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Applied Energy

journal homepage: www.elsevier.com/locate/apenergy

A new asymmetric twin-scroll turbine with two wastegates for energyimprovements in diesel engines

Dengting Zhu, Xinqian Zheng⁎

Turbomachinery Laboratory, State Key Laboratory of Automotive Safety and Energy, Tsinghua University, Beijing 100084, China

H I G H L I G H T S

• A new asymmetric twin-scroll turbine with two wastegates (ATST-2WG) has been firstly presented.• Experiment and simulation are combined on the diesel engine with asymmetric turbocharger.• Wastegates control strategy and impact laws of asymmetry are studied.• The engine with ATST-2WG has the maximum fuel economy improvement of 2.91% compared to the engine with ATST-1WG.

A R T I C L E I N F O

Keywords:ATST-1WGATST-2WGAsymmetric turbineTwo wastegatesDiesel engineFuel economyEmission

A B S T R A C T

This paper first presented a new asymmetric twin-scroll turbine with two wastegates (ATST-2WG) for energyimprovements. An asymmetric twin-scroll turbine with one wastegate (ATST-1WG) is relatively simple and caneffectively solve the contradiction between low nitrogen oxide emissions and low fuel consumption when ex-haust gas recirculation is employed. However, its disadvantage is that the fuel economy will decrease at a partialopening degree of the exhaust gas recirculation valve, especially at a high-speed engine range. An experimentalinvestigation has been performed to calibrate the numerical model of a diesel engine equipped with an asym-metric twin-scroll turbine with one wastegate, and the engine with an asymmetric twin-scroll turbine with twowastegates model has also been especially established. Based on the models, both the wastegates control strategyand the critical parameter ASY turbine asymmetry (ASY, the ratio of the throat areas of the two scrolls) effectlaws have been studied, and they are different from the asymmetric twin-scroll turbine with one wastegate. Thebrake specific fuel consumption advantage first remains unchanged and then decreases as the engine speedincreases, and the maximum fuel economy improvement is 2.91% at the rated power point. The asymmetrictwin-scroll turbine with two wastegates has great advantages to achieve a better balance of engine emissions andenergy.

1. Introduction

Internal combustion engines are widely used and play an importantrole in industry. During the last two decades, engines have consumed alarge amount of fuel and led to considerable environmental pollution.At present, energy conservation and emission reduction are essentialwith greater energy shortages and environmental problems being,especially in the automobile and marine industries [1,2]. Since theCorporate Average Fuel Economy (CAFE) was first established in theUnited States in 1970s, the standards to improve fuel economy havebeen spreading worldwide [3,4]. The Euro 6 nitrogen oxide (NOx) limitfor diesel cars is 80mg/km, a reduction of over 95% compared to Euro1 emissions legislation [5,6]. Euro 6-compliant diesel passenger cars

feature lean NOx traps to satisfy the increasingly stringent NOx reg-ulations [7]. It is hard for the automakers to secure an optimal portfolioof fuel-efficient and emission reduced technologies that complies withtighter emission regulations and addresses rising fuel costs. Thestrengthened standards have driven engine manufacturers to use ex-haust gas recirculation (EGR) and turbocharger technologies in an ever-increasing number [8].

EGR is a well-accepted method to transport a fraction of exhaust gasback to the combustion chambers. Exhaust temperature is the key factorand the facet effect for diesel engine NOx emissions [9,10]. EGR de-creases the oxygen fraction inside the chambers as well as the peaktemperature during the combustion processes, so it effectively reducesNOx in the research of Raptotasios et al. [11] and Zhong et al. [12]. The

https://doi.org/10.1016/j.apenergy.2018.04.078Received 19 January 2018; Received in revised form 19 April 2018; Accepted 27 April 2018

⁎ Corresponding author.E-mail address: zhengxq@tsinghua.edu.cn (X. Zheng).

Applied Energy 223 (2018) 263–272

0306-2619/ © 2018 Published by Elsevier Ltd.

