Experimental response surface study of the effects of low-pressure exhaust gas recirculation mixing on turbocharger compressor performance

Applied Energy 261 (2020) 114349

Contents lists available at ScienceDirect

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

Experimental response surface study of the effects of low-pressure exhaust gas recirculation mixing on turbocharger compressor performance

T

⁎

Amin Reihania,b, , John Hoarda, Stefan Klinkertc, Chih-Kuang Kuanc, Daniel Stylesc, Greg McConvillec a

Department of Mechanical Engineering, University of Michigan, Ann Arbor, MI, United States Electrical Engineering and Computer Science Department, University of Michigan, Ann Arbor, MI, United States c Ford Research and Innovation Center, Dearborn, MI, United States b

H I GH L IG H T S

performance is significantly affected by LP-EGR mixing flow field. • Compressor isentropic efficiency drops (up to 25% rel.) at high EGR momentum ratio. • Compressor perturbs the local inlet incidence angle reducing compressor efficiency. • Vorticity mixing uniformity does not show significant effect on compressor performance. • LP-EGR • Eccentric EGR generates a pre-whirl acting similar to a compressor inlet guide vane.

A R T I C LE I N FO

A B S T R A C T

Keywords: Energy efficiency Mixing flow field Diesel engine Low-pressure EGR Turbocharger compressor

In Low-Pressure Exhaust Gas Recirculation (LP-EGR), clean exhaust gas is extracted downstream of the after treatment, and reintroduced upstream of the turbocharger compressor. A major pathway for engine fuel economy improvement, by employing LP-EGR, is the enhancement of compressor and turbine efficiencies by increased flow rates which moves the operating points towards higher efficiencies. However, what is often overlooked in the literature is the influence of LP-EGR/air mixing flow field on the compressor performance. Here, we systematically study this effect on a turbocharger unit for a diesel engine on a hot gas stand using response surface methodology. In addition, the mixing flow field of LP-EGR and air upstream of the compressor was scanned using a 3-dimensional directional probe. A reconfigurable T-junction mixer geometry was used, enabling the study of major mixing parameters such as mixing length, and EGR introduction angle. The mixing flow field showed a strong dependence on EGR-to-air momentum ratio, significantly affecting the EGR uniformity and axial velocity uniformity. Furthermore, multiscale stream-wise vortices were generated in the mixing section, with an intensity that increased with momentum ratio, and decreased with mixing length. In addition, by eccentric LP-EGR introduction a bulk swirl was generated in the mixing section similar to the flow field obtained using an intake guide vane. A decrease in compressor efficiency with LP-EGR introduction, compared to baseline compressor map, was observed. The efficiency degradation was larger at higher EGR momentum ratios and flow rates. In the range of mixing lengths limited by engine packaging, up to length-to-diameter ratio (L/D) = 2.5, the flow perturbations were not damped, and a decline in both compressor pressure ratio and efficiency was observed. Efficiency degradation mechanism is found to be the formation of strong vortices in the mixing zone upstream of the compressor, which are advected to the compressor impeller inlet and perturb the local incidence angle. In contrast, the EGR and axial velocity non-uniformities did not show a negative impact on the compressor. This study identifies the major flow parameters that cause significant degradation of compressor performance, and proposes a figure of merit for design of efficient LP-EGR mixers to benefit from the fuel economy advantages of LP-EGR architecture in diesel and gasoline direct injection engines.

⁎

Corresponding author at: Department of Mechanical Engineering, University of Michigan, Ann Arbor, MI, United States. E-mail address: [email protected] (A. Reihani).

https://doi.org/10.1016/j.apenergy.2019.114349 Received 1 September 2019; Received in revised form 5 December 2019; Accepted 9 December 2019 0306-2619/ © 2019 Elsevier Ltd. All rights reserved.

Applied Energy 261 (2020) 114349

A. Reihani, et al.

Nomenclature

ṁ μ N → n ω p01 p02 Pref Tref Q r RANS Re ρ T01 T02 T02, isen

Symbol α¯

Description Mean angular perturbation of the velocity vector from axial direction AIS air induction system c molar concentration PR pressure ratio D air duct diameter at compressor inlet DEGR LP-EGR introduction port diameter Dim impeller diameter DL-EGR dual-loop exhaust gas recirculation EGR exhaust gas recirculation ηt-t isentropic total-to-total compressor efficiency Γ velocity uniformity index γ ratio of specific heats HP-EGR high-pressure exhaust gas recirculation Lmix mixing length of LP-EGR and air upstream of the compressor LP-EGR low-pressure exhaust gas recirculation

θEGR UI v ω¯ z ωz, rel

1. Introduction

mass flow rate of compressor dynamic viscosity compressor shaft angular velocity normal vector to flow path cross section angular velocity total pressure at inlet/outlet of the compressor reference pressure (1 atm)/temperature (298.15 K) volumetric flow rate of compressor EGR-to-air momentum ratio Reynolds-Averaged Navier-Stokes Reynolds number density total temperature at inlet/outlet of the compressor total temperature at the outlet of compressor in a isentropic compression process LP-EGR introduction angle (introduction eccentricity) EGR mixing uniformity index flow velocity mean axial flow vorticity relative axial flow vorticity

specific heats, and (III) altering chemical composition which increasing the endothermic dissociation of H2O and CO2 molecules. All the aforementioned effects result in a lower flame temperature, which in turn leads to lower NOx emissions and higher hydrocarbon (HC), soot, and CO formation [9]. Diesel engines can operate at very high levels of EGR (more than 50% EGR fraction) to suppress NOx emissions; however, the EGR fraction is limited in gasoline engines to 20–25% to prevent combustion instability [10]. The two main EGR architectures are the low-pressure (LP-EGR) and high-pressure EGR (HP-EGR) [7]. A system combining both HP- and LPEGR architectures is referred to as dual loop EGR (DL-EGR), which is shown schematically in Fig. 1. The HP-EGR is the commonly used configuration, where the exhaust is extracted from the high-pressure side of the turbine (turbine inlet or exhaust manifold), commonly passed through an EGR cooler, and introduced downstream of the compressor in the intake manifold. The maximum flow rate of HP-EGR is limited at high engine loads due to elevated intake manifold pressure and opposite pressure differential between the exhaust and intake manifold. In this condition, partially closing an intake throttle valve and/or variable geometry turbine (VGT) are often employed to provide sufficient pressure differential to achieve the target EGR flow and control EGR flow rate. However, both of these approaches restrict the intake air or exhaust gases and result in increased pumping losses [11].

