingthe engine pumping work and boost-pressure.

The ATST is a new technology compared to thetwo technologies above. Since the concept was firstproposed in 1954,34 twin-entry turbochargers havebeen widely applicable in turbocharged engines forimproving the potential of pulsating flow energy.This design surpasses the design of traditional turbo-chargers, which usually only considers the steady-state flow performance and can improve the designof turbochargers to enhance the performance of pul-sating flow turbines.35 Romagnoli et al.36 and Rajooet al.,37 respectively, analyzed the steady and unsteadyperformances of the variable geometry twin-entry

2 Proc IMechE Part A: J Power and Energy 0(0)

turbine. The results showed that twin-entry turbineshad a better performance at improving the cycle-aver-aged efficiency compared with single-scroll turbines.

To meet more stringent demand of emissions andfuel consumption, ATSTs have been used to realizeHP EGR using only one scroll while designing theother scroll for optimal scavenging. Muller et al.38

first described the principle mechanism of the ATSTand presented 3D CFD flow simulations to demon-strate the turbine characteristic and the underlyingcorrelation between the two scrolls. Hand et al.39 pre-sented a model developed component-wise for theengine, and special attention was given to the flowsof the ATST. The developed model showed goodagreement with steady-state measurements. Brinkertet al.40 experimentally studied the internal flow simi-larities of the ATST under different unequal turbineadmission conditions. Since the beginning of the 21stcentury, Daimler has developed a series of differentasymmetric twin-scroll turbocharged diesel engines.Kruger et al.41 showed the 10.7 L heavy-duty dieselengine named OM 470 employed with an asymmetrictwin-scroll turbocharger, which was destined forworldwide applications. Later, the 12.8 L OM 471,14.8 L OM 472, and 15.6 L OM 473 graduallybecame the products for decreasing NOx emis-sions.42,43 Zhu and Zheng44 experimentally studiedthe differences in engine performance between theconventional symmetric twin-scroll turbines andATSTs and concluded that the ATST could achievea better trade-off between engine fuel consumptionand emissions. After that, they proposed the new con-cepts of the ATST with two wastegates (WG)45 andwith two EGR circuits.46 Compared with the trad-itional ATST, they could increase the fuel economyby 2.91 and 1.98%, respectively.

At present, VGT, 2ST, and ATST are threeadvanced turbocharging technologies combined withEGR, and they are increasingly used in automobiles.However, the EGR requirements are different underdifferent after-treatment technology routes in ICEs.There is no clear guideline of the turbocharging

strategy for different EGR. This paper will addressthis issue. In this study, a comparison of the VGT,2ST, and ATST on engine performance is first inves-tigated, and the effects of the engine speed and EGRrate are explored. This article consists of three parts:first, a test bench experiment of the diesel engineemployed with an ATST system and an EGRsystem is presented, and the simulation models ofthe three engines with the VGT, 2ST, and ATST areestablished and validated. Second, the turbochargingroutes among the three turbocharging approachesunder different requirements for EGR rates are stu-died. Last, the other significant performances (cost,transient response, reliability, and complexity) of thethree turbines are discussed and compared.

Experimental and simulative methods

This section performs an experiment on an asymmet-ric twin-scroll turbocharged diesel engine. The simu-lation model of the diesel engine with the ATST hasbeen established and validated according to the testdata. Meanwhile, keeping the engine system and theEGR system unchanged, the variable and two-stageturbocharged diesel engine models have been com-pleted, and all working lines of the engines are inhigh-efficiency areas of the compressors.

The experiment of the engine with an ATST

The studies above show that asymmetric twin-scrollturbochargers have considerable potential for energyand emission improvements. An ATST has two tur-bine scrolls with different throat areas (Figure 1(a)).Generally, ATSTs are adopted in multi-cylinderengines, and Figure 1(b) shows a schematic diagramof a six-cylinder diesel engine employed with anATST. During on-engine application, the turbinetwo scrolls are, respectively, linked with the two cylin-der groups. For a given exhaust flow (from the enginecylinders), the intake pressure p3ð Þ will be lower thanthe operating pressure of the small scroll p1ð Þ for

Figure 1. (a) Asymmetric twin-scroll turbine and (b) schematic diagram of a six-cylinder diesel engine with an ATSTand EGR system.