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EGR rate, defined as the mass percent of the recirculated exhaust in thetotal intake mixture, is the general parameter controlled by the positionof the EGR valve. In a proper range, NOx decreases with the increasingEGR rate. Ref. [13] investigated the effects of the proportion of highpressure and low pressure (HP/LP) EGR on engine operation. HP EGRsystems are most common for turbines, whereby exhaust gas is drawnfrom upstream of the turbocharger. EGR and turbocharging systemcontrol offers broad potential to lower NOx emissions and fuel con-sumption; the control reduced NOx emissions above 50% compared toEuro 5 levels [14]. Wei et al. [15] concluded that EGR techniques canreduce engine fuel consumption and meet more stringent emissionregulations in addition to other advanced techniques. At present, manyturbocharging technologies, including two-stage turbocharging, vari-able geometry turbines (VGT), symmetric twin-scroll turbines (STST)and ATST-1WG are widely combined with EGR in diesel engines.

To further increase waste energy recovery and improve engineperformance, two turbochargers of different sizes can be connected toform a two-stage turbocharging system. In a two-stage turbochargingsystem, the HP turbocharger is smaller than that of the LP in order toachieve a better transient response at low speeds; the LP turbocharger islarge and is optimized for maximum power output operation [16].Single-stage turbocharging does not typically maintain high boostpressure and a heavy EGR rate due to limited overall turbochargingefficiency, especially at low-speed engine ranges [17]. Therefore, two-stage turbocharging is widely adopted for vehicles and small aircrafts[18]. Compared with single stage turbocharging, two-stage turbochar-ging provides flexibility to meet engine requirements at both low andhigh speeds because of load split. Both LP and HP stages can operate atreduced flow and pressure ratio ranges. However, two-stage turbo-charging has more complicated mechanical structures and control sys-tems to achieve smooth operation during stage switching. The perfor-mance accuracy measurement for mapping turbocharging systems insteady turbocharger gas-stands is difficult to ensure due to aero-thermalinter-stage phenomena [19]. The disadvantages of two-stage turbo-charging are complicated piping, valve and seal systems, and a con-siderable weight penalty. Two-stage turbocharging systems also havelarger flow passage volume and more metal surface than single stagesystems, and this can affect the time taken by the turbocharger to warmup from the cold start, thus affecting the operation of the downstreamcatalyst converter and engine cold start emissions [20].

The most widely recognized problem with fixed geometry devices isturbocharger lag, which is the poor transient response of the turbo-charger at low engine loads [21]. Therefore, VGT is a well-accepted andpotential technology to increase boost-pressure at low speeds and re-duce response times [22]. VGT can change the turbine throat area andprovide enough backpressure to drive EGR and allows good handling of

fuel injection and inlet air charge flow into the combustion chamber[23,24]. In VGT devices, the aspect ratio will determine the EGR flow,and the EGR rates are fixed by adjusting the VGT position, since itgoverns the pressure difference between the inlet manifold and exhaustmanifold [25]. Therefore, EGR and VGT are combined to control andoptimize the fuel consumption by minimizing pumping losses [26,27].VGT offers improved turbocharger rotational speed, engine speed andboost-pressure over a regular turbocharger and allows the performanceof the turbocharger to be optimized across the whole engine range[28,29]. Furthermore, the trend of actuating VGT devices is shiftingfurther towards electrical and hydraulic variants that allow more deli-cate control than pneumatic controls. Variable two-stage turbochargingsystems that may regulate exhaust enthalpy and matching points to thehigh efficient zone under different operating conditions will be widelyused [16]. However, VGT has very sophisticated control systems tomatch with the EGR system and the engine system. The strength andreliability of the adjustable vanes are very fundamental [30], and thevanes are expensive. In the same production volume, the cost of a ty-pical VGT ranges from 270% to 300% of the cost of the same sizesystem and can offer gains of approximately 20% over comparable fixedgeometry turbocharger systems [31].