The automotive industry has faced increasing concerns regarding vehicular greenhouse gas emission and fossil fuel consumption. Future fuel economy standards, such as U.S. Corporate Average Fuel Economy (CAFE) which requires achieving fleet-wide fuel economy of 48.7 mpg by 2025 [1], and European regulations targeting 15% and 30% reductions in fleet average CO2 emissions from new light-duty vehicles in 2025 and 2030 respectively [2]. Both standards require significant improvement in powertrain efficiency of passenger and light duty vehicles. One of the methods of achieving this goal is the use of lean burn gasoline and diesel engines for their inherently higher fuel conversion efficiency [3,4,5]. Simultaneously, stringent regulations on NOx emissions, such as U.S. Tier 3, as well as more realistic emission testing drive cycles such as European RDE, impose a challenge for the use of lean burn engines and require higher levels of charge dilution and low temperature combustion [6]. The prevalent approach for charge dilution of both stoichiometric and lean burn engines and inhibition of NOx formation in lean burn engines is the use of exhaust gas recirculation (EGR) [7,8]. EGR has the following effects on combustion process: (I) charge dilution by means of intake oxygen displacement, (II) altering the thermodynamic properties of charge by increasing the average charge heat capacity and ratio of

Fig. 1. Schematic of dual-loop EGR architecture including HP and LP-EGR loops, colors indicate the temperature of the flow. 2

Applied Energy 261 (2020) 114349

A. Reihani, et al.

isentropic efficiency degradation at low EGR fractions which increases to 5% at higher EGR rates. Also, they indicated that achieving high mixing uniformity generally leads to higher mixer pressure drop which can outweigh the losses due to disturbed compressor inlet flow if the entire compressor plus mixer efficiency is considered. It should be noted that, Bruneh et al. used work averaged total pressure, and massweighted temperature as the fluid properties at compressor inlet which were calculated using air and EGR properties before entering the mixer. Although, the efficiency calculated using these averaged properties enables evaluation of the entire compressor plus mixer system, it is inaccurate for determination of the compressor efficiency as a standalone unit. Forlese and Spoleti [28] numerically studied the impact of coherent flow motion at compressor inlet that may occur in an automotive engine induction system, for instance as a result of air intake valve or LPEGR introduction. They evaluated the performance of compressor by imposing either bulk swirl or twin swirl at compressor inlet at two intensity levels using a 3D Computational Fluid Dynamics simulation under steady-state assumption with a RANS-based turbulent model. It was reported that the twin swirl leads to a mass-flow-rate/pressureratio map very close to the baseline but with a loss of efficiency growing with the increasing the swirl intensity. In addition, the bulk co-rotating swirl (with respect to compressor impeller) improved the efficiency but decreasing the mass flow rate of the compressor. The opposite trend of increased mass flow rate and pressure-ratio but decreased efficiency was reported in the case of bulk counter-rotating swirl. Based on the available literature, the following questions need to be addressed:

A second drawback of the HP-EGR configuration is the cylinder-to-cylinder EGR maldistribution due to limited mixing length of EGR with intake air in the intake manifold [12]. In the LP-EGR configuration, exhaust gas is recirculated from downstream of the turbine, and often downstream of aftertreatment (A/ T) system, and fed back upstream of the compressor. This typically requires a longer return tube, which includes an EGR valve and can include an EGR cooler. This configuration has some advantages compared to the HP-EGR configuration as listed below. (a) Higher EGR rates at high engine load/speeds are achievable due to low pressure point of introduction at compressor inlet [13]. (b) Less EGR cooling requirement due to lower post-turbine temperatures, as well as additional cooling in the long EGR return tube. This can lead to lower intake manifold temperature, which is desired for improved thermal and volumetric efficiencies [14]. Lower intake temperature can also lead to improved combustion phasing and reduced in-cylinder heat losses, as well as lower NOx emissions [13,15]. (c) More uniform cylinder-to-cylinder EGR distribution due to longer mixing length of EGR and air [16]. (d) Less exposure of EGR loop to fouling. (e) Improved compressor and turbine performance, as a result of increased turbine and compressor flow rates [17,18]. The effect on compressor performance is specifically important since it has significant influence on the overall engine performance and fuel conversion efficiency of the engine. On the turbine side, increased flow rate can lead to a more open VGT state while maintaining the target boost pressure, which results in lower backpressure and a reduction in pumping losses [19,13,20]. On the other hand, the LP-EGR configuration has the following potential concerns: (a) slower airpath transient response to changes in EGR rate due to larger volume of the EGR path, which includes the longer return tube, compressor, and the charge air cooler. Slower response makes EGR rate control and EGR level estimations challenging [21]. (b) Water condensation in the LP-EGR return tube, LP-EGR cooler, or downstream of the compressor in the charge air cooler due to lower gas temperature in the LP-EGR path and the water vapor content in diesel exhaust [22,23,24]. Should water condense in the system, water droplets or a wall film could risk damage to the compressor wheel [25]. (c) Exposure of the LP-EGR loop, compressor, and engine intake to small particles eroded from the catalysts. To discuss the effect of LP-EGR on compressor performance, consider the compressor operating points on the FTP-75 drive cycle for a medium-duty diesel engine, shown in Fig. 2 which is obtained using a validated 1D gas-dynamics and thermodynamics model [26]. The HP and LP-EGR data points on Fig. 2 are obtained under the same EGR fraction, while the rest of operating parameters of the engine are kept constant, in order to demonstrate the component-wise impact of the LPEGR on compressor performance. As can be seen, most of the operating points for the HP-EGR configuration are concentrated towards the lower flow rates; since the compressor is sized to cover the entire speed/load map of the engine. Therefore, in typical driving conditions by increasing the LP-EGR rate, which increases the compressor flow rate, the compressor operating point can be shifted toward higher efficiency islands, as indicated by blue operating points representing the LP-EGR configuration in Fig. 2. In the figure, the dashed lines are lines of constant compressor efficiency, with the center island enclosed by a dashed line having the highest efficiency. The analysis above assumes that the perturbation of the flow field by mixing of LP-EGR and air upstream of the turbocharger compressor does not affect the compressor performance (i.e. the compressor map does not vary by introducing LP-EGR); however, as will be discussed this is not the case. A limited number of research studies have investigated the effect of LP-EGR mixing on compressor performance. Bruneh et al. [27] experimentally evaluated the mixing flow field of LPEGR using three candidate mixer geometries including a Lobed mixer, a radial mixer, and a perpendicular (T-junction) mixer. They identified that the Lobbed mixer and the T-junction resulted in the highest and lowest mixing uniformity respectively. In addition, they report 3%

• To what extent does the LP-EGR mixing flow field influence the • •

turbocharger compressor performance? This is a challenging measurement given that the fluid properties at compressor inlet is unknown. What flow structures, and which operating conditions cause degradation/improvement in compressor performance? This requires first a mixer setup where different flow field condition can be generated. Secondly, a minimally invasive, in-situ measurement of mixing flow field close to compressor inlet. What is a suitable figure of merit for LP-EGR mixer design in order to maximize the system efficiency of compressor and mixer?