Zhu et al. 3

enabling EGR, even at low engine load conditions.47

Meanwhile, the large scroll pressure p2ð Þ is lower todecrease pumping work. Typically, there is a relation-ship as follows

p1 4 p2 4 p3 ð1Þ

An important parameter called turbine asymmetry(ASY) is defined as the ratio of the small volute throatarea TA1ð Þ and the large volute throat area TA2ð Þ. TheASY has a significant influence on the trade-offbetween engine fuel economy and NOx emissions

ASY ¼TA1

TA2ð2Þ

In the study, an experiment was carried out on a12.55 L, six-cylinder diesel engine equipped with anintercooled EGR system and an asymmetric twin-scroll turbocharger (ASY¼ 53%). Table 1 shows thedetailed test engine specifications, and for this engine,the rated power and the maximum torque are 351 kW(1900 r/min) and 2380Nm (1000–1400 r/min), respect-ively. The entire experimental engine system wasestablished on a dynamometer test bench accordingto the structural layout shown in Figure 1(b). Table 2presents the primary testing instruments in the experi-ment. The speed and the load of the engine areregulated by using a C 500 eddy current dynamometerwhose accuracy is �10 r/min and �1.25Nm. AnAVL735S fuel consumption measuring instrumentwhose precision is 0.12% of the recorded value isused to meter the engine fuel consumption. The dataacquisition system is a PUMA OPEN 1.2 all-in-onebench-top instrument, and data are processed by asoftware CONCERTO-P. The engine intake massflow is collected by a Sensyflow P/4000 air flowmeter(accuracy:�5mg). In the turbocharging system, the

ATST and the compressor have been tested on theturbocharged test bed, and the maps are shown inFigure 2. The maximum isentropic efficiencies of theATST and compressor are both over 76%. The test

Table 1. Test engine specifications.

Item Value and units

Engine type Inline six-cylinder DI diesel

Number of valves

per cylinder

Four (two inlet/two exhaust)

Bore 129 mm

Stroke 160 mm

Displacement 12.55 L

Compression ratio 18.2:1

Cooling system Water cooled

Air intake system Intercooled asymmetric twin-scroll

turbocharger (ASY¼ 53%)

EGR system Water intercooled EGR

Rated power 351 kW (1900 r/min)

Maximum torque 2380 Nm (1000–1400 r/min)

Table 2. Test instrument specifications.

Instrument Type (range and accuracy)

Eddy current dynamometer C 500 (0–4000 Nm

and 0–720 kW;

�10 r/min and�1.25 Nm)

Environment simulation

system

ACS2400 (298�1 K

and 100� 1 kPa)

Intake system Sensyflow P/4000 (�5 mg)

Coolant constant tempera-

ture control device

AVL553 (70–120�C,�1�C)

Gas emission analyzer MEXA-7100DEGR

Fuel constant temperature

control device

AVL753 (15–80�C,�1�)

Data acquisition system PUMA OPEN 1.2

Fuel consumption measuring

instrument

AVL735S

(0.1–110 kg/h, 0.12%)

(a)

(b)

Figure 2. Turbocharger maps: (a) asymmetric twin-scroll

turbine (ASY¼ 53%) and (b) compressor.


engine operated at full load and various speeds(800–1900 r/min) and its working conditions arelisted in Table 3. The engine performances measuredin the test will be used for the model calibration in thenext part.

Asymmetric twin-scroll turbocharged engine model

According to the experiment, the numerical model ofthe test asymmetric twin-scroll turbocharged dieselengine is developed and calibrated by the simulationtool GT-POWER v7.3.0 (Figure 3). GT-POWER, aregistered trademark of Gamma Technologies, iscommonly used for turbocharger matching andengine cycle simulation. This software covers sixaspects of the engine body, drive system, coolingsystem, fuel supply system, crankshaft mechanism,and valve mechanism.

In the model, all pipes have the same shape, size,and length as the test parts, and their friction coeffi-cient and heat transfer coefficient are determinedaccording to the materials. For the EGR system, themodel has an EGR circuit connected with the smallscroll of the ATST. The exhaust flow can be adjustedby the EGR valve, and meanwhile, it must be cooledin the EGR cooler before back to the cylinders.

In the engine model, the heat transfer can be cal-culated using the empirical WoschniGT. The instant-aneous average heat transfer coefficient for theworking gas and chamber wall �g

� �is very critical

and is calculated as equation (3). Then, the heatexchange for the gas and wall per unit engine crank

angle dQwd’

� �is naturally obtained using equation (4).

The two formulas above will be used to calculate heattransfer from the engine cylinders and crankcase,

Figure 3. GT-POWER model of the test asymmetric twin-scroll turbocharged diesel engine.

Table 3. Experimental engine operating conditions.

Speed

(r/min)

Atmospheric

pressure (kPa)

Atmospheric

temperature (K)

Fuel-injection

mass (kg/h)

Air-fuel

ratio

Torque

(Nm)

Power

(kW)