The twin-scroll turbine is a meridionally divided turbine, and thescroll has a single divider around the entire perimeter of the housing.Each inlet feeds the entire rotor circumference. It was first proposed in1954 [32,33]. The STST, which has two inlet scrolls whose shapes andareas are uniform, has traditionally seen wider use on multiple-cylinderengines by turbocharger manufacturers due to its inexpensive andsimple design. A comparison between the twin-scroll (meridionallydivided) and double-scroll (circumferentially divided) turbines revealedtheir very distinct efficiency characteristic [34]. A double-scroll turbinehas been shown to deliver higher peak efficiency at full admissionconditions. Therefore, a twin-scroll turbine showed lesser deteriorationat partial admission conditions because the flow was still capable ofexpanding into the larger rotor inducer area, even though not entirely[35,36]. Chiong et al. [37,38] presented a revised one-dimensionalpulse flow modeling of twin-scroll turbocharger turbine under pulseflow operating conditions. The results showed that a twin-scroll turbinedoes not operate at full admission throughout the in-phase pulse flowconditions. Instead, the turbine worked at an unequal admission statedue to the magnitude disparity of the turbine inlet flow. Rajoo et al.[39] discussed the details of unsteady experimentation and analysis of atwin-scroll variable geometry turbine for an automotive turbocharger.The cycle-averaged efficiency of the twin or single-scroll nozzled tur-bine was found to depart significantly from the equivalent quasi-steady.In comparison to the nozzleless single-scroll turbine, the departure wasas much as 32%. When STST is used to drive EGR, both two-exhaust

Nomenclature

φ turbine throat arearpm revolutions per minuterps revolutions per second

Subscripts

1 small scroll inlet2 large scroll inlet

Abbreviations

ASY turbine scroll asymmetryATST asymmetric twin-scroll turbineATST-1WG asymmetric twin-scroll turbine with one wastegateATST-2WG asymmetric twin-scroll turbine with two wastegates

BSFC brake specific fuel consumptionC compressorDI direct injectionDL dual loopEGR exhaust gas recirculationHP high pressureLP low pressureNOx nitrogen oxidesOPD the opening degree of the EGR valvePMEP pumping mean effective pressureRES the relative engine speedSTST symmetric twin-scroll turbineT turbineVGT variable geometry turbineWG wastegate

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passages are linked with the EGR passage [40], which deteriorates fueleconomy because of high pumping losses.

The ATST is a potential turbocharging technology that can effec-tively solve the problem in the STST [41]. The ATST first emerged inthe last century, which was used to keep the overall fuel consumptionincrease as low as possible on the 6-cylinder truck engines equippedwith an EGR system by Daimler-Benz. It has two scrolls with differentthroat areas, and the two scrolls are separately connected with the twogroups of cylinders. In the engine, one group of cylinders linked withthe small scroll is operated with a high exhaust gas backpressure, andthe other is operated with a fuel-saving low exhaust gas backpressure[42]. Therefore, the small scroll can offer a higher backpressure to driveEGR. On the other hand, the large scroll is isolated from the EGR pas-sage and can reduce the average backpressure for better fuel economy.Daimler Trucks launched a generation of new Mercedes-Benz dieselengines for medium-duty [43,44] and heavy-duty [45,46] commercialvehicles with the institution of the Euro 6 emission standard. ATST hasbeen developed as the core of the EGR system for the reduction of NOxemissions and applied the system to the modern generation of Daimlercommercial vehicle engines [47]. In 2007, a heavy-duty diesel enginewith 14.8 L displacement called OM 472 was introduced into themarket. After that, Daimler continued launching new products, in-cluding the 12.8 L OM 471, the 15.6 L OM 473 and the 10.7 L OM 470[48,49]. The OM 470, which was initially developed to meet the Euro 6emissions limits, has lower NOx emissions and fuel economy advantageof up to 5% compared to OM 457 in the Euro 5 setting [50]. At present,ATST normally has one WG linked with the large scroll to avoid over-boost pressure. However, when the EGR valve is not completely open atthe medium and high-speed range of the engine, the backpressure of thesmall scroll will increase to cause adverse fuel economy because ofmore pumping losses.

This paper first presented a new ATST-2WG, which has two WGscalled WG1 and WG2. The two WGs are connected with the small scrolland the large scroll, respectively. When the backpressure of the smallscroll is so high that it causes too high an EGR rate, the ATST-2WG canopen the WG1 to decrease the backpressure for better fuel economy toensure engine power. Compared with VGT, two-stage turbochargers,STST and ATST-1WG, the ATST-2WG are relatively simple and can ef-fectively solve the problem that fuel economy decreases at partial EGRvalve opening degrees. This paper consists of three parts. First, an en-gine experiment with ATST is presented, and the simulation models areestablished. Second, both the wastegates control strategy and the cri-tical parameter ASY impact laws are explored. Finally, the advantagesof ATST-2WG are evaluated in comparison with ATST-1WG.

