Original article

Proc IMechE Part DJ Automobile Engineering2019 Vol 233(8) 2246ndash2256 IMechE 2018Article reuse guidelinessagepubcomjournals-permissionsDOI 1011770954407018757926journalssagepubcomhomepid

Design criterion for asymmetrictwin-entry radial turbine for efficiencyunder steady and pulsating inletconditions

Aolin Wang and Xinqian Zheng

AbstractThe internal combustion engine plays an important role in energy conservation and environmental protection An effi-cient way to simultaneously reduce the engine fuel consumption and the emissions (NOx) is an asymmetric twin-entryturbine turbocharging system The asymmetric twin-entry turbine uses a volute that has two scrolls in the axial directionwith different throat areas The smaller area scroll can increase the backpressure to support the exhaust gas recircula-tion system for lower NOx emissions meanwhile the larger area scroll can decrease the backpressure to reduce theexhaust resistance for lower fuel consumption The area ratio is dependent on the required exhaust gas recirculationrate and fuel consumption However there are no guidelines for how to arrange the scrolls and whether placing the bigscroll on the hub side or the shroud side can provide improved results remains unknown In this paper two differentdesign types are discussed for the asymmetric turbine one with the big scroll on the hub side while the other is on theshroud side Also the unsteady computational fluid dynamics (CFD) simulation is under steady and pulsating inlet condi-tions The results show that when under steady inlet conditions the efficiency of the asymmetric turbine with the bigscroll on the shroud side is nearly 16 higher However when facing pulsating inlet conditions the efficiency of theasymmetric turbine with the big scroll on the hub side is almost 11 higher The difference is due to the different massflow storage capacities of the two scrolls under pulsating inlet conditions The design criterion for asymmetric twin-entry turbine states that if the turbocharger is designed for a large volume exhaust manifold it is better to place the bigscroll on the shroud side If for a small volume exhaust manifold it is better to set the big scroll on the hub side

KeywordsAsymmetric turbine twin-entry volute efficiency steady and pulsating exhaust gas recirculation

Date received 23 July 2017 accepted 10 January 2018

Introduction

Turbocharger system has been an essential part of die-sel engine due to its ability to increase power outputand engine efficiency and exhaust gas recirculation(EGR) system has been an efficient way to decrease theNOx emission of diesel engine1 Lapuerta et al2 evalu-ated a diesel engine NOx emission and found that EGRsystem together with turbocharger system can reducethe emission effectively Zamboni et al3 got the similarconclusion and found that the turbocharger systemhelped the engine reach better fuel consumptionFigure 1 shows an exhaust system of a diesel engineincluding EGR and turbocharger system The EGRsystem takes part of exhaust gas from turbine inlet tocompressor outlet depressing the oxygen concentrationof the inlet air to cut down the production of NOx

Therefore the back-pressure of EGR inlet (also turbineinlet) is required to be high enough to prevent backflowfrom EGR out (also the compressor outlet) As for tur-bocharger system especially the turbine part it isknown that the back-pressure of turbine inlet is relatedto the inlet throat area4 and the higher pressure requiressmaller throat area while the decrease of throat areawill enhance the flow resistance and worse the enginescavenging process5 Based on this turbine requires big

Turbomachinery Laboratory State Key Laboratory of Automotive Safety

and Energy Tsinghua University Beijing China

Corresponding author

Xinqian Zheng Turbomachinery Laboratory State Key Laboratory of

Automotive Safety and Energy Tsinghua University Beijing 100084 China

Email zhengxqtsinghuaeducn

throat area to reduce back-pressure and this is oppo-site to the requirement of EGR system neither thetraditional single-entry turbine nor the symmetric twin-entry turbine can meet their demand at the same timeHowever the asymmetric twin-entry turbine can makeit as shown in Figure 1 this kind of turbine has twoscrolls with two different throat areas the larger oneprovides sufficient flow capacity for power output andthe smaller one can increase the back-pressure for highenough EGR rate Zhu and Zheng6 studied the asym-metric twin scroll turbocharging system and found itcan improve the engine efficiency and emission perfor-mance Schmidt et al7 named the two different scrollsas lsquolsquoEGR scrollrsquorsquo for the smaller and lsquolsquoLambda scrollrsquorsquofor the larger to represent their different functions