800 99.34 298.4 30.92 17.39 1818 152

1000 99.31 298.6 48.72 17.59 2385 250

1100 99.33 298.5 52.76 19.21 2385 275

1200 99.33 298.1 57.68 19.02 2386 300

1300 99.33 298.5 63.00 19.97 2385 325

1400 99.30 298.6 68.85 19.80 2385 350

1600 99.34 298.7 71.07 20.37 2096 351

1900 99.37 298.8 75.52 21.93 1761 350

Zhu et al. 5

which are very crucial for the engine output powerand fuel consumption

�g ¼ 820p0:8T�0:53D�0:2½C1Cm þ C2TaVs

paVað p� p0Þ�

0:8

�g � the instantaneous average heat

transfer coefficient for the

working gas and chamber wall

p,T� the working gaspressure and temperature

D� the cylinder diameter

Cm � the average piston speed

C1 � the gas velocity coefficient

pa,Ta,Va � the working gas pressure, temperature

and cylinder volume at the beginning of

compression, respectively

p0 � the cylinder pressure of the inverted engine

ð3Þ

dQw

d’¼X31

dQwi

d’¼

1

�

X31

�g � AiðT� TwiÞ

Qw � the heat exchange for the gas and wall

’� the crank angle

� � the angular velocity

A� the heat transfer area

Tw � the average wall temperature and

¼ 1, 2 and 3 respectively represent the cylinder

head, piston and cylinder liner:

ð4Þ

In particular, when the combustion object in thediesel engine model uses more than one temperaturezone, the convective heat transfer is calculated usingan effective gas temperature. The resulting heat trans-fer is applied to the burned and unburned zones bylinear weighting with the respective to mass in eachzone. For all multi-zone combustion models, thiseffective gas temperature Tgð Þ is as shown below

Tg ¼ Tbmb

mt

� �n

þTu 1�mb

mt

� �n� �

Tb � the burned zone temperature

Tu � the unburned zone temperature

mb � the burned mass

mt � the total mass

ð5Þ

Here, n is the weighting exponent that is calculatedfrom linear to quadratic as follows

n ¼ 1þmb

mt

� �2

ð6Þ

According to Table 3, the data of the environmen-tal conditions and fuel-injection mass are put into the

simulation models. The validations for engine per-formances including the engine torque, power,BSFC, and EGR rate are shown in Figure 4.The abscissas are dimensionless engine speeds (DES)that are dimensionless to the rated power point(1900 r/min). The ordinates are normalized to thetest data at 800 r/min. There is only a low deviationin the results of the experiment and simulation, andthe relative errors are within 2.0%. For this study, thedifferences are considered to be acceptable.

Variable geometry turbocharged engine model

Figure 5(a) presents the six-cylinder diesel engine withthe VGT component. In 2011, Romagnoli et al.36

presented the results from an experimental studyconducted on variable geometry single and twin-entry turbine designs for an automotive turbocharger,and they concluded that the variable geometry twin-entry turbine could make better use of the impulseenergy. Therefore, in this work, the VGT model hastwo identical entries, and both entries are used todrive EGR. Generally, the ASTS has a WG on theturbine large passage. This WG will keep closed atengine low speeds for enough boost-pressure andgradually open with the increasing speed for avoidingover-boosting. However, the VGT has no WGbecause it can control the intake pressure by adjustingthe turbine output power. Keeping the engine andEGR system unchanged, the simulation model ofthe variable turbocharged diesel engine is establishedin Figure 5(b). The throat area of the VGT is adjust-able, and the WG diameter is always set to zero.

Two-stage turbocharged engine model

The schematic diagram of a six-cylinder diesel enginewith a 2ST device is shown in Figure 6(a). As previ-ously described, the two-stage turbocharged enginesystem has two turbochargers with different sizes.The HP turbine has a small throat area to achieve ahigher boost-pressure at the low-range engine speed,and meanwhile, the LP turbine throat area is largerfor reducing the exhaust back-pressure at the high-range engine speed. To use exhaust pulse energy, theHP turbine has two identical scrolls, and both scrollsare linked with EGR. The two turbines both haveWGs (WG1 with the HP turbine; WG2 with the LPturbine). At low engine speeds, both the WGs arenormally closed, and all the engine exhaust gas goesthrough the HP turbine, resulting in a quick boostpressure rise on the air side. With increasing speeds,WG2 stays closed and WG1 gradually opens, and theexpansion ratio of the LP turbine rises constantly.To avoid excessive intake pressure, the two WGs areopen at the high-range engine speed, and the boost-pressure remains at the maximum value. Similarly, theturbocharger system is only changed according toFigure 3, and the two-stage turbocharged diesel


engine model is finished in Figure 6(b). The twoturbocharger maps are scaled for a proper range,and the validation of the models will be achieved inthe next part.

Model validation

The numerical models of the diesel engines with theATST, the VGT, and the 2ST have already been

Figure 5. Six-cylinder diesel engine with the VGT and EGR system: (a) schematic diagram and (b) GT-POWER model.

(a) (b)

(c) (d)

Figure 4. Performance calibration of the engine system: (a) torque, (b) power, (c) BSFC, and (d) EGR rate (all values are dimen-

sionless due to division by the experimental value at 42% DES).