2. Experiment and simulation methods

A schematic diagram of a 6-cylinder inline diesel engine with anATST-1WG is shown in Fig. 1(a). The six cylinders are equally dividedinto two groups, which are respectively connected with two scrollpassages. The backpressure p1 of the small scroll is higher than p2 of thelarge scroll to drive a high enough EGR rate; therefore, the small scrollis also called the EGR scroll [51,52]. There is only one WG on the largescroll passage for an acceptable boost-pressure. Different engine op-erations always lead to unequal two scrolls admission conditions, and itis difficult to match the ATST with the engine. The turbine asymmetry(ASY), which is commonly defined by the ratio of the throat areas of thetwo scrolls, is a key parameter in the ATST design.

= ×ASYφφ

100%12 (1)

In which φ1 and φ2 are the throat areas of the two scrolls. ASY canquantitatively evaluate the difference between the two scrolls.

In the study, an inline 6-cylinder direct injection (DI) diesel engineequipped with an ATST-1WG (ASY=53%) is employed on a dynam-ometer test bench according to the structure of illustrated in Fig. 1(a).

Details of the engine specifications are given in Table 1. The speed andload of the engine are regulated using a C 500 eddy current dynam-ometer with an accuracy of± 10 rpm and±1.25 Nm, and the max-imum speed is 4500 rpm. The maximum brake torque is 4000 Nm(1000–1700 rpm) and the rated power is 720 kW (1700–4500 rpm). AnAVL735S fuel consumption measuring instrument with a precision of0.12% of the recorded value is used to meter the engine fuel con-sumption, and it has a measuring range of 0.1–110 kg/h. The data ac-quisition system is a PUMA OPEN 1.2 all-in-one bench-top instrument,and the data is processed by CONCERTO-P software. The engine intakemass flow is collected by a Sensyflow P/4000 air flow meter (accu-racy:± 5 mg), and the temperature of the intake gas boosted is con-trolled within 30–55 °C by a BATCON device. The engine fuel andcoolant have constant temperature regulated by an AVL753 fuel con-stant temperature control device and an AVL553 coolant constanttemperature control device. The fuel and coolant temperature arecontrolled within the range of 15–80 °C and 70–120 °C, respectively.The experiment engine has an intercooled EGR system whose EGR ratescan be sampled in real time by a gas emission analyzer MEXA-7100DEGR. Moreover, the environmental conditions are accuratelycontrolled by the engine intake and exhaust environment simulationsystem ACS2400. The environmental temperature and pressure are setto 298 ± 1 K and 100 ± 1 kPa, respectively. The diesel engine oper-ates at full load and different speeds.

According to the experiment, the engine simulation model is ac-complished in Fig. 1(b) by GT-POWER v7.3.0 software, which is com-monly used in engine cycle simulations [53]. The pipes are modeledaccording to the experimental pipe shape and length, and the frictionand heat transfer multiplier are within a reasonable range in reference

Fig. 1. A 6-cylinder diesel engine with an ATST-1WG: (a) the schematic dia-gram; (b) the simulation model.

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to experimental pipes materials. An orifice joins the two entry pipes atthe turbine inlet, which can model backflow [54]. In the turbochargermodel, the compressor and ATST (ASY=53%) maps are obtained fromturbocharger experiments on the gas test stand during the entire de-velopment phase, as shown in Fig. 2. The parameter “Wastegate AreaFraction for Entry 1” is set 0 in the turbine model, as shown in Fig. 1(b),which means the WG is only linked with the large scroll. In the enginecombustion model, the experimental combustion rate is adopted. It isdefined as the fuel combustion quality with unit time and can bemeasured by an AVL641 combustion analyzer. At full load, the ex-perimental combustion data is as shown in Fig. 3. The relative enginespeed (RES) is normalized to the maximum engine speed at the full load

(1900 rpm).

