A new asymmetric twin-scroll turbine with two wastegates for energy ...

[Pages:11]Applied Energy 223 (2018) 263?272

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Applied Energy

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A new asymmetric twin-scroll turbine with two wastegates for energy T

improvements in diesel engines

Dengting Zhu, Xinqian Zheng

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

HIGHLIGHTS 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.

ARTICLE INFO

Keywords: ATST-1WG ATST-2WG Asymmetric turbine Two wastegates Diesel engine Fuel economy Emission

ABSTRACT

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

1. Introduction

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

feature lean NOx traps to satisfy the increasingly stringent NOx regulations [7]. It is hard for the automakers to secure an optimal portfolio of fuel-efficient and emission reduced technologies that complies with tighter emission regulations and addresses rising fuel costs. The strengthened standards have driven engine manufacturers to use exhaust gas recirculation (EGR) and turbocharger technologies in an everincreasing number [8].

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

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

Received 19 January 2018; Received in revised form 19 April 2018; Accepted 27 April 2018 0306-2619/ ? 2018 Published by Elsevier Ltd.

D. Zhu, X. Zheng

Nomenclature

turbine throat area

rpm

revolutions per minute

rps

revolutions per second

Subscripts

1

small scroll inlet

2

large scroll inlet

Abbreviations

ASY

turbine scroll asymmetry

ATST asymmetric twin-scroll turbine

ATST-1WG asymmetric twin-scroll turbine with one wastegate

ATST-2WG asymmetric twin-scroll turbine with two wastegates

Applied Energy 223 (2018) 263?272

BSFC C DI DL EGR HP LP NOx OPD PMEP RES STST T VGT WG

brake specific fuel consumption compressor direct injection dual loop exhaust gas recirculation high pressure low pressure nitrogen oxides the opening degree of the EGR valve pumping mean effective pressure the relative engine speed symmetric twin-scroll turbine turbine variable geometry turbine wastegate

EGR rate, defined as the mass percent of the recirculated exhaust in the total intake mixture, is the general parameter controlled by the position of the EGR valve. In a proper range, NOx decreases with the increasing EGR rate. Ref. [13] investigated the effects of the proportion of high pressure and low pressure (HP/LP) EGR on engine operation. HP EGR systems are most common for turbines, whereby exhaust gas is drawn from upstream of the turbocharger. EGR and turbocharging system control offers broad potential to lower NOx emissions and fuel consumption; the control reduced NOx emissions above 50% compared to Euro 5 levels [14]. Wei et al. [15] concluded that EGR techniques can reduce engine fuel consumption and meet more stringent emission regulations in addition to other advanced techniques. At present, many turbocharging technologies, including two-stage turbocharging, variable 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 engine performance, two turbochargers of different sizes can be connected to form a two-stage turbocharging system. In a two-stage turbocharging system, the HP turbocharger is smaller than that of the LP in order to achieve a better transient response at low speeds; the LP turbocharger is large and is optimized for maximum power output operation [16]. Single-stage turbocharging does not typically maintain high boost pressure and a heavy EGR rate due to limited overall turbocharging efficiency, especially at low-speed engine ranges [17]. Therefore, twostage turbocharging is widely adopted for vehicles and small aircrafts [18]. Compared with single stage turbocharging, two-stage turbocharging provides flexibility to meet engine requirements at both low and high speeds because of load split. Both LP and HP stages can operate at reduced flow and pressure ratio ranges. However, two-stage turbocharging has more complicated mechanical structures and control systems to achieve smooth operation during stage switching. The performance accuracy measurement for mapping turbocharging systems in steady turbocharger gas-stands is difficult to ensure due to aero-thermal inter-stage phenomena [19]. The disadvantages of two-stage turbocharging are complicated piping, valve and seal systems, and a considerable weight penalty. Two-stage turbocharging systems also have larger flow passage volume and more metal surface than single stage systems, and this can affect the time taken by the turbocharger to warm up from the cold start, thus affecting the operation of the downstream catalyst converter and engine cold start emissions [20].