In order to address these questions, in this study we use a T-junction LP-EGR mixer with configurable geometry to systematically vary the mixing flow field and observe its effect on compressor performance. The flow field upstream of the compressor was scanned using a directional probe with an in-house fabricated positioning mechanism with redundant degrees of freedom, enabling scanning the flow field with

Fig. 2. Compressor efficiency and operating points during FTP-75 drive-cycle for HP-EGR and LP-EGR configurations delivering equal EGR fraction on a medium-duty diesel engine. 3

Applied Energy 261 (2020) 114349

A. Reihani, et al.

section and the turbocharger unit, which was constrained by space and would have impeded the reconfiguration of the mixer parameters. In addition, the 1 DOF redundancy in the mechanism allowed for positioning the rod of the probe as close as possible to the duct walls to minimize flow perturbation during measurement. This probe can also be used for measurement of flow field on engine dynamometers, where the packaging of turbocharger and LP-EGR mixer do not provide access to the mixing section from the sides.

minimal perturbation. Importantly, calculation of the compressor performance metrics was done by averaging the scanned flow properties at compressor inlet rather than using flow properties upstream of the mixing point. This is an important issue in calculation of efficiency, as using properties upstream of mixing point would lead to over-estimation of mass-averaged temperature at compressor inlet (due to heat transfer in mixing section) which leads to significant over-prediction of the compressor efficiency. Finally the correlation between flow field and compressor response was used to identify the flow structures and flow field properties that have significant effect on compressor performance.

3. LP-EGR mixer geometric configuration The geometry of the T-junction mixer is shown in Fig. 4, with an air path diameter Dair = 9.7 cm and EGR port diameter of DEGR = 2.9 cm. The geometric mixer parameters studied here are mixing length, Lmix, EGR introduction eccentricity (or introduction angle), θEGR. The prototype mixer was designed as a 5-piece part, where the combination of these parts could form discrete levels of mixing length, and introduction angles. The discrete levels of geometric parameters, as well as the criteria for selecting their range are listed in Table 1.

2. Experimental setup The experimental setup is schematically shown in Fig. 3, which consists of a turbocharger unit installed on a hot gas test stand. As a case study, we have considered a T-junction as the LP-EGR mixer, which was selected for its low implementation cost, ease of prototyping and reconfiguration, and compact packaging. Heated air was used to simulate EGR flow, upstream of the compressor, which was mixed with air at room temperature and flowed to the compressor inlet. The static and total properties of air upstream of the LP-EGR mixer, at the EGR inlet, compressor outlet, turbine inlet, and turbine outlet were recorded, as well as the air, EGR, and turbine mass flow rates. The mixture flow field was also measured at different cross sections upstream of the compressor as will be discussed in the next section.

3.1. Design of experiment The EGR/air mixing flow field is mainly governed by two dimensionless numbers: (1) Air Reynolds number, Reair (Eq. (1)), representing the airflow rate and main stream turbulence intensity, and (2) EGR-to-air momentum ratio, r (Eq. (2)), representing the ratio of momentum flux of the EGR stream with respect to the air stream as defined in [29]. Note that in the Jet-in-cross flow literature, the momentum ratio is also defined without the square root as ρEGR VEGR2/ρair Vair2 [30].

2.1. Mixing flow field measurement The LP-EGR/air mixing flow field was scanned at different cross sections of the mixer upstream of the compressor inlet for two main reasons; (a) calculation of compressor performance metrics, such as isentropic efficiency, requiring the mass-flow-averaged inlet flow properties which cannot be found from bulk flow measurements due to high flow non-uniformity. (b) Obtaining the correlation between flow structures at compressor inlet and variation in compressor performance. The flow field was measured using a custom 5-hole probe (U.S. Sensors Corp.) with an integrated tip thermocouple enabling the measurement of three velocity components and temperature at the tip location. A 90° bend at a distance of 50 mm from the probe tip allowed the introduction of the probe to the mixing section from further upstream as shown in Fig. 4. The probe tip was then positioned in the mixing section to the desired (x, y, z) coordinate using an in-house fabricated delta-robot slider mechanism with four degrees of freedom (DOF). This mechanism enabled the positioning of the tip with 0.25 mm accuracy without the need for adding actuators close to the mixing

ReAir =

ρV ¯ ¯D μ¯

(1) 1

2 2

ρ VEGR r = ⎜⎛ EGR 2 ⎟⎞ ⎝ ρair Vair ⎠

(2)

A partial factorial design of experiment with a total number of 41 EGR introduction experiments based on the Box Behnken method was used [31] as listed in Table 2. The highest momentum ratio tested here, r = 2.66, corresponds to a LP-EGR fraction of 18% which is relevant to both diesel and gasoline direct injection (GDI) engine application. In addition, in Table 2 the total compressor flow rate is specified by the normalized flow defined as (ṁ − ṁ surge ) (ṁ choke − ṁ surge ) at a given compressor speed line. At the bottom of this table, a set of baseline experiments is listed which includes testing points at the same

Fig. 3. Schematic of the hot gas stand for turbocharger testing with LP-EGR introduction. 4

Applied Energy 261 (2020) 114349

A. Reihani, et al.

Fig. 4. Schematic of the T-junction LP-EGR mixer, showing the coordinate system and the geometric parameters. The mixing flow field measurement probe is shown in red.

Table 1 Geometric parameters of the T-junction EGR mixer used in the experimental study. Parameter

Discrete values

Range criteria (min/ max values)

Mixing length, Lmix Introduction angle, θEGR EGR-to-air momentum ratio, r

8, 16, 24 cm 0, 60° 0.45, 1.20, 2.66

~/packaging constraint ~/geometric constraint for maintaining EGR introduction area Pressure loss in EGR delivery tube/ geometric constraints

compressor total flow rate and speeds of the LP-EGR introduction experiments, where no EGR was introduced. These points were measured on the exact same experimental setup as the LP-EGR experiments (which includes the mixer), and will serve as the baseline for evaluation of the effect of EGR introduction. The compressor total mass flow rate, LP-EGR fraction, LP-EGR total temperature, and turbocharger shaft speed were fixed at each experimental point. Also, the turbine inlet temperature was fixed at 600 °C according to SAE standard for all measurements [32]. The response parameters of interest were primarily, compressor total-to-total efficiency and pressure ratio, as well as turbine inlet pressure, mass flow rate, and power.

gas stand employs temperature and pressure measurement stations at inlet/outlet of the compressor to obtain the thermodynamic state of the gas at these locations. Together with the mass flow measurement upstream of the compressor, the compressor isentropic efficiency and pressure ratio can be calculated. In the LP-EGR introduction experiments, the temperature, pressure, and mass flow measurements were also added on the EGR introduction port, as shown in Fig. 4. However, the value of Tmix cannot be estimated using a first law-analysis of the control volume enclosing the mixing section, since the assumption of adiabaticity and negligible viscous dissipation are not valid in the mixing section. Those assumptions would cause over-prediction of the average compressor inlet temperature and would lead to artifact of unrealistically high compressor efficiencies. Therefore, the mass-flow-averaged inlet temperature is calculated directly from probe measurements at the compressor inlet as shown in Eq. (7).

4. Results and discussion In this section, the gas stand measurements of the effect of LP-EGR on compressor performance will be presented, then they will be explained by analysis of the mixing flow field obtained from the scanning probe.

→

T¯inlet =

n dA ∫ T (x , y ) ρV . → (7)

ṁ c

4.1. Measurement of compressor performance 4.2. Regression analysis

The main metrics of compressor performance are the pressure ratio, PR (Eq. (3)), and isentropic total-to-total efficiency with ideal gas approximation, ηt-t (Eq. (4)). These metrics are functions of reduced-mass flow rate, ṁ cor (Eq. (5)), and reduced compressor speed, Nred (Eq. (6)) [19].