= ×RESEngine speed

rpm1900100%

(2)

The results of the single-zone and two-zone combustion models areshown below. It can be seen that two combustion models have the sametrend. Meanwhile, the Woschni Formula is chosen reasonably in thecylinder heat transfer model.

= ⎡⎣⎢

+ − ⎤⎦⎥

− −α p T D C C C T Vp V

p p820 ( )g m a sa a

0.8 0.53 0.21 2 0

0.8

(3)

where αg is the instantaneous average heat transfer coefficient for theworking gas and chamber wall; p and T are the working gas pressureand temperature, respectively, D is the cylinder diameter, Cm is theaverage piston speed, C1 is the gas velocity coefficient, pa, Ta and Vaare working gas pressure, temperature and cylinder volume at the be-ginning of compression, respectively, andp0is cylinder pressure of theinverted engine. Therefore, the heat exchange for the gas and wall (Qw)are calculated as Eq. (4):

∑ ∑= = −dQwdφ

dQwidφ θ

αg Ai T T1 · ( )wi1

3

1

3

(4)

in which θ is the angular velocity, A is the heat transfer area, Tw is theaverage wall temperature and i =1, 2 and 3 respectively represent thecylinder head, piston and cylinder liner.

The engine model calculation is finished, and the experimental andsimulation results including torque, power, brake specific fuel con-sumption (BSFC) and EGR rate are compared, as shown in Fig. 4. Thesimulation results and the experiment results at full load have goodagreement, with only a small deviation. For the purpose of this work,the deviations are acceptable.

This paper proposes a new ATST-2WG concept and the whole en-gine system schematic is shown in Fig. 5(a). This ATST has two WGs,which are connected with the small scroll and the large scroll, respec-tively. Based on the engine simulation model with an ATST-1WG, theengine model equipped with an ATST-2WG is established in Fig. 5(b).The engine system, EGR system and compressor are unchanged, andonly the turbine structure is different compared to Fig. 1(b). The“Wastegate Area Fraction for Entry 1” parameter is also set as 0 in theturbine model, which means this WG is only linked with the large scroll.WG1 connects the small scroll to the atmosphere and is as shown inFig. 5(a).

3. EGR valve effects of ATST-1WG on engine performance

For the engine with an ATST-1WG, EGR valve adjustment can

Table 1Test engine specifications.

Items Value and unit

Engine type Inline 6-cylinder DI dieselNumber of valves per

cylinder4 (2 inlet / 2 exhaust)

Bore 129mmStroke 160mmDisplacement 12.55 LCompression ratio 18.2:1Cooling system Water cooledAir intake system Intercooled asymmetric twin-scroll turbocharger

(ASY=53%)EGR system Intercooled EGRRated power 351 kW (1900 rpm)Maximum torque 2380 Nm (1000–1400 rpm)

(a)

(b)Fig. 2. Turbocharger maps: (a) compressor; (b) ATST (ASY=53%).

0.00

0.01

0.02

0.03

0.04

0.05

0.06

0.07

120

Com

bust

ion

rate

[%]

Crank angle [deg]

42%

53%

58%

68%

84%

100%

RES

Fig. 3. Experimental combustion rate data.

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change the EGR rates. This section mainly discusses the EGR valve ef-fects of ATST-1WG on engine performance. In the model, the EGR valveis an orifice model that will be adjusted by modifying the hole diameter.The opening degree of the EGR valve (OPD) is a key parameter, which isdefined as the ratio of the hole diameter and the maximum diameter inthe experiment. Based on the model in Fig. 1(b), the remaining ASY is53%, and by changing OPD at a range of 0–100% at different enginespeeds, the effects on EGR rates and BSFC are presented in Fig. 6.