Since the two scrolls are designed asymmetrically inthe axial direction it is interesting to think about theeffect it may have on turbine performance We knowthat the impeller of a turbine is also asymmetric in theaxial direction and the shroud side has a more signifi-cant curvature than hub side is this feature has anyeffect on turbine efficiency or flow field Unfortunatelyit remains unclear nowadays Based on this it is impor-tant to address which arrangement of scrolls the bigscroll on the impeller hub side or the shroud side canprovide better efficiency Currently the most popularkind of asymmetric twin-entry turbine places the bigscroll on the shroud side as shown in Figure 1 Thistype of arrangement is based on a list of previous stud-ies Dale and Watson9 measured a symmetric twin-entry turbinersquos performance over a wide range of steadypartial inlet conditions They found that for the sym-metric twin-entry turbine the peak efficiency pointoccurred when the scroll at the impeller shroud side hasmore mass flow than the hub side scroll instead of thefull inlet condition However the authors did notexplain this phenomenon Baines and colleagues10ndash12

found the same phenomenon whereby the flow pre-ferred to enter the impeller from the shroud side Theyperformed experiments and measured the performanceand the flow field of a twin-entry radial turbine understeady full and partial inlet conditions And they

hypothesized that the impeller hub side leading edge ismore sensitive to positive incidence angle and this maycause a penalty for efficiency Still no experimental evi-dence can support the hypothesis Based on the conclu-sion above the asymmetric twin-entry turbine is mainlydesigned with the big scroll on the shroud sideTherefore the more mass flow goes into impeller in thehigh-efficiency region and this can provide high perfor-mance Mualler et al13 give the same explanation fortheir asymmetric twin-entry turbine on a Benz 11-L die-sel engine Brinkert et al8 also used the same asym-metric model to represent the asymmetric twin-entryturbine

However the research works above are based onlyon steady inlet conditions without any further investi-gation under pulsating inlet conditions Since the twin-entry turbocharger is wildly used on gasoline engineswhich almost adopt the pulsating turbocharging systemto decrease the turbine lag it is essential to investigatethe turbine performance under pulsating inlet condi-tions to make clear whether it keeps the same as or dif-fer from the steady inlet conditions Quasi-steadyassumption is one of the available methods to analyzeturbine pulsating performance if the turbine operationwere quasi-steady the pulsating flow results would fol-low the steady-flow curve that is to say if the quasi-steady assumption is satisfied the turbine performancekeeps the same under steady and pulsating inlet condi-tions However research works of Wallace and col-leagues14ndash16 and Woods and Norbury17 reported thatthe quasi-steady assumption tends to be unsatisfactorywhen pulsating frequency increases Dale and Watson9

tested the pulsating turbine performance on a test rigand their results showed a noticeable deviation fromthe steady-flow curve demonstrating the inadequacy ofthe quasi-steady assumption The more recent investi-gations on the unsteady performance of twin-entry tur-bines were given by Galindo et al18 they used the CFDmethod to simulate the flow characteristics of a radialturbine in pulsating flow and found that the primarysource of non-quasi-steadiness of the turbine is the tur-bine volute If the turbine volute has greater length andvolume the degree of non-quasi-steadiness becomessevere Other work also focused on the quasi-steadyassumption such as the research work in Martinez-Botas and colleagues19ndash21 Based on these studies DeBellis et al22 carried out one-dimensional (1D) simula-tion and experiment to analyze turbine performanceunder steady and unsteady conditions and found thatunder a pulsating flow the turbine rotor can be seen asa steady state while the volute showed storage effectswithin it The flow in the volute is influenced byunsteady effects of pulsating inlet and the flow field atthe rotor inlet is affected which may lead to a changein turbine efficiency

Therefore compared with under steady inlet condi-tions the performance of the turbine is quite differentwhen under pulsating inlet conditions and it is thevolute that causes the difference Due to this the

Figure 1 EGR and turbocharger system on diesel engine8

EGR exhaust gas recirculation

Wang and Zheng 2247

popular type of asymmetric turbine shown in Figure 1may not be suitable for pulsating turbocharger systemso it is important to investigate the effect of the asym-metric volute on turbine performance and find outwhich type of arrangement can better suit pulsatinginlet conditions Unfortunately there is relatively lessresearch on asymmetric turbine under pulsating inletcondition and even rare study on another type of asym-metric volute that places the small scroll on the shroudside

In the present paper the authors compared the per-formance and flow field of two different types of asym-metric turbines together with another symmetricturbine under steady and pulsating inlet conditionsBased on the different performances of the three tur-bines under different inlet conditions a brief explana-tion is given and a design criterion for the asymmetrictwin-entry turbine when under steady or pulsating inletconditions is described

Test facility

The turbine dynamometer test rig on which the turbinecan be tested in isolation from the rest of the turbochar-ger is shown in Figure 2 There are four pressure sensorsat the turbine inlet the arrangement is equispaced at thecircumferential direction and the measurement precisionis 004 The number of inlet temperature sensors is 3and they have the same arrangement the precision is025C At the turbine outlet the arrangement of thepressure sensors is same as inlet condition but there isonly one temperature sensor The hydraulic dynam-ometer is used to absorb the power output and the for-mula to calculate the turbine efficiency ht is as follows

ht =Pt

mf3Cp3DT0eth1THORN

where Pt is the turbine output power that is tested bydynamometer and mf is mass flow rate DT0 is the

difference between the total temperature of the inletT01 and the outlet T03

For further explanation the parameter that thedynamometer can measure directly is the torque so Pt

is defined as follows

Pt =Tq3n

9550eth2THORN

where Tq is the torque and n is the rotational speedWhatrsquos more considering the inaccuracy of the tem-perature measurement caused by the uniformity of theturbine outlet condition the value of the outlet totaltemperature T03 is calculated as follows