PR =

p02 p01

(3)

ηt−t =

Δh 0s T − T01 (T02 T01) − 1 ≈ 02 = Δh 0 T02s − T01 (p02 p01 )γ − 1 γ − 1

(4)

ṁ cor =

Nred =

In order to obtain the effect of each mixer and flow parameter on the turbocharger performance, a large set of statistical models was fitted to the data using least squares regression. As shown in Eq. (8), the function space of the fit includes up to fourth order polynomial to model the main effects of independent parameter (first term) and up to second order for interaction effects (second term) where M is the number of predictor variables, in this case M = 5 corresponding to variables listed in Table 2.

y=

ṁ T01 Tref P01 Pref

4

i=1

k=0

ai, k x ik +

+ ei, j x i2 x j2

⎞ ⎟ ⎠

⎝

(5)

N T01 Tref

M

∑ ⎛⎜ ∑

(6)

∑

bi, j x i x j + ci, j x i x j2 + di, j x i2 x j

j ∊ {0,1, ⋯ , M },(j ≠ i)

(8)

Eq. (9) shows the error function. The first term is a weighted squared error with the weights (Wi) distributed proportional to the deviation of the compressor metric from the baseline value, with a

In a typical experiment to obtain the turbocharger performance, without any mixing flow at the inlet of the compressor, turbocharger 5

Applied Energy 261 (2020) 114349

A. Reihani, et al.

Table 2 Design of experiment points for LP-EGR introduction on turbocharger gas stand. Test number

Input parameters N (krpm)

Normalized flow (%)

Momentum ratio, r

Mixing L (cm)

Eccentricity (deg)

LP-EGR Experiments

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41

15 15 70 70 70 40 40 40 40 15 70 70 40 40 40 40 40 15 15 70 70 70 70 40 40 40 40 40 40 40 40 40 40 40 40 40 40 40 40 15 70

10 80 10 80 80 40 40 40 40 40 40 40 10 10 80 80 80 40 40 40 40 40 40 10 10 80 80 40 40 10 80 40 40 40 40 40 40 40 40 40 40

1.20 1.20 1.20 1.20 1.20 0.45 0.45 2.66 2.66 0.45 0.45 2.66 0.45 2.66 0.45 2.66 2.66 1.20 1.20 1.20 1.20 1.20 1.20 1.20 1.20 1.20 1.20 1.20 1.20 1.20 1.20 0.4 1.20 2.66 0.45 1.20 2.66 0.45 2.66 1.20 1.20

14 14 14 14 14 8 24 8 24 14 14 14 14 14 14 14 14 8 24 8 8 8 24 8 24 8 24 14 14 14 14 8 8 8 24 24 24 14 14 14 14

0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 60 60 60 60 60 60 60 60 60 60 60 60

Baseline Experiments

1 2 3 4 5 6 7 8 9

15 15 15 40 40 40 70 70 70

10 40 80 10 40 80 10 40 80

0 0 0 0 0 0 0 0 0

14 14 14 14 14 14 14 14 14

0 0 0 0 0 0 0 0 0

constant c = 0.5 here, in order to put more emphasis on determining the effect of operating points with the largest influence on compressor performance. The second term in the error function is the sum of all orders (km, n) with a properly chosen constant, h, which is a regularization term to force the algorithm to obtain the simplest model. For instance, if the fit function is second order with respect to (w.r.t.) the mth variable, then km, m = 2, and if it contains an interaction term with second order w.r.t. the mth variable and first order w.r.t. nth variable then km, n = 3. Using this error function, a k-fold cross-validation was done and the simplest function which achieved a low testing error was selected. The purpose of cross-validation and imposing a penalty on higher order models was to avoid overfitting. It is noteworthy that, since the input parameters are in different numerical ranges, in the function fitting process, the predictor variables listed in Table 2 were first normalized by z-score scaling [33]. Fig. 5 shows the predicted vs true value of the experimental parameters using the fit described above.

e=

∑ Wi (yi − yi )̂ 2 + h ∑ km,n, Wi = 1 + c |yi | i

m, n

(9)

4.3. Response surface of compressor performance The results presented in this section are obtained from the best fitted response function (described in previous section) on gas stand measurements. Fig. 6 shows the response surface with error bars of relative change in compressor pressure ratio with respect to the baseline, i.e. (PR − PRbaseline)/PRbaseline, as a function of operating conditions and EGR mixing parameters. It can be seen that the eccentric EGR introduction with 60° introduction angle shows lower pressure ratio across the entire parameter space compared to concentric EGR. Also, Fig. 6 (a) and (b) show that the EGR introduction decreases the pressure ratio of the compressor as the compressor speed and total flow rate increases. In addition, the negative effect of EGR on pressure ratio 6

Applied Energy 261 (2020) 114349

A. Reihani, et al.

Fig. 5. Predicted value obtained from the best fitted model versus the experimental value of (a) relative change in compressor pressure ratio, (b) relative change in compressor isentropic efficiency ηt-t, and (c) relative change in turbine power.

(Pt − Pt.baseline)/Pt.baseline, at three levels of flow rate. As can be seen the turbine power closely reflects the opposite trends of compressor efficiency. This behavior is expected, since a decrease in compressor efficiency - at constant compressor flow rate and pressure ratio - leads to larger demanded shaft power from turbine.

increases at higher momentum ratios and slightly increases as the mixing length increases. Fig. 7 shows the response surface with error bars of relative change in compressor isentropic efficiency, ηt-t, with respect to the baseline, i.e. (η − ηbaseline)/ηbaseline, at three levels of flow rate. The effect of EGR introduction on compressor efficiency shows a parabolic dependence on compressor speed, indicating a larger decrease in compressor efficiency at lower compressor speeds and at low to mid flow rates. It can be seen that the compressor efficiency shows a strong negative dependence on EGR momentum ratio across all flow rates. The EGR introduction causes a larger drop in compressor efficiency as the mixing length increases within the maximum value limited by packaging. The EGR introduction angle does not demonstrate a strong effect on compressor efficiency as shown in Fig. 7(d). Fig. 8 shows the response surface with error bars of relative change in turbine power, Pt = ṁ t (hout − hin ) , with respect to baseline, i.e.