In Fig. 6(a) and (b), the relative EGR rate and relative BSFC are non-dimensionalized by dividing the EGR rate and BSFC at the point whenboth the OPD and RES are 100%. It can be seen that the EGR ratescontinue with increasing OPD, and the increasing trend becomes largerwith increasing engine speed. On one hand, the EGR rate first increasesand then essentially remains unchanged, and BSFC has a small upwardtrend as the OPD increases at low engine speeds. It is well known thatthe exhaust gas is not usually enough for engines to boost inlet gas;therefore, the turbine WG is usually closed at low engine speeds. In thisdiesel engine, the WG, which is connected with the large volute, iscompletely closed under RES 58% (engine maximum torque point).Given a smaller OPD, there is more boosting exhaust gas; therefore, theturbine power will rise. The engine intake pressure increases resultingin better power and BSFC. When the OPD is beyond 60%, the EGRdriving pressure comes to a maximum value, so the EGR rate remainsunchanged. On the other hand, the WG is gradually open as the enginespeed increases at high speeds to avoid overboost pressure. The exhaustbackpressure is enough to drive EGR, and it is higher with increasingspeed; thus, the EGR rate continually increases. When the OPD reaches

(a)

(b)

(c)

(d)Fig. 4. Comparison of simulation and experiment results: (a) torque, (b) power,(c) BSFC, and (d) EGR rates.

p1p2EGR valve

WG2

T

EGR Cooler

Charge Air

Cooler

C

EGR linked with the small scroll

WG1

ATST-2WG

Engine system

EGR systemWG linked with the small scroll

(a)

(b)Fig. 5. A 6-cylinder diesel engine with an ATST-2WG: (a) the schematic dia-gram; (b) the simulation model.

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100%, the EGR rate will remain unchanged. Meanwhile, the back-pressure of the small scroll decreases with the increasing OPD, resultingin smaller pumping losses and better fuel economy. At 100% RES, theBSFC at 0 OPD is 14.6% higher than that at 100% OPD.

Fig. 6(c) illustrates the engine pumping mean effective pressure(PMEP) versus an OPD range from 0 to 100% at different engine speedsin the engine cycle simulation. PMEP increases when OPD increasesbecause the smaller throat area leads to a higher exhaust backpressure.This influence law is more apparent at the high-speed engine range.Using RES=100% as an example, it can be seen that the two exhaustpassages have higher pressures compared to the intake pressure inFig. 7. The differences are smaller when OPD increases. It was men-tioned before that the EGR circuit is only linked with the small volute.

The EGR valve diameter is the smallest in all passages of the EGRsystem, and the diameter directly influences the exhaust backpressureof the small scroll, whose effect is the same as the throat area of thesmall scroll. Meanwhile, the exhaust backpressure of the large scrollalso changes because of backflow and the interaction of the two scrolls.For example, comparing the pressure of the small scroll with the intakepressure, the relative pressure is 1.18 at 0 OPD but 0.14 at 100% OPD;the non-dimensionalized value is determined by dividing the intakepressure value at 100% OPD. The values are much smaller whencomparing the pressure of the large scroll with that of the intake as 0.18at 0 OPD and 0.12 at 100% OPD. Therefore, the average exhaustbackpressure is obviously higher than the intake gas pressure. As pre-viously mentioned, a smaller OPD means a smaller throat area, re-sulting in a high exhaust backpressure and a bad gas exchange condi-tion.

4. Comparison of ATST-1WG and ATST-2WG

In this section, the cycle simulation results of the engine modelsequipped with an ATST-1WG and an ATST-2WG are compared to de-termine the potential of ATST-2WG over ATST-1WG. To achieve abetter balance between NOx emissions, fuel economy and power output,it is important to study the wastegates control strategy and the effects ofthe turbocharger key parameter ASY, which are also explored in thefollowing sections.

(a)

(b)

(c)Fig. 6. Engine performances versus an OPD range of 0–100% at different enginespeeds in engine cycle simulation: (a) relative EGR rate; (b) relative BSFC; (c)PMEP.

Fig. 7. Intake air pressure and two exhaust gas scrolls backpressures at 100%RES.

Fig. 8. WG relative opening degree in the engine with an ATST-1WG at dif-ferent engine speeds and EGR valve OPDs.