The most widely recognized problem with fixed geometry devices is turbocharger lag, which is the poor transient response of the turbocharger at low engine loads [21]. Therefore, VGT is a well-accepted and potential technology to increase boost-pressure at low speeds and reduce response times [22]. VGT can change the turbine throat area and provide 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 it governs the pressure difference between the inlet manifold and exhaust manifold [25]. Therefore, EGR and VGT are combined to control and optimize the fuel consumption by minimizing pumping losses [26,27]. VGT offers improved turbocharger rotational speed, engine speed and boost-pressure over a regular turbocharger and allows the performance of the turbocharger to be optimized across the whole engine range [28,29]. Furthermore, the trend of actuating VGT devices is shifting further towards electrical and hydraulic variants that allow more delicate control than pneumatic controls. Variable two-stage turbocharging systems that may regulate exhaust enthalpy and matching points to the high efficient zone under different operating conditions will be widely used [16]. However, VGT has very sophisticated control systems to match with the EGR system and the engine system. The strength and reliability of the adjustable vanes are very fundamental [30], and the vanes are expensive. In the same production volume, the cost of a typical VGT ranges from 270% to 300% of the cost of the same size system and can offer gains of approximately 20% over comparable fixed geometry turbocharger systems [31].

The twin-scroll turbine is a meridionally divided turbine, and the scroll has a single divider around the entire perimeter of the housing. Each inlet feeds the entire rotor circumference. It was first proposed in 1954 [32,33]. The STST, which has two inlet scrolls whose shapes and areas are uniform, has traditionally seen wider use on multiple-cylinder engines by turbocharger manufacturers due to its inexpensive and simple design. A comparison between the twin-scroll (meridionally divided) and double-scroll (circumferentially divided) turbines revealed their very distinct efficiency characteristic [34]. A double-scroll turbine has been shown to deliver higher peak efficiency at full admission conditions. Therefore, a twin-scroll turbine showed lesser deterioration at partial admission conditions because the flow was still capable of expanding into the larger rotor inducer area, even though not entirely [35,36]. Chiong et al. [37,38] presented a revised one-dimensional pulse flow modeling of twin-scroll turbocharger turbine under pulse flow operating conditions. The results showed that a twin-scroll turbine does not operate at full admission throughout the in-phase pulse flow conditions. Instead, the turbine worked at an unequal admission state due to the magnitude disparity of the turbine inlet flow. Rajoo et al. [39] discussed the details of unsteady experimentation and analysis of a twin-scroll variable geometry turbine for an automotive turbocharger. The cycle-averaged efficiency of the twin or single-scroll nozzled turbine was found to depart significantly from the equivalent quasi-steady. In comparison to the nozzleless single-scroll turbine, the departure was as much as 32%. When STST is used to drive EGR, both two-exhaust

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Applied Energy 223 (2018) 263?272

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

The ATST is a potential turbocharging technology that can effectively solve the problem in the STST [41]. The ATST first emerged in the last century, which was used to keep the overall fuel consumption increase as low as possible on the 6-cylinder truck engines equipped with an EGR system by Daimler-Benz. It has two scrolls with different throat areas, and the two scrolls are separately connected with the two groups of cylinders. In the engine, one group of cylinders linked with the small scroll is operated with a high exhaust gas backpressure, and the other is operated with a fuel-saving low exhaust gas backpressure [42]. Therefore, the small scroll can offer a higher backpressure to drive EGR. On the other hand, the large scroll is isolated from the EGR passage and can reduce the average backpressure for better fuel economy. Daimler Trucks launched a generation of new Mercedes-Benz diesel engines for medium-duty [43,44] and heavy-duty [45,46] commercial vehicles with the institution of the Euro 6 emission standard. ATST has been developed as the core of the EGR system for the reduction of NOx emissions and applied the system to the modern generation of Daimler commercial vehicle engines [47]. In 2007, a heavy-duty diesel engine with 14.8 L displacement called OM 472 was introduced into the market. After that, Daimler continued launching new products, including 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 6 emissions limits, has lower NOx emissions and fuel economy advantage of 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 overboost pressure. However, when the EGR valve is not completely open at the medium and high-speed range of the engine, the backpressure of the small scroll will increase to cause adverse fuel economy because of more pumping losses.