4.4. Mixing flow field The time-averaged velocity field was measured by scanning a grid of 73 points radially positioned as shown in Fig. 9 in the cross section of the EGR mixer at different distances downstream of the EGR port using a 3D directional 5-hole probe. Simultaneously, the temperature field was measured using a thermocouple embedded at the tip of the probe. The grid of points for temperature scanning were shifted due to the distance between the thermocouple junction and the pressure sensing holes on the probe (~9 mm). In the following we will present scanned

Fig. 6. Variation of experimental response surface showing relative change in compressor pressure ratio, as a function of operating condition and EGR mixing parameters. 7

Applied Energy 261 (2020) 114349

A. Reihani, et al.

Fig. 7. Variation of experimental response surface showing relative change in the compressor isentropic total-to-total efficiency, ηt-t, as a function of operating condition and EGR mixing parameters.

flow field of LP-EGR/air mixing upstream of a turbocharger compressor during operation.

mixing. Since the EGR inlet temperature was feedback controlled at 80 °C, significantly different from the air flow temperature at 25 °C, a measurable temperature field is formed as a result of EGR and air mixing. Given that the value of Lewis number is close to unity in such gaseous flows, the rates of thermal and mass transport are comparable;

4.4.1. Temperature and velocity fields The temperature field is particularly useful for the study of EGR

Fig. 8. Variation of experimental response surface showing relative change in turbocharger turbine power as a function of operating condition and EGR mixing parameters. 8

Applied Energy 261 (2020) 114349

A. Reihani, et al.

axial velocity is observed in the top right part of the cross section in Fig. 12(a) which is due to the blockage of the air path with EGR jet which has low axial momentum. The movement of this region from top to the top-right part of the cross section can be attributed to the clockwise impeller-induced swirl. At high momentum ratio of r = 2.66, Fig. 12(b) shows a significantly larger blockage of the flow in the top part of the cross section and a region of high axial velocity surrounding the blocked region. This high velocity region is formed by the incoming air stream which is diverted around the EGR jet. Notice that at high momentum ratios, the range of values of axial velocity is significantly larger than low momentum ratios, showing a variation of ~4 m/s corresponding to ~30% of mean flow velocity at z = 5 cm. As the flow is moved downstream, at z = 10 cm, the variation is reduced to less than 3 m/s, indicating axial momentum exchange during the mixing process which leads to lower velocity non-uniformity. Fig. 13 shows the temperature contour, as well as in-plane and outof-plane (axial) velocity field of eccentric LP-EGR introduction for θEGR = 60°, where EGR is introduced in the top right part of the cross section which is apparent from the region of high temperature in Fig. 13 (a). The in-plane velocity vector shown in Fig. 13 (b) at z = 5 cm, shows an eccentric downward velocity of EGR at the right part of the cross section. This eccentric jet leads to the clock-wise circulation (corotating with impeller) at z = 10 cm. This indicates that the eccentric EGR introduction indeed acts as an inlet guide vane and can generate either counter or co-rotating swirl depending on the direction of the EGR eccentricity.

Fig. 9. Grid of points in a cross section of the LP-EGR (such as the cross section A-A in Fig. 4) mixer for scanning of velocity field using a 3D directional probe.

therefore, the temperature field is a good proxy of EGR distribution. Fig. 10 shows the scanned mixing temperature field obtained in a concentric LP-EGR mixer with mixing length of 14 cm at two cross sections of 5 and 10 cm downstream of the introduction port. The measurements are done at a compressor speed of 40 krpm, at three levels of EGR momentum ratio of r = 0.45, 1.2, and 2.66. The observed temperature profile in Fig. 10 (a) indicates that at r = 0.45, a creeping EGR jet is formed which moves along the wall of the duct and is not separated downstream. At r = 1.2 (Fig. 10 (b)), an initial jet is formed moving towards the center of the air path. As the jet moves downstream some part of it is detached from the walls whereas the rest is moved along the wall. At r = 2.66 (Fig. 10 (c)) the EGR jet is partially detached from the top wall at z = 5 cm, and fully detached at z = 10 cm, showing a larger spread in the cross section. Fig. 11 shows the scanned 2D in-plane velocity vector corresponding to different EGR momentum ratio values, at the same cross sections and operating conditions of Fig. 10. At momentum ratio of r = 0.45, Fig. 12(a), at a plane of z = 5 cm downstream the LP-EGR port, the top region of the cross section shows a small downward velocity, which is related to the downward momentum of the EGR jet. Further downstream at z = 10 cm, this downward velocity of the EGR jet is not observed, indicating complete vertical momentum exchange between the EGR and air flow at this mixing length. Also, the streamlines at the center of the duct both at z = 5 and z = 10 cm show formation of weak counter rotating vortices. At r = 1.2, Fig. 11(b), a larger region compared to the case of r = 0.45, with downward velocity is observed located at top of the cross section of z = 5 cm. This is explained by the larger downward momentum of EGR jet. Moving downstream, at z = 10 cm, this downward flow of EGR is mostly damped by momentum exchange with air flow, leading to clear formation of strong counter-rotating vortices with length scale of 1–2 cm. Fig. 11 (c) shows the high momentum ratio of r = 2.66, in which at z = 5 cm a very large downward flow is observed in the upper half of the cross section. This flow field is in agreement with the trend observed above, and indicates the large momentum of EGR jet moving downward to the air path. Similar to previous cases, the downward flow is mostly damped at z = 10 cm, and formation of vortex pairs with length scale of 1–3 cm is observed. In addition to above observations, a small clockwise bulk swirl is observed in all the cases in Fig. 11 specially close to the compressor inlet at z = 10 cm, which may be induced by impeller rotation to the upstream field. The out-of-plane (axial) velocity contour is shown in Fig. 12 at different values of EGR momentum ratio. At r = 0.45, a region of low

4.4.2. Quantitative analysis of mixing flow field For quantitative evaluation of the LP-EGR mixing flow field, three metrics are defined and calculated at each cross-sectional plane downstream of the EGR introduction point. (1) The EGR mixing uniformity index (UI), shown in Eq. (10), with a formulation defined in [34], which takes a value of unity in case of completely uniform EGR, and a value of zero in case of completely unmixed EGR, where regions in the cross section either contain a 100% concentration of EGR or a 100% concentration of air. In Eq. (10), c is the local molar concentration of EGR in the mixture, c¯ is the average concentration of EGR in a given cross section and Atot is the total area of the cross section. (2) The velocity uniformity index, Γ, Eq. (11), is defined in a similar fashion reflecting the non-uniformities of axial velocity component w.

UI = 1 −

Γ=1−

∫ |c − c¯| dA 2Atot c¯

(10)

∫ |w − w¯| dA 2Atot w¯

(11)

An important flow quantity is the axial vorticity field which has a spatial distribution ωz (x, y) at the plane of interest upstream of the compressor (for instance the compressor inlet plane). To distinguish between the bulk swirl and smaller scale axial vorticity we decompose the axial vorticity field as a mean and a deviation from the mean, ωz (x , y ) = ω¯ z + ωz, rel (x , y ) . The mean axial vorticity is calculated using area averaging in Eq. (12), which is essentially equal to the flow circulation divided by cross sectional area. This quantity reflects the bulk swirl present in the flow. Here, the sign of circulation is positive if the flow is rotating in the opposite direction of the impeller. Note that the circulation in a closed boundary flow with non-slip wall boundary condition would always be zero. However, the flow field scanning in this study did not include a 6 mm ring near wall region due to the dimensions of the probe tip; therefore, the flow field measurements here exclude the boundary layer and a finite value of total circulation is obtained by integration in the scanned area in Eq. (12).

ω¯ z =

9

∫ ωz dA Atot

(12)

Applied Energy 261 (2020) 114349

A. Reihani, et al.

Fig. 10. Scanned temperature field in the LP-EGR mixing zone at two plane sections, 5 and 10 cm downstream of the EGR port. LP-EGR inlet temperature is 80 °C. Measurements at compressor speed of 40 krpm, normalized flow rate of 40%, and EGR momentum ratios of: (a) r = 0.45, (b) r = 1.2, and (c) r = 2.66.