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4.1. Wastegates control strategy

For the engine with the ATST-1WG with one wastegate and one EGRvalve, the boost-pressure and EGR rate can be adjusted under variableoperations. In contract, the EGR valve is unnecessary in the engine withthe ATST-2WG, and the WG1 can vary the small scroll backpressure fordifferent EGR rates. Therefore, the ATST-2WG has a new controlstrategy with respect to the ATST-1WG. At different OPDs of the EGRvalve, the WG relative opening degrees in the ATST-1WG are as shownin Fig. 8. The WG relative opening degrees are non-dimensionalized bydividing the WG opening degree at 60% OPD and 100% RES. Within thespeed range at the maximum torque point, the WG is completely closedfor boosting. Beyond that, an increasing engine speed opens the WGbecause the exhaust backpressure and mass flow rate increase, and theboost-pressure is limited to a constant value. Meanwhile, the exhaustbackpressure and mass flow rate will also increase when the EGR valveis gradually closed; therefore, the WG opening degree increases.

Given the same EGR rates of the engine with the ATST-1WG, theWG1 and WG2 relative opening degrees in the engine with the ATST-2WG are taken from Fig. 9(a) and (b), respectively. At 100% OPD, theWG1 is fully closed, and the EGR rate maximizes. Beyond the speed atthe maximum torque point, the WG1 opening degree can regulate thesmall scroll pressure, and its increasing can decrease the EGR rate. Toensure the boost-pressure, the WG2 needs to be closed little by little asmore exhaust gas passes through the WG1 instead of the turbine.

4.2. ASY effects on engine performances

As mentioned earlier, ASY is a very critical parameter, whichcharacterizes the interrelationships of the two scrolls and impacts the

balance between the engine emissions and fuel consumption. Thissection will investigate the ASY effects on engine performances com-paring the ATST-1WG and ATST-2WG. First, based on the enginemodels with an ATST-1WG and an ATST-2WG discussed in Section 2,the engine cycle simulations are conducted using the same enginesystem, EGR system and compressor. Given the same EGR rate at thesame OPD and 100% RES for ATST-1WG and ATST-2WG, all modelsoperate at the full load engine condition, and the ASY values are from30% to 80%.

The change rates of the power and BSFC in the engine with an ATST-

(a)

(b)Fig. 9. WGs relative opening degree in the engine with an ATST-2WG at dif-ferent engine speeds and EGR valve OPDs: (a) WG1 and (b) WG2.

(a)

(b)

Fig. 10. Engine performance change rates in the engine with an ATST-2WGwith respect to the engine with an ATST-1WG at different ASY and OPDs: (a)power; (b) BSFC.

Fig. 11. The intake and two exhaust gas scrolls backpressures of the ATST-1WGand ATST-2WG.

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2WG with respect to the engine with an ATST-1WG are shown inFig. 10. Under the same conditions, the engines with an ATST-2WG orATST-1WG have the same power and BSFC at 100% OPD from 30% to80% ASY. At 100% OPD, the WG1 is completely closed to offer themaximum EGR rate. Therefore, the WG1 has no influence on engineperformance. However, at partial OPD conditions, the dynamic per-formance and engine fuel economy with an ATST-2WG are clearlybetter than those of the engine with an ATST-1WG. The power hasimprovements of 2.76% and 2.69% for BSFC at 60% OPD. With in-creasing ASY, the power and BSFC improvements decrease since ahigher ASY usually means a larger small scroll throat area, thus

decreasing the backpressure. For example, at 100% RES and 70% OPD,the pressures of the small scroll, large scroll and intake passage areshown in Fig. 11. It can be seen that the pressures of the two turbinevolutes in ATST-2WG are lower than that in ATST-1WG and thereforethe ATST-2WG has lower pumping loss, which is presented at 80% and70% OPD in Fig. 12. A smaller large scroll throat area leads to a higheraverage pressure and a bad charge air condition as ASY rises. WhenOPD is 80%, the exhaust backpressure is lower than that at 70% OPD.The pumping loops at 70% OPD are shown in comparison between theATST-2WG and the ATST-1WG in Fig. 13. Clearly, it can be seen thatpumping losses in the cylinders with the large scroll are similar butmuch less in the cylinders with the small scroll, which also proves thatthe EGR valve has a great impact on the small scroll performance.