This paper first presented a new ATST-2WG, which has two WGs called WG1 and WG2. The two WGs are connected with the small scroll and the large scroll, respectively. When the backpressure of the small scroll is so high that it causes too high an EGR rate, the ATST-2WG can open the WG1 to decrease the backpressure for better fuel economy to ensure engine power. Compared with VGT, two-stage turbochargers, STST and ATST-1WG, the ATST-2WG are relatively simple and can effectively solve the problem that fuel economy decreases at partial EGR valve opening degrees. This paper consists of three parts. First, an engine experiment with ATST is presented, and the simulation models are established. Second, both the wastegates control strategy and the critical parameter ASY impact laws are explored. Finally, the advantages of ATST-2WG are evaluated in comparison with ATST-1WG.

Details of the engine specifications are given in Table 1. The speed and load of the engine are regulated using a C 500 eddy current dynamometer with an accuracy of ? 10 rpm and ? 1.25 Nm, and the maximum speed is 4500 rpm. The maximum brake torque is 4000 Nm (1000?1700 rpm) and the rated power is 720 kW (1700?4500 rpm). An AVL735S fuel consumption measuring instrument with a precision of 0.12% of the recorded value is used to meter the engine fuel consumption, and it has a measuring range of 0.1?110 kg/h. The data acquisition system is a PUMA OPEN 1.2 all-in-one bench-top instrument, and the data is processed by CONCERTO-P software. The engine intake mass flow is collected by a Sensyflow P/4000 air flow meter (accuracy: ? 5 mg), and the temperature of the intake gas boosted is controlled within 30?55 ?C by a BATCON device. The engine fuel and coolant have constant temperature regulated by an AVL753 fuel constant temperature control device and an AVL553 coolant constant temperature control device. The fuel and coolant temperature are controlled within the range of 15?80 ?C and 70?120 ?C, respectively. The experiment engine has an intercooled EGR system whose EGR rates can be sampled in real time by a gas emission analyzer MEXA7100DEGR. Moreover, the environmental conditions are accurately controlled by the engine intake and exhaust environment simulation system ACS2400. The environmental temperature and pressure are set to 298 ? 1 K and 100 ? 1 kPa, respectively. The diesel engine operates at full load and different speeds.

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

2. Experiment and simulation methods

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

ASY = 1 ? 100%

2

(1)

In which 1 and 2 are the throat areas of the two scrolls. ASY can quantitatively evaluate the difference between the two scrolls.

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

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

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D. Zhu, X. Zheng

Table 1 Test engine specifications.

Items

Engine type Number of valves per

cylinder Bore Stroke Displacement Compression ratio Cooling system Air intake system

EGR system Rated power Maximum torque

Value and unit

Inline 6-cylinder DI diesel 4 (2 inlet / 2 exhaust)

129 mm 160 mm 12.55 L 18.2:1 Water cooled Intercooled asymmetric twin-scroll turbocharger (ASY = 53%) Intercooled EGR 351 kW (1900 rpm) 2380 Nm (1000?1400 rpm)

Combustion rate [%]

Applied Energy 223 (2018) 263?272

0.07

RES

0.06

42%

53%

0.05

58%

68%

0.04

84%

100% 0.03

0.02

0.01

0.00 120

Crank angle [deg]

Fig. 3. Experimental combustion rate data.

(a)

(b)

Fig. 2. Turbocharger maps: (a) compressor; (b) ATST (ASY = 53%). to experimental pipes materials. An orifice joins the two entry pipes at the turbine inlet, which can model backflow [54]. In the turbocharger model, the compressor and ATST (ASY = 53%) maps are obtained from turbocharger experiments on the gas test stand during the entire development phase, as shown in Fig. 2. The parameter "Wastegate Area Fraction 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 engine combustion model, the experimental combustion rate is adopted. It is defined as the fuel combustion quality with unit time and can be measured by an AVL641 combustion analyzer. At full load, the experimental combustion data is as shown in Fig. 3. The relative engine speed (RES) is normalized to the maximum engine speed at the full load

(1900 rpm).