Zhu et al. 7

established. In the variable geometry turbocharged dieselengine model and two-stage turbocharged diesel enginemodel, the EGR valve remains entirely open. Their tur-bine and compressor maps are obtained by multiplyingthe whole mass flow rate data of the asymmetric twin-scroll turbocharger maps with a factor (’/’0) and

keeping efficiency and pressure ratio or expansion ratiounchanged. When the three engine models have thesame EGR rate, the factor ’/’0 is 1.4 in the VGT mapat the rated power point (Figure 7(a)) and is 0.8 and 1.2in the HP turbine and LP turbine maps of the two-stageturbocharger, respectively (Figure 7(b) and (c)).

(a) (b)

(c)

Figure 7. Turbine maps from the experiment: (a) VGTat the rated power point, (b) HP turbine of the 2ST, and (c) LP turbine of the 2ST.

Figure 6. Six-cylinder diesel engine with the 2ST and EGR system: (a) schematic diagram and (b) GT-POWER model.


When the EGR valve is fully open, the working linein the compressor map of the engine with theATST (engine-ATST) is shown in Figure 8(a). Withthe engine speed increasing, the pressure ratio of theworking points grows and then remains constant.Because at the high range of the engine speed, theWG is open to avoid over-boosting. In the enginewith the VGT (engine-VGT), the position of the tur-bine vanes is adjusted to achieve the same EGRrate with the engine-ATST, and its turbine throatarea is changed. Therefore, the factor ’/’0 of the tur-bine is changing, but the factor ’/’0 of the compressorkeeps constant, which is the same as the compressor inthe engine-ATST. Figure 8(b) presents the workingline in the compressor map of the engine-VGT atthe whole range of the engine full load. The factor’/’0 of the compressor is 1.0, and it can be seenthat the working points are all in the high efficiencyareas of the compressors. For the engine with the 2ST(engine-2ST), the two WGs determine the distributionof energy between the HP and LP stages, and mean-while, they can adjust the EGR rate. Similarly, keep-ing the same EGR rates, the working lines of the HP

and LP compressors are shown in Figure 8(c) and (d),respectively. In the HP stage, the pressure ratio of theworking points first increases and then drops.Meanwhile, it always goes up in the LP stage.Because the HP turbine has more output power atthe low engine speeds. The WG1 must open for redu-cing the pumping work. Both the working lines are inthe high efficiency areas of the compressors.

Therefore, the mass flow ranges of all maps are rea-sonable. Meanwhile, the effect of pulse flow on theefficiency of the turbine maps is not considered in thepaper. First, the quantitative influence laws of the tur-bine efficiency on engine performances are studied inZhu and Zheng.44 The results showed that the turbineefficiency growth of 1% results in BSFC at the ratedpower point possibly decreasing by approximately0.12%, which is smaller for this work. Second, thiswork mainly compares engine performances with dif-ferent turbocharging technologies. All engines with theATST, VGT, and 2ST have pulse flow, and therefore,the influence of pulse flow on the performance differ-ences will be weakened. Therefore, the results of themodels are reliable and acceptable.

(a) (b)

(c) (d)

Figure 8. Compressor maps and working lines from the simulation: (a) engine-ATST, (b) engine-VGT, (c) HP stage of the engine-2ST,

and (d) LP stage of the engine-2ST.

Zhu et al. 9

Engine performance comparison

In this section, the performances of the engine withthe ATST, VGT, and 2ST are compared. The relativemagnitude trend will change with the engine speedand EGR rate. Meanwhile, this section will discussthe other main performances (cost, transient response,reliability, and complexity) of the three turbochargingtechnologies for the strategy.

Engine speed effect

In ‘‘Model validation’’ section, all models of the dieselengines with the ATST, VGT, and 2ST have the sameEGR rate at the same engine speed. Figure 9 showsthe EGR rate of three engine models versus 40–100%DES. The EGR rate rises with the engine speedincreasing, and the maximum value is about 31.0%at the rated power point (1900 r/min). Under this con-dition, the engine’s performances including enginetorque and BSFC versus 40–100% DES are containedin Figure 10(a) and (b). It can be clearly seen that therelative trend of the three engine performances haschanged with the engine speed. Before about 80%DES, the engine-2ST has the best performance, andthe engine-ATST has the minimum torque and high-est BSFC. At the maximum torque point (58% DES),the engine-2ST has torque and fuel economy improve-ments of approximately 7.9 and 1.0%, respectively,compared with the engine-ATST. By contrast, theengine-ATST has the best dynamic property andfuel economy beyond 80% DES. It has 1.0% lowerBSFC than the engine-2ST at the rated power point(100% DES).