Similarly, we can calculate the total amount of vorticity present in the cross-sectional plane, excluding the bulk swirl, using Eq. (13), which we call relative axial vorticity. Please note the absolute value in Eq. (13), which essentially sums all the vorticity in the flow irrespective of the direction. This includes all the smaller scale vorticity in the mixing field, such as the counter-rotating vortex pair present in jet-incross flow.

ω¯ z, rel =

∫ |ωz − ω¯ z | dA Atot

(13)

The EGR (temperature) uniformity index (UI) is shown in Fig. 14 (a) as a function of EGR momentum ratio. It can be seen that the UI increases monotonically as the momentum ratio increases at z = 10 cm downstream the EGR introduction. This trend is expected as the higher momentum ratio leads to formation of stronger eddies in the flow which improves the turbulent mixing. However, this trend is not seen at lower mixing length of z = 5 cm, where the point of r = 0.45 shows an 10

Applied Energy 261 (2020) 114349

A. Reihani, et al.

Fig. 11. Scanned in-plane 2D velocity vectors and streamlines in the LP-EGR mixing zone at two plane sections, 5 cm and 10 cm downstream of the EGR port. Measurements at compressor speed of 40 krpm, normalized flow rate of 40%, and EGR momentum ratios of: (a) LP-EGR momentum ratio r = 0.45, (b) r = 1.2, and (c) r = 2.66.

compressor efficiency in Fig. 7 (b), and pressure ratio in Fig. 6 (c) one can conclude that lower EGR mixing uniformity does not lead to a reduction in compressor efficiency or pressure ratio. Based on these results, spatial variation in temperature, or density at compressor inlet caused by EGR non-uniformity do not cause a measurable reduction in the compressor performance. The axial velocity uniformity in Fig. 14 (b) shows a non-monotonic behavior vs momentum ratio at both cross sections of z = 5 and 10 cm

unexpectedly high value of UI. By checking the corresponding temperature distribution in Fig. 11 (a), one can see that the creeping jet of EGR near the top wall might not have been completely detected by the EGR probe due to the dimensions of the probe which limits its minimum distance from the wall. Therefore, the UI at r = 0.45, might not reflect the mixing uniformity of the entire cross section and may not be a valid measurement point. By comparing the trend of EGR mixing uniformity in Fig. 14 (a) with 11

Applied Energy 261 (2020) 114349

A. Reihani, et al.

Fig. 12. Scanned out of-plane velocity magnitude in the LP-EGR mixing zone at two plane sections, 5 and 10 cm downstream of the EGR port. LP-EGR inlet temperature is 80 °C. Measurements at compressor speed of 40 krpm, normalized flow rate of 40%, and EGR momentum ratios of: (a) r = 0.45, and (b) r = 2.66.

total circulation as the mixture travels downstream. For analysis of these results one should keep in mind that the cross sections of velocity measurement do not include the 6 mm of near wall region; therefore, the reported circulation does not reflect the boundary layer profile which if included would reduce the total vorticity. Hence, the increase in circulation in this central region can be explained by conversion of linear motion of the EGR jet to angular motion as the mixture moves downstream. In addition, other mechanisms, such as vortex stretching, exist in 3D flow fields which can increase the circulation of stream-wise vortices [35]. This increase in circulation can explain the ambiguous decreasing response of compressor pressure ratio as a function of mixing length (up to the measurement limit of L/D = 2.5) in Fig. 6 (d).

and shows a peak at r = 1.2, and a lowest value at r = 2.66. The axial velocity non-uniformity is generated by partial blockage of the air stream by the EGR jet which has initially zero axial momentum. Fig. 14 (c) demonstrates the relative axial vorticity at each cross section (computed using Eq. (13)) as a function of momentum ratio. Since the velocity vector is mostly in z direction, the vorticity shown in this plot is mostly a stream-wise vorticity. This stream-wise vorticity shows an exponential increase with increasing momentum ratio at z = 5, indicates generation of strong vortices in this mixing length. The vorticity decreases at z = 10 cm especially at r = 2.66 indicating viscous damping of vortices in the mixing section. However, there is still strong vorticity at the mixing length of z = 10 cm. Fig. 15 shows the EGR (temperature) uniformity as a function of total compressor flow rate at constant speed of 40 krpm and EGR momentum ratio of r = 1.2. At z = 5 cm, UI shows a strong dependence on total flow rate, which indicates a large drop in UI as flow rate increases from 10% to 40%, then shows a small increase as the flow rate is increased to 80%. The same trend with lower variation is observed at z = 10 cm; however, in this case the value UI has increased significantly since the mixing length is increased. Fig. 16 indicates the total circulation (or bulk swirl) at different cross sections as a function of EGR introduction angle. It can be seen that with concentric EGR introduction (θEGR = 0), there is small value of negative circulation (co-rotating with the impeller) which increases in magnitude as the cross section gets closer to the impeller. This rotation can be attributed to rotor-induced rotation which can be transmitted upstream the compressor. At the eccentric EGR introduction of (θEGR = 60 deg) the circulation magnitude significantly increases which is expected as the EGR introduction in this case imparts an angular momentum to the flow field. It is noteworthy that the magnitude of total circulation, measured in the cross section, is increased as the mixture moves downstream from z = 5 cm to 10 cm. This trend seems counterintuitive, since the viscous damping is expected to reduce the

4.5. Figure of merit for LP-EGR mixer design In this section, we aim to obtain the main mechanism of compressor performance degradation as a result of perturbed flow field at compressor inlet, and leverage this finding to obtain an efficient figure of merit for LP-EGR mixer design. This is done by analyzing the compressor response presented in Fig. 6 and Fig. 7 as well as the flow field quantities close to compressor inlet (z = 10 cm) shown in Fig. 14. Note that, the response of a centrifugal compressor to a bulk swirl at the inlet is well known in the literature [36,37]. A co-rotating swirl (w.r.t. impeller) reduces the flow rate and the pressure ratio, and could increase the efficiency and surge margin. A counter-rotating swirl has the opposite effects [28]. However, the compressor efficiency degradation is observed here with very small bulk swirl values in the concentric EGR introduction cases. Therefore, the mechanism is related to smaller scale flow structures and perturbations at compressor inlet. As can be seen in Fig. 17 a strong correlation is observed between the relative axial vorticity ωz,rel and change in compressor pressure ratio and isentropic total-to-total efficiency. The relative vorticity, defined in Eq. (13), excludes the bulk flow pre-whirl. This quantity 12

Applied Energy 261 (2020) 114349

A. Reihani, et al.

Fig. 13. Mixing flow field with eccentric LP-EGR introduction. (a) Scanned temperature field, (b) in-plane velocity vector, and (c) out-of-plane velocity magnitude in the mixing zone at two plane sections, z = 5 and 10 cm downstream of the EGR port. LP-EGR inlet temperature is 80 °C. Measurements at compressor speed of 40 krpm, and normalized flow rate of 40% and EGR momentum ratios of r = 1.2 at EGR introduction angle of θEGR = 60°.

impeller inlet, decreasing the compressor efficiency and pressure ratio. Therefore we define α¯ as a measure of the mean angular perturbation of the velocity vector from axial direction caused by ωz,rel in Eq. (14). Where Dim is the inlet duct diameter, and the quantity v¯tan = ω¯ z, rel Dim 4 , is the mean tangential velocity caused by relative vorticity at the inlet of compressor obtained from Stokes’ theorem

essentially captures the sum of all axial vortices at all length scales captured by the scanning probes, for instance this includes the counterrotating vortex pairs observed in Fig. 11. This observation is in agreement with previous numerical simulations by the author [26], indicating that stream-wise vortices can be sustained in the mixing section for relatively long distances and therefore can reach the impeller inlet. At the impeller inlet, depending on vorticity magnitude and length scale, these vortices can perturb the local incidence angle at

(ω¯

z , rel

=

∮ Vtan dl Atot

=

4 v¯ Dim tan

). The value of α¯ is included as the upper

horizontal axis of Fig. 17.