Engine speed

RES =

? 100%

1900rpm

(2)

The results of the single-zone and two-zone combustion models are shown below. It can be seen that two combustion models have the same trend. Meanwhile, the Woschni Formula is chosen reasonably in the cylinder heat transfer model.

0.8

g

=

820p0.8T-0.53D-0.2 C1Cm

+

C2

Ta pa

Vs Va

(p-p0

)

(3)

where g is the instantaneous average heat transfer coefficient for the working gas and chamber wall; p and T are the working gas pressure and temperature, respectively, D is the cylinder diameter, Cm is the average piston speed, C1 is the gas velocity coefficient, pa, Ta and Va are working gas pressure, temperature and cylinder volume at the beginning of compression, respectively, andp0is cylinder pressure of the inverted engine. Therefore, the heat exchange for the gas and wall (Qw) are calculated as Eq. (4):

dQw d

3

=

1

dQwi d

=

1

3

1

g?Ai (T-Twi)

(4)

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

The engine model calculation is finished, and the experimental and simulation results including torque, power, brake specific fuel consumption (BSFC) and EGR rate are compared, as shown in Fig. 4. The simulation results and the experiment results at full load have good agreement, 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 engine system schematic is shown in Fig. 5(a). This ATST has two WGs, which are connected with the small scroll and the large scroll, respectively. Based on the engine simulation model with an ATST-1WG, the engine model equipped with an ATST-2WG is established in Fig. 5(b). The engine system, EGR system and compressor are unchanged, and only the turbine structure is different compared to Fig. 1(b). The "Wastegate Area Fraction for Entry 1" parameter is also set as 0 in the turbine model, which means this WG is only linked with the large scroll. WG1 connects the small scroll to the atmosphere and is as shown in Fig. 5(a).

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

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

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D. Zhu, X. Zheng

(a)

Applied Energy 223 (2018) 263?272

EGR Cooler

EGR valve

p2

p1

EGR linked with the small scroll

Charge Air

Cooler

T

C

WG1

WG2

(a)

Engine system

WG linked with the small scroll

EGR system

(b)

ATST-2WG

(b)

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

(c)

(d)

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

change the EGR rates. This section mainly discusses the EGR valve effects of ATST-1WG on engine performance. In the model, the EGR valve is 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 is defined as the ratio of the hole diameter and the maximum diameter in the experiment. Based on the model in Fig. 1(b), the remaining ASY is 53%, and by changing OPD at a range of 0?100% at different engine speeds, 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 nondimensionalized by dividing the EGR rate and BSFC at the point when both the OPD and RES are 100%. It can be seen that the EGR rates continue with increasing OPD, and the increasing trend becomes larger with increasing engine speed. On one hand, the EGR rate first increases and then essentially remains unchanged, and BSFC has a small upward trend as the OPD increases at low engine speeds. It is well known that the exhaust gas is not usually enough for engines to boost inlet gas; therefore, the turbine WG is usually closed at low engine speeds. In this diesel engine, the WG, which is connected with the large volute, is completely closed under RES 58% (engine maximum torque point). Given a smaller OPD, there is more boosting exhaust gas; therefore, the turbine power will rise. The engine intake pressure increases resulting in better power and BSFC. When the OPD is beyond 60%, the EGR driving pressure comes to a maximum value, so the EGR rate remains unchanged. On the other hand, the WG is gradually open as the engine speed increases at high speeds to avoid overboost pressure. The exhaust backpressure is enough to drive EGR, and it is higher with increasing speed; thus, the EGR rate continually increases. When the OPD reaches

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Applied Energy 223 (2018) 263?272

(a)

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

(b)

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

(c)

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

100%, the EGR rate will remain unchanged. Meanwhile, the backpressure of the small scroll decreases with the increasing OPD, resulting in smaller pumping losses and better fuel economy. At 100% RES, the BSFC 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 speeds in the engine cycle simulation. PMEP increases when OPD increases because 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 exhaust passages have higher pressures compared to the intake pressure in Fig. 7. The differences are smaller when OPD increases. It was mentioned before that the EGR circuit is only linked with the small volute.