Figure 11(a) and (b) illustrates the intake pressureand the pumping mean effective pressure (PMEP) ofthe three engines. At the engine low and mediumspeeds, engine-2ST can make better use of the exhaustgas energy because of the two turbines, and therefore,

it has the maximum pressurization capability. For theengine-VGT, it can decrease the turbine throat area toincrease the turbine power. Hence, it has a higherintake pressure than the engine-ATST. At the max-imum torque point, the maximum pressure ratio ofthe engine-ATST is only 2.9, but 3.2 and 3.5 forVGT and two-stage sections, respectively. Meanwhile,the engine-ATST has the minimum PMEP, so theBSFC is worst. However, at high-range enginespeeds, the intake pressure is limited for avoidingover boosting, and the three engines have the sameboost pressure. Because they have the same EGRrate, which is very high, the VGT and 2ST mustremain at high back-pressure. Therefore, they havesmaller PMEP resulting in a bad fuel economy. Forthe engine-ATST, the large scroll does not driveEGR, so it decreases the average exhaust pressure,resulting in lower fuel consumption. For example, thePMEP of the engine-ATST is 0.5 bar higher than thatof the engine-VGT at 100% DES.

EGR rate effect

Usually, at the low-range speed of the engine, theexhaust energy is not enough to boost for

Figure 9. EGR rate of the engine models with the ATST, VGT,

and 2ST versus the engine speed range of 40–100% from the

simulation.

(a)

(b)

Figure 10. Engine performance comparison of the three

models versus the engine speed range of 40–100% from the

simulation: (a) engine torque and (b) engine BSFC.


turbochargers. As mentioned in ‘‘Discussion of otherperformances’’ section, the VGT and 2ST can achievea higher pressure ratio. They have turbines with smal-ler throat areas for more output turbine power and ahigher EGR rate. Therefore, they have better engineperformances when the EGR is employed. However,the EGR rate needs to be higher and higher to meetstrict emission laws at high engine speeds.

Figure 12(a) and (b) shows the engine power andBSFC comparison of the three engine models versusthe EGR rate of 22–34% at 100% DES. The ordin-ates are the performance change rates relative to thevalues of the engine-ATST at the EGR rate of 22%.The performances of the three engines change with theEGR rate. Remaining the same EGR rate, the engine-2ST has better performances compared with theengine-VGT. This is because the two-stage sectioncan achieve higher boost-pressure and cascade useof energy. When the EGR rate is less than about27%, the engine-VGT has higher output power andlower fuel consumption than the engine-ATST.The former has higher boost-pressure because it candecrease the turbine throat area. However, the trend isinverse beyond 27% EGR rate, and the difference ismore prominent with increasing EGR. At 34% EGRrate, the engine-ATST has power and fuel economy

improvements of approximately 8.0 and 2.6%,respectively. The engine-ATST and the engine-2SThave a similar law, and the turning point is 29%EGR rate. At 22% EGR rate, the engine-2ST hasabout 7.0% higher power and 3.7% lower BSFC

(a)

(b)


models versus the engine speed range of 40–100% from the

simulation: (a) engine intake pressure and (b) engine PMEP.

(a)

(b)


models versus the EGR rate of 22–34% at 100% DES from the

simulation: (a) engine power and (b) engine BSFC (the ordin-

ates are the change rates relative to the values of the ATST at

the EGR rate of 22%).

Figure 13. Engine PMEP of the three models versus the EGR

rate of 22–34% from the simulation.

Zhu et al. 11

than the engine-ATST. In Figure 4(d), it can be cal-culated that the error of the EGR rate is �1%.

The engine PMEP of the three models versus theEGR rate of 22–34% is presented in Figure 13. When

the EGR rate is low, the VGT can adjust the throatarea to a more significant level, and the 2ST can changethe WG1 opening degree to adjust the exhaust pressurefor a higher PMEP. However, when the EGR rate ishigh, to increase the back-pressure, the VGT mustdecrease the throat area, and meanwhile, the 2STneeds to reduce the WG1 opening degree. This causesthe pumping loss to rising. For the ATST, this influencewill be diminished because the large scroll can decreasethe average back-pressure, and the slope of the PMEPversus the EGR rate is lower.

So, overall, the 2ST is better (<29% EGR rate),and the ATST is better (>29% EGR rate) when con-sidering the engine power and fuel consumption. IfEGR is not employed, the ASY of the ATST will be100%. At the rated power point, the 2ST and VGThave better performances (Figure 14(a)) as they havebrake mean effective pressure and PMEP improve-ments (Figure 14(b)). Therefore, the ATST is moresuitable to drive high EGR.

Discussion of other performances

The sections above have studied the engine dynamicperformance and fuel economy of engine-VGT,engine-2ST, and engine-ATST. This part mainly dis-cusses the other performances including the turbo-charger response time, complexity, reliability, andcost, which will provide better guidance for enginedesigners and manufacturers to choose the turbochar-ging technologies.

The performance comparison of the three turbo-charging technologies is shown in Figure 15(a) and(b). The ATST has an advantage in cost and reliabilitybecause of the simple mechanical structure and con-trol system, but the boost-pressure is lower because itsthroat area is unchanged. The ATST has one maindifference; it has two volutes with different throat

(a)

(b)


models at 100% DES when no EGR from the simulation:

(a) engine power and BSFC and (b) PMEP and BMEP.