13

Applied Energy 261 (2020) 114349

A. Reihani, et al.

Fig. 16. Total flow axial circulation or bulk swirl (obtained using Eq. (12)) as a function of LP-EGR introduction angle (negative value indicates co-swirl with respect to the impeller rotation).

Fig. 17. Relative change in compressor pressure ratio and isentropic total-tototal efficiency as a function of relative axial vorticity (bottom horizontal axis) or velocity perturbation angle (top horizontal axis) at compressor inlet. Measurements at compressor speed of 40 krpm, and normalized flow rate of 40% corresponding to mean axial velocity of 12.2 m/s.

Fig. 14. Mixing flow field parameters calculated at z = 5 and 10 cm cross sections downstream of the EGR port showing: (a) EGR uniformity index, (b) Axial velocity uniformity index, and (c) relative axial vorticity obtained using Eq. (13).

Interestingly, there is no negative correlation between EGR nonuniformity at compressor inlet and compressor performance. Nor there is a clear negative correlation between axial velocity non-uniformity and compressor performance metrics. These results are counter-intuitive in the sense that the EGR non-uniformity does not have any measurable negative effect on compressor performance. This observation is contrary to the literature for LP-EGR mixer design where obtaining EGR mixing uniformity index at low pressure drop is used as the design objective [27]. Given the tight packaging constraints, obtaining high mixing uniformity is challenging and leads to higher mixer pressure drop. Based on the results in Fig. 17, a more efficient objective for LP-EGR mixer design is obtaining a vortex-free flow field at compressor inlet (i.e. minimizing ωz, rel), relaxing the requirement on the LP-EGR mixing uniformity which enables design of mixers with lower pressure drop.

Fig. 15. LP-EGR uniformity index vs total compressor flow rate at speed of N = 40 krpm, and momentum ratio of r = 1.2.

5. Conclusions

ω¯ z, rel Dim ⎞ v¯ α¯ = tan−1 ⎛ tan ⎞ = tan−1 ⎛ ⎝ v¯axial ⎠ ⎝ 4v¯axial ⎠

The effect of Low-Pressure Exhaust Gas Recirculation (LP-EGR) introduction on the turbocharger compressor performance was experimentally studied on a hot gas stand using response surface methodology. The 3D flow field inside the LP-EGR mixer was measured using

⎜

⎟

⎜

⎟

(14)

14

Applied Energy 261 (2020) 114349

A. Reihani, et al.

Declaration of Competing Interest

a scanning probe. As a case study, a reconfigurable T-junction mixer geometry was used. The effect of various geometric parameters such as mixing length, Lmix, and LP-EGR introduction angle, θEGR, as well as operating parameters including total compressor flow rate, LP-EGR fraction, and compressor speed on mixing flow field and the compressor performance was investigated. The mixing flow field of LP-EGR and air upstream of the compressor was significantly affected by the dimensionless quantity EGR-to-air momentum ratio. Scanning the temperature distribution in the mixing section was used as a surrogate for measurement of EGR distribution. The EGR uniformity, close to compressor inlet, was measured to be 60.5% at low momentum ratio of 0.45 with a monotonic increase to 91.8% at momentum ratio of 2.66. The axial flow velocity was also perturbed by LP-EGR introduction showing regions of low velocity near the EGR stream due to low axial momentum of EGR jet. However, close to the compressor inlet, the axial velocity uniformity was higher than 96% at all measured conditions. Introduction of LP-EGR generated significant stream-wise vorticity in the mixing section with a length scale significantly smaller than the compressor inlet diameter. The vorticity intensity increases with EGRto-air momentum ratio and decreases with mixing length. These streamwise vortices can reach the impeller inlet (at mixing lengths up to 2.5), and depending on intensity and length scale, perturb the local incidence angle at impeller leading edge. In addition to the aforementioned small-scale vorticity, with eccentric LP-EGR introduction a bulk swirl with lengths scale comparable to mixer duct diameter was induced in the mixing section, which was advected to the impeller inlet, globally altering the inlet incidence angle similar to the effects of an inlet guide vane. The compressor isentropic total-to-total efficiency, and pressure ratio are altered by the EGR-induced flow field as summarized below. The pressure ratio decreases (up to 3%), with a larger decline at higher compressor speeds, higher flow rates, and higher LP-EGR momentum ratios. The pressure ratio decreases further by introducing eccentric EGR with θEGR > 0, generating pre-whirl co-rotating with compressor impeller, which can be explained by modification of global inlet incidence angle. Compressor isentropic total-to-total efficiency, ηt-t, degrades significantly (up to 25% rel.) by introducing LP-EGR compared to no LPEGR. The efficiency degradation increases at higher EGR momentum ratios, and at higher compressor flow rates; however, it is almost insensitive to EGR eccentricity. Significant efficiency drop is observed at longer mixing lengths up to length-to-diameter ratio (L/D) = 2.5. The mechanism of compressor efficiency reduction is suggested to be the formation of strong (mostly stream-wise) vortices in the EGR-air upstream of the compressor, leading to perturbation of the local incidence angle at the impeller inlet. Contrary to previous studies which use EGR uniformity as a figure of merit for LP-EGR mixer design, in this study it was shown that EGR (or temperature) non-uniformity does not show a negative impact on compressor performance. Therefore, here we propose a more efficient figure of merit for LP-EGR mixer design which is obtaining a vortex-free flow at compressor inlet in order to minimize the perturbation of local incidence angle at leading edge of the impeller. This objective is generally less stringent than obtaining high EGR uniformity which enables design of efficient LP-EGR mixers with lower pressure drop and therefore higher system efficiency. In summary, the findings in this study provide insights into the significant effects of LP-EGR induced flow field on the turbocharger compressor performance which is of paramount importance to the engine fuel conversion efficiency. Further it puts emphasis on importance of efficient LP-EGR mixer design, and proposes an efficient figure of merit for mixers in order to benefit from the fuel economy advantages of LP-EGR architecture in diesel and GDI engines.