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

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

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

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Applied Energy 223 (2018) 263?272

(a)

(a)

(b)

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

4.1. Wastegates control strategy

(b)

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

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

Given the same EGR rates of the engine with the ATST-1WG, the WG1 and WG2 relative opening degrees in the engine with the ATST2WG are taken from Fig. 9(a) and (b), respectively. At 100% OPD, the WG1 is fully closed, and the EGR rate maximizes. Beyond the speed at the maximum torque point, the WG1 opening degree can regulate the small scroll pressure, and its increasing can decrease the EGR rate. To ensure the boost-pressure, the WG2 needs to be closed little by little as more 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, which characterizes the interrelationships of the two scrolls and impacts the

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

balance between the engine emissions and fuel consumption. This section will investigate the ASY effects on engine performances comparing the ATST-1WG and ATST-2WG. First, based on the engine models with an ATST-1WG and an ATST-2WG discussed in Section 2, the engine cycle simulations are conducted using the same engine system, EGR system and compressor. Given the same EGR rate at the same OPD and 100% RES for ATST-1WG and ATST-2WG, all models operate at the full load engine condition, and the ASY values are from 30% to 80%.

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

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Applied Energy 223 (2018) 263?272

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

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

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

(a)

4.3. Fuel economy improvements in the ATST-2WG

(b)

Fig. 13. Pumping loop diagrams of ATST-1WG and ATST-2WG at 50% ASY and 70% OPD: (a) cylinder with large scroll; (b) cylinder with small scroll.

2WG with respect to the engine with an ATST-1WG are shown in Fig. 10. Under the same conditions, the engines with an ATST-2WG or ATST-1WG have the same power and BSFC at 100% OPD from 30% to 80% ASY. At 100% OPD, the WG1 is completely closed to offer the maximum EGR rate. Therefore, the WG1 has no influence on engine performance. However, at partial OPD conditions, the dynamic performance and engine fuel economy with an ATST-2WG are clearly better than those of the engine with an ATST-1WG. The power has improvements of 2.76% and 2.69% for BSFC at 60% OPD. With increasing ASY, the power and BSFC improvements decrease since a higher ASY usually means a larger small scroll throat area, thus

In the sections above, it is clearly demonstrated that the ATST-2WG has greater potential than ATST-1WG on engine performances. Given 53% ASY and the same EGR rates, the maximum fuel economy improvements in the ATST-2WG over the ATST-1WG at different engine speeds are shown in Fig. 14. All engine models operate at full load, and the engines with the ATST-1WG have 60% OPD. The BSFC change rate first remains unchanged and then decreases as engine speed increases. The maximum fuel economy improvement is 2.91% at 100% RES (the rated power point). When the engine speed is beyond 58% RES (the maximum torque point), its growth of 10% results in a decrease in the BSFC rate by approximately 0.7%. For the diesel engines with an ATST1WG, the WG2 is completely closed at low engine speeds and gradually opens over the maximum torque point speed. Therefore, for engines with an ATST-2WG, the WG1 is also completely closed to ensure engine power leading to no BSFC advantage under 58% RES. With the increasing engine speed, the WG in ATST-1WG is usually open to avoid overboost pressure, and the 60% OPD will increase the backpressure of the small volute, which causes the fuel economy to deteriorate. If the EGR rate demand is lower, the OPD will be smaller and the fuel economy will further deteriorate. By contrast, in the ATST-2WG, the EGR valve is completely open, and the WG1 is properly open to achieve the same EGR rates as the ATST-1WG. Therefore, the exhaust backpressures of the two scrolls will decrease for better charge air conditions, especially at high engine speeds.

5. Conclusions and remarks

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

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