(a) (b)

Figure 15. Performance comparison of the three turbocharging technologies: (a) low EGR requirement and (b) high EGR

requirement.


areas on the structure compared with the fixed geom-etry symmetric twin-scroll turbine, so they have simi-lar turbine performances. Saidur et al.5 stated that themost widely recognized problem with FGTs is turbo-charger lag: the poor transient response of the turbo-charger at low engine loads. Aghaali and Angstrom48

pointed out that the VGT and 2ST devices aredesigned to increase boost pressure at low speeds,reduce response times, and increase available torque.However, the more complex structures and layouts ofthe VGT and 2ST impose a burden of the control,reliability, and cost. He et al.49 were able to controlthe complexity and used a lookup table of enginespeed against fuel quantity to decide VGT positions inan open-loop manner, with EGR flow controlled in aclosed-loop fashion. It is very challenging for the 2ST toprove properties such as constraint satisfaction becausethe considered control structure is very complicated.The cost of a typical VGT ranges from 270 to 300%of the cost of the same size system and can offer gains ofapproximately 20% over comparable fixed geometryturbocharger systems, and meanwhile, the 2ST costsmore due to two turbochargers and the sophisticatedcontrol system.50 For engine performances, engine-2STand engine-VGT have better low-speed torque and fueleconomy through the above study in this paper.However, high-speed performances depend on theEGR rate. At low EGR requirement (Figure 15(a)),both engine-2ST and engine-VGT can decrease theexhaust back-pressure by regulating the flow distribu-tion of the 2STs and the turbine throat area, achievingpower, and BSFC improvements. However, the oppos-ite occurs at a high EGR requirement of about 29%(Figure 15(b)). The ATST is designed for high EGRbecause the turbine small volute provides back-pressurewhile the large volute decreases the average pressure.

Conclusions

This work established three engine models with theVGT, 2ST, and ATST, which were validated by theexperimental data. The comparison of the ATST,2ST, and the VGT on diesel engine performanceswas investigated in detail. The engine speed andEGR rate had important influences for the choice ofthe three turbocharging technologies, and the otherperformances including the cost, transient response,reliability, and complexity were also discussed. Themost significant results are as follows:

1. At low speeds, the VGT and 2ST have bettertorque and fuel economy, because they canachieve higher boost-pressure. By contrast, theATST has a large scroll without EGR to adjustthe average back-pressure, improving engine per-formances at the high-range of the engine speed.

2. The EGR has significant effects on the engine per-formance employed with the three turbochargingapproaches. For engine performances, the 2ST is

better than the VGT. Meanwhile, compared withthe VGT and 2ST, the ATST has performanceadvantages at high EGR rates.

3. Turbocharging routes are different under differentrequirements for EGR rates. There is an inflectionpoint in the relative advantages of asymmetric,variable geometry, and two-stage turbochargedengines. In the study, when the EGR rate islower than 29%, the two-stage turbochargingtechnology has the best performance. However,when the EGR rate is higher than 29%, the asym-metric turbocharging technology is the bestchoice. When considering other performances ofthe three turbocharging methods, the advantagesof the VGT and 2ST are the better transientresponse and low-speed performances. However,the ATST is more simple and cheaper.

The three turbocharging technologies (VGT, 2ST,and ATST) all have their own characteristics andsuperiority. For different EGR requirements, this art-icle may provide guidance to engine designers andmanufacturers on the turbocharging strategy undermore stringent legislation of ICEs.

Declaration of Conflicting Interests

The author(s) declared no potential conflicts of interest with

respect to the research, authorship, and/or publication ofthis article.

Funding

The author(s) disclosed receipt of the following financial

support for the research, authorship, and/or publicationof this article: National Science and Technology MajorProject (Grant No. 2017-II-0004-0016) and NationalNatural Science Foundation of China (GrantNo. 51876097).

ORCID iD

Xinqian Zheng https://orcid.org/0000-0003-1765-0416

References

1. Kumar BR and Saravanan S. Effect of iso-butanol

addition to diesel fuel on performance and emissionsof a DI diesel engine with exhaust gas recirculation.Proc IMechE, Part A: J Power and Energy 2016; 230:

112–125.2. Hatami M, Ganji DD and Gorji-Bandpy M. Numerical

study of finned type heat exchangers for ICEs exhaust

waste heat recovery. Case Studies Therm Eng 2014; 4:53–64.

3. Conklin JC and Szybist JP. A highly efficient six-strokeinternal combustion engine cycle with water injection for

in-cylinder exhaust heat recovery. Energy 2010; 35:1658–1664.

4. Wang T, Zhang Y, Zhang J, et al. Analysis of recover-

able exhaust energy from a light-duty gasoline engine.Appl Therm Eng 2013; 53: 414–419.

5. Saidur R, Rezaei M, Muzammil WK, et al. Technologies

to recover exhaust heat from internal combustionengines. Renew Sustain Energy Rev 2012; 16: 5649–5659.