The authors declared that there is no conflict of interest. Acknowledgment Funding for this research carried out at the University of Michigan Walter E. Lay Automotive Laboratory was provided by Ford Motor Company under the Ford/UM Alliance Program. The authors would like to thank Keith Plagens, Craig Stanfill, and Timothy Gardner of Ford’s Gas Turbine Laboratory, for their help and support conducting the gas stand experiments. References [1] EPA. Final Determination on the Appropriateness of the Model Year 2022-2025 Light-Duty Vehicle Greenhouse Gas Emissions Standards under the Midterm Evaluation. Epa; 2017. [Online]. Available: https://www.epa.gov/regulationsemissions-vehicles-and-engines/midterm-evaluation-light-duty-vehicle-greenhousegas. [2] Johnson TV, Joshi A. Review of deNO x Technology for Mobile Applications, no. 33, 2018. [3] Sellnau M, Hoyer K, Moore W, Foster M, Sinnamon J, Klemm W. Advancement of GDCI engine technology for US 2025 CAFE and Tier 3 emissions; 2018. [4] Koerfer T, Tomazic D, Bick W, Menne C, Franke M. Meeting 2025 CAFE Standards for LDT with Fuel-Efficient Diesel Powertrains-Approaches and Solutions; 2017. [5] Stanton D, Charlton S, Vajapeyazula P. Diesel engine technologies enabling powertrain optimization to meet US greenhouse gas emissions. SAE Int J Engines 2013;6(3):1757–70. [6] Johnson T, Joshi A. Review of vehicle engine efficiency and emissions. SAE Int J Engine 2018;11. [7] Zheng M, Reader GT, Hawley JG. Diesel engine exhaust gas recirculation––a review on advanced and novel concepts. Energy Convers Manage 2004;45(6):883–900. [8] Wei H, Zhu T, Shu G, Tan L, Wang Y. Gasoline engine exhaust gas recirculation–a review. Appl. Energy 2012;99:534–44. [9] Mao B, Yao M, Zheng Z, Li Y, Liu H, Yan B. Effects of dual loop EGR on performance and emissions of a diesel engine; 2015. [10] Abd-Alla GH. Using exhaust gas recirculation in internal combustion engines: a review. Energy Convers Manage 2002;43(8):1027–42. [11] Wahlstrom J, Eriksson L, Nielsen L. EGR-VGT control and tuning for pumping work minimization and emission control. IEEE Trans Control Syst Technol 2010;18(4):993–1003. [12] Bittle J, Zheng J, Xue X, Song H, Jacobs T. Cylinder-to-cylinder variation sources in diesel low temperature combustion and the influence they have on emissions. Int J Engine Res 2014;15(1):112–22. [13] Park Y, Bae C. Experimental study on the effects of high/low pressure EGR proportion in a passenger car diesel engine. Appl Energy 2014;133:308–16. [14] Torregrosa AJ, Olmeda P, Martin J, Degraeuwe B. Experiments on the influence of inlet charge and coolant temperature on performance and emissions of a DI Diesel engine. Exp Therm Fluid Sci 2006;30(7):633–41. [15] Luján JM, Climent H, Novella R, Rivas-Perea ME. Influence of a low pressure EGR loop on a gasoline turbocharged direct injection engine. Appl Therm Eng 2015;89:432–43. [16] Van Aken M, Willems F, de Jong DJ. Appliance of high EGR rates with a short and long route EGR system on a heavy duty diesel engine; 2007. [17] Zamboni G, Capobianco M. Experimental study on the effects of HP and LP EGR in an automotive turbocharged diesel engine. Appl Energy 2012;94:117–28. [18] Cornolti L, Onorati A, Cerri T, Montenegro G, Piscaglia F. 1D simulation of a turbocharged Diesel engine with comparison of short and long EGR route solutions. Appl Energy 2013;111:1–15. [19] Heuwetter D, Glewen W, Meyer C, Foster DE, Andrie M, Krieger R. Effects of low pressure EGR on transient air system performance and emissions for low temperature diesel combustion; 2011. [20] Zamboni G, Moggia S, Capobianco M. Hybrid EGR and turbocharging systems control for low NOX and fuel consumption in an automotive diesel engine. Appl Energy 2016;165:839–48. [21] Wang J. Air fraction estimation for multiple combustion mode diesel engines with dual-loop EGR systems. 2007 46th IEEE conference on decision and control. 2007. p. 2862–7. [22] Serrano JR, Piqueras P, Angiolini E, Meano C, De La Morena J. On cooler and mixing condensation phenomena in the long-route exhaust gas recirculation line; 2015. [23] Siokos K, Koli R, Prucka R, Schwanke J, Miersch J. Assessment of cooled low pressure egr in a turbocharged direct injection gasoline engine. SAE Int J Engines 2015;8(4):1535–43. [24] Hoard J, Reihani A, Singh MR, Styles DJ. Method for utilizing condensate to improve engine efficiency. Google Patents; 2018. [25] Karstadt S, Werner J, Münz S, Aymanns R. Effect of water droplets caused by low pressure EGR on spinning compressor wheels. 19th Supercharging Conference Dresden. 2014. [26] Reihani A, Hoard J, Klinkert S, Kuan C-K, Styles D. Numerical evaluation of the

15

Applied Energy 261 (2020) 114349

A. Reihani, et al.

[27]

[28] [29]

[30]

[31] Box GEP, Behnken DW. Some new three level designs for the study of quantitative variables. Technometrics 1960;2(4):455–75. [32] TURBOCHARGER GAS STAND TEST CODE, SAE, vol. J1826; 1995. [33] Devore JL. Probability and statistics for engineering and the sciences. Cengage Learning 2011. [34] Reihani A, Corson B, Hoard JW, Fisher GB, Smirnov E, Roemer D, et al. Rapidly pulsed reductants in diesel NOx reduction by Lean NOx traps: Effects of mixing uniformity and reductant type. SAE Int J Engines 2016;9(3):1630–41. [35] Jiménez J. Turbulence and vortex dynamics. Notes Polytecnique Course Turbul 2004. [36] Kyrtatos N, Watson N. Application of aerodynamically induced prewhirl to a small turbocharger compressor; 1980. [37] Mohtar H, Chesse P, Yammine A, Hetet JF. Variable inlet guide vanes in a turbocharger centrifugal compressor: local and global study; 2008.

effects of low pressure EGR mixer configuration on turbocharger compressor performance. ASME 2018 internal combustion engine division fall technical conference. 2018. p. V002T06A017–V002T06A017. Brune K-H, Schiffer H-P, Christmann R, Gnewikow M. Experimental investigations of the disturbed inlet-flow structure caused by mixing geometries and its influence on the performance of a turbocharger centrifugal compressor. ASME Turbo Expo 2009: Power for Land, Sea, and Air. 2009. p. 1295–304. Forlese J, Spoleti G. Impact of inlet coherent motions on compressor performance. IOP Conf Series: Mater Sci Eng 2017;232(1):12043. Cambonie T, Gautier N, Aider J-L. Experimental study of counter-rotating vortex pair trajectories induced by a round jet in cross-flow at low velocity ratios. Exp Fluids 2013;54(3):1475. Kamotani Y, Greber I. Experiments on a turbulent jet in a cross flow. AIAA J 1972;10(11):1425–9.

16