4.3. Fuel economy improvements in the ATST-2WG

In the sections above, it is clearly demonstrated that the ATST-2WGhas greater potential than ATST-1WG on engine performances. Given53% ASY and the same EGR rates, the maximum fuel economy im-provements in the ATST-2WG over the ATST-1WG at different enginespeeds are shown in Fig. 14. All engine models operate at full load, andthe engines with the ATST-1WG have 60% OPD. The BSFC change ratefirst remains unchanged and then decreases as engine speed increases.The maximum fuel economy improvement is 2.91% at 100% RES (therated power point). When the engine speed is beyond 58% RES (themaximum torque point), its growth of 10% results in a decrease in theBSFC rate by approximately 0.7%. For the diesel engines with an ATST-1WG, the WG2 is completely closed at low engine speeds and graduallyopens over the maximum torque point speed. Therefore, for engineswith an ATST-2WG, the WG1 is also completely closed to ensure enginepower leading to no BSFC advantage under 58% RES. With the in-creasing engine speed, the WG in ATST-1WG is usually open to avoidoverboost pressure, and the 60% OPD will increase the backpressure ofthe small volute, which causes the fuel economy to deteriorate. If theEGR rate demand is lower, the OPD will be smaller and the fueleconomy will further deteriorate. By contrast, in the ATST-2WG, theEGR valve is completely open, and the WG1 is properly open to achievethe same EGR rates as the ATST-1WG. Therefore, the exhaust back-pressures of the two scrolls will decrease for better charge air condi-tions, especially at high engine speeds.

5. Conclusions and remarks

A new asymmetric twin-scroll turbine with two wastegates is firstintroduced to achieve a better balance between energy and NOx emis-sions under more stringent legislation. To fully exploit its potential, this

Fig. 12. Engine PMEP at 80% and 70% OPD in the engines with an ATST-1WGand an ATST-2WG versus the ASY range of 30–80%.

(a)

(b)

Fig. 13. Pumping loop diagrams of ATST-1WG and ATST-2WG at 50% ASY and70% OPD: (a) cylinder with large scroll; (b) cylinder with small scroll.

Fig. 14. BSFC change rates in the ATST-2WG over the ATST-1WG at 53% ASYand 60% OPD versus different engine speeds.

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work compared the performance with that of engines with an asym-metric twin-scroll turbine with one wastegate and investigated thewastegates control strategy and the critical ASY influence parameter.The most significant results can be summarized as follows:

(1) The asymmetric twin-scroll turbine with one wastegate has an in-herent vice that the fuel economy deteriorates when the EGR valveis partially open at the high-speed range of the engine because ofhigher exhaust backpressure, especially for the turbine small scroll.

(2) The asymmetric twin-scroll turbine with two wastegates has a newwastegates control strategy. The EGR valve is needless, and the EGRrate can be adjusted by the WG1. Within the speed at the maximumtorque point, the two WGs are fully closed. Beyond that, an in-creasing WG1 opening degree decreases the EGR rate, and the WG2must be gradually closed to ensure the boost-pressure at the sametime.

(3) The turbine asymmetry has great effects on diesel engine perfor-mances. The asymmetric twin-scroll turbine with two wastegateshas energy improvements compared with the asymmetric twin-scroll turbine with one wastegate, and the power and fuel economyimprovements decrease with asymmetry increasing. The BSFCchange rate first remains unchanged and then decreases as theengine speed increases and the maximum fuel economy improve-ment is 2.91% at the rated power point.

Compared with VGT, two-stage turbochargers, symmetric twin-scroll turbines and asymmetric twin-scroll turbines with one wastegate,the asymmetric twin-scroll turbine with two wastegates is relativelysimple and can effectively solve the problem of decreasing fueleconomy at partial EGR valve opening degree. This work can providedesign guidelines for turbocharged engine designers, and the tech-nology can be equipped with internal combustion engines to meet morestringent fuel consumption and emissions legislations. The engine op-erating points at variable loads and speeds will be considered andfurther studied to achieve better engine emissions and energy balance.

Acknowledgments

This research was supported by the National Natural ScienceFoundation of China (Grant No. 51176087).

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A new asymmetric twin-scroll turbine with two wastegates for energy improvements in diesel enginesIntroductionExperiment and simulation methodsEGR valve effects of ATST-1WG on engine performanceComparison of ATST-1WG and ATST-2WGWastegates control strategyASY effects on engine performancesFuel economy improvements in the ATST-2WG

Conclusions and remarksAcknowledgmentsReferences