Zhu et al. 13

https://orcid.org/0000-0003-1765-0416

6. Ghazikhani M, Hatami M, Ganji DD, et al. Exergyrecovery from the exhaust cooling in a DI dieselengine for BSFC reduction purposes. Energy 2014; 65:

44–51.7. Hatami M, Ganji DD and Gorji-Bandpy M.

Experimental and numerical analysis of the optimized

finned-tube heat exchanger for OM314 diesel exhaustexergy recovery. Energy Convers Manag 2015; 97:26–41.

8. Hatami M, Jafaryar M, Ganji DD, et al. Optimizationof finned-tube heat exchangers for diesel exhaust wasteheat recovery using CFD and CCD techniques. IntCommun Heat Mass Transfer 2014; 57: 254–263.

9. Guo J, Ge Y, Hao L, et al. Comparison of real-worldfuel economy and emissions from parallel hybrid andconventional diesel buses fitted with selective catalytic

reduction systems. Appl Energy 2015; 159: 433–441.10. Liu B, Zhang F, Zhao C, et al. A novel lambda-based

EGR (exhaust gas recirculation) modulation method

for a turbocharged diesel engine under transient oper-ation. Energy 2016; 96: 521–530.

11. Almeida FL, Zoldak P, Wang Y, et al. Multi-dimen-

sional engine modeling study of EGR, fuel pressure,post-injection and compression ratio for a lightduty diesel engine. In: Proceedings of the ASME 2014internal combustion engine division fall technical confer-

ence, Columbus, IA, USA, 19–22 October 2014,ICEF2014-5661.

12. Raptotasios SI, Sakellaridis NF, Papagiannakis RG,

et al. Application of a multi-zone combustion modelto investigate the NOx reduction potential of twostroke marine diesel engines using EGR. Appl Energy

2015; 157: 814–823.13. Cloudt IR, Baert IR, Willems IF, et al. SCR-only

concept for heavy-duty Euro VI applications. MTZWorldwide 2009; 70: 58–63.

14. Gao J, Fu W, Wang F, et al. Experimental and numer-ical investigations of tip clearance flow and loss in avariable geometry turbine cascade. Proc IMechE, Part

A: J Power and Energy 2018; 232: 157–169.15. Desantes JM, Benajes J, Garcıa A, et al. The role of the

in-cylinder gas temperature and oxygen concentration

over low load reactivity controlled compression ignitioncombustion efficiency. Energy 2014; 78: 854–868.

16. Tang H, Pennycott A, Akehurst S, et al. A review ofthe application of variable geometry turbines to thedownsized gasoline engine. Int J Engine Res 2015; 16:

810–825.

17. Lee M and Sunwoo M. Model-based control of a die-sel engine variable geometry turbine and exhaustgas recirculation system using a linear parameter vary-

ing methodology. Proc IMechE, Part D: J AutomobileEngineering 2014; 228: 1633–1643.

18. Ahmed FS, Laghrouche S, Mehmood A, et al.Estimation of exhaust gas aerodynamic force onthe variable geometry turbocharger actuator: 1D

flow model approach. Energy Convers Manag 2014;84: 436–447.

19. Galindo J, Tiseira A, Fajardo P, et al. Development and

validation of a radial variable geometry turbinemodel for transient pulsating flow applications.Energy Convers Manag 2014; 85: 190–203.

20. Hatami M, Cuijpers MCM and Boot MD.

Experimental optimization of the vanes geometry for

a variable geometry turbocharger (VGT) using adesign of experiment (DoE) approach. Energy ConversManag 2015; 106: 1057–1070.

21. Glenn BC, Upadhyay D, Utkin VI, et al. Observerdesign of critical states for air path flow regulationin a variable geometry turbocharger exhaust gas

recirculation diesel engine. Int J Engine Res 2011; 12:501–512.

22. Zamboni G and Capobianco M. Influence of high and

low pressure EGR and VGT control on in-cylinderpressure diagrams and rate of heat release in an auto-motive turbocharged diesel engine. Appl Therm Eng2013; 51: 586–596.

23. Llamas X and Eriksson L. Optimal transient control ofa heavy duty diesel engine with EGR and VGT. In: 19thIFAC World Congress, Cape Town, South Africa,

24–29 August, 2014, vol. 47, pp.11854–11859.24. Gelso ER and Dahl J. Air-path control of a heavy-duty

EGR-VGT diesel engine. In: 8th IFAC symposium on

advances in automotive control, Norrkoping, Sweden,20–23 June 2016, vol. 49, pp.589–595.

25. Wieske P, Luddecke B, Ewert S, et al. Optimisation of

gasoline engine performance and fuel consumptionthrough combination of technologies. MTZWorldwide 2009; 70: 30–36.

26. Zamboni G, Moggia S and Capobianco M. Hybrid

EGR and turbocharging systems control for low NOx

and fuel consumption in an automotive diesel engine.Appl Energy 2016; 165: 839–848.

27. Li H, Zhang G, Zhang H, et al. Equivalent matchingmodel of a regulated two-stage turbocharging systemfor the plateau adaptability. Proc IMechE, Part D: J

Automobile Engineering 2016; 230: 1654–1669.28. Faustmann C, Zerobin S, Spataro R, et al. On the

acoustics of a turning mid turbine frame with embeddeddesign in a two-stage test-turbine. Proc IMechE, Part

A: J Power and Energy 2015; 229: 529–538.29. Galindo J, Serrano JR, Climent H, et al. Impact of two-

stage turbocharging architectures on pumping losses of

automotive engines based on an analytical model.Energy Convers Manag 2010; 51: 1958–1969.

30. Avola C, Copeland C, Burke R, et al. Effect ofinter-stage phenomena on the performance predictionof two-stage turbocharging systems. Energy 2017; 134:

743–756.31. Yang M, Gu Y, Deng K, et al. Influence of altitude on

two-stage turbocharging system in a heavy-duty diesel

engine based on analysis of available flow energy. ApplTherm Eng 2018; 129: 12–21.

32. Zhao R, Li W, Zhuge W, et al. Characterization of two-stage turbine system under steady and pulsating flow

conditions. Energy 2018; 148: 407–423.

33. Zheng Z, Feng H, Mao B, et al. A theoretical and

experimental study on the effects of parameters oftwo-stage turbocharging system on performance of aheavy-duty diesel engine. Appl Therm Eng 2018; 129:

822–832.34. Schorn NA. The radial turbine for small turbocharger

applications: evolution and analytical methods for twin-

entry turbine turbochargers. SAE Int J Engines 2014;27: 1422–1442.

35. Chiong MS, Rajoo S, Romagnoli A, et al. One-dimen-sional pulse-flow modeling of a twin-scroll turbine.

Energy 2016; 115: 1291–1304.


36. Romagnoli A, Martinez-Botas RF and Rajoo S. Steadystate performance evaluation of variable geometry twin-entry turbine. Int J Heat Fluid Flow 2011; 32: 477–489.

37. Rajoo S, Romagnoli A and Martinez-Botas RF.Unsteady performance analysis of a twin-entry variablegeometry turbocharger turbine. Energy 2012; 38:

176–189.38. Muller M, Streule T, Sumser S, et al. The asymmetric

twin scroll turbine for exhaust gas turbochargers. In:

Proceedings of the ASME turbo expo 2008: power forland, sea and air, Berlin, 9–13 June 2008, paper GT2008-50614.

39. Hand MJ, Hellstrom E, Kim D, et al. Model and cali-

bration of a diesel engine air path with an asymmetrictwin scroll turbine. In: Proceedings of the ASME inter-nal combustion engine division fall technical conference,

Long Beach, CA, 2013. Paper ICEF2013-19134.40. Brinkert N, Sumser S, Weber S, et al. Understanding

the twin scroll turbine: flow similarity. J Turbomach

2013; 135: 021039.41. Kruger W, Kleffel J, Dietrich P, et al. 10.7-l Daimler

HD truck engine for Euro VI and Tier 4. MTZ world-

wide 2012; 73: 4–10.42. Chebli IE, Muller IM and Leweux DI. Development of

an exhaust-gas turbocharger for HD Daimler CVengines. Auto Tech Rev 2013; 2: 34–39.

43. Burny BD. New 14.8-l HD truck engine from Daimlerfor NAFTA. MTZ worldwide 2013; 74: 12–19.

44. Zhu D and Zheng X. Asymmetric twin-scroll turbo-

charging in diesel engines for energy and emissionimprovement. Energy 2017; 141: 702–714.

45. Zhu D and Zheng X. A new asymmetric twin-scroll

turbine with two wastegates for energy improvementsin diesel engines. Appl Energy 2018; 233: 263–272.

46. Zhu D and Zheng X. Fuel consumption and emissioncharacteristics in asymmetric twin-scroll turbochargeddiesel engine with two exhaust gas recirculation circuits.

Appl Energy 2019; 238: 985–995.47. Palenschat T, Mueller M, Rajoo S, et al. Steady-

state experimental and meanline study of an asym-

metric twin-scroll turbine at full and unequal andpartial admission conditions. SAE Technical Paper,2018-01-0971.

48. Aghaali H and Angstrom HE. A review of turbocom-pounding as a waste heat recovery system for internalcombustion engines. Renew Sustain Energy Rev 2015;49: 813–824.

49. He Y, Lin C and Gangopadhyay A. Integrated simula-tion of the engine and control system of a turbochargeddiesel engine. SAE Technical Paper, 2006-01-0439.

50. Martinez-Botas RF, Pesiridis A and Yang M. Overviewof boosting options for future downsized engines. SciChina Technol Sci 2011; 54: 318–331.

Appendix

Notation

r/min revolutions per minuter/s revolutions per second

Subscripts

1 turbine small scroll inlet2 turbine large scroll inlet

Zhu et al. 15

proc imeche part a: turbocharging strategy among variable

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