T03 =T01

P01

P03

g

g1eth3THORN

where P01 and P03 are the inlet and outlet total pres-sures respectively As the static inlet temperature of thetest rig is controlled under 40C heat exchange withthe ambient environment can be neglected and the flowcan be considered as isentropic

Numerical methodology

Numerical model

There are three types of models used to analyze thesymmetric and asymmetric twin-entry scroll turbinesThe turbine geometry models which are shown inFigure 3 consist of the volute and the impeller Thedetailed parameters of the geometry are shown inTable 1 The three models use the same impeller butare combined with different volutes The asymmetricdesign refers to that where the throat area of the voluteshroud side scroll is bigger than that of the hub sideand the asymmetric OP is the model with the oppositefeatures as the asymmetric design The total throat areaof these three models remains the same and they alsohave the same volute cross-sectional area distributionfeature to ensure a similar flow capacity

The prototype of the asymmetric turbine is a com-mercial turbine installed at a 11-L inline six-cylinderdiesel engine The six cylinders are connected in twogroups of three and each group feeds one inlet of the

Figure 2 Sketch map of the test benchFigure 3 Turbine geometry models (a) asymmetric(b) symmetric and (c) asymmetric OP

2248 Proc IMechE Part D J Automobile Engineering 233(8)

twin-entry turbine This paper focuses on the differentperformance of the asymmetric and asymmetric OP tur-bine the symmetric type works as a reference

Considering the nonuniform circumferential flowfield caused by the volute at the impeller inlet it is nec-essary to simulate the entire impeller flow domain asopposed to the single passage domain The mesh of theimpeller and volute are shown in Figure 4 The struc-tured mesh is used for the impeller with a total of4360760 grids points The computational domain ofthe volute adopted the unstructured mesh with a totalof 2037995 grids points Therefore the total gridnumber of the model is more than 6 million

Numerical method

The simulation was conducted with the ANSYS CFXsolver which is based on the three-dimensional com-pressible finite volume scheme to solve steady-stateReynolds-averaged NavierndashStokes equations with theconservative formulation Considering the calculationcost and time the kndashe turbulence model is selected andcan accurately simulate the flow separation and vor-tices The validation of the mesh independence is shownin Figure 5 and as seen 5 or 6 million grids points canprovide reasonably accurate numerical results

Simulation procedure

The absolute total pressure and total temperature areimposed as the inlet boundary conditions the absolute

static pressure is imposed as the outlet boundary condi-tion and the data are derived from the experimentalmeasurements Inlet1 and inlet2 employ the boundaryconditions separately to make different inlet conditionsfor each scroll when under the pulsating inflow condi-tion At solid boundaries as the turbine test rig adoptsthe cold inlet air and there is heat insulator outside thepipe adiabatic wall conditions are imposed for thesimulation which means no slip and no heat transferThe frozen rotor model is imposed at the interface

Simulation validation

The results comparing the simulated total-to-staticisentropic efficiency and the mass flow rate under fulladmission conditions with the experiments are shownin Figure 6 The simulated efficiency is defined asfollows

hts =T01 T03

T01 1 P3

P01

(g1)=g eth4THORN

where T01 and T03 are the inlet and outlet total tem-peratures respectively P3 and P01 are the static outletpressure and total inlet pressure respectively

The experiments concern the asymmetric turbine(a) and the four lines refer to the rotational speed64000 80000 96000 and 112000 rmin respectively(Figure 3) The results indicate that the simulation is inreasonable agreement with the experimental data bothfor the efficiency and for the flow capacity with a maxi-mum discrepancy of 09 The noticeable differencesare toward the lower speed and lower expansion ratiorange of the efficiency This may be a result of theexperiment deviation because it is difficult to maintainthe test speed at the required value at lower rotatingspeeds and the deviation would be more significantthan at higher rotating speeds The good agreement ofthe flow capacity can also prove the conclusion becausethe test of the mass flow rate is negligibly affected bythe change in speed

Table 1 Turbine geometric parameters

Parameter Value

Blade numbers 10Volute outlet radius 4280 mmVolute outlet width 1138 mmImpeller inlet radius 4175 mmImpeller outlet radius 2766 mmTip clearance 045 mmInlet blade angle 0Throat area 1168 mm2

Figure 4 Single passage of the impeller mesh (a) and thedomain and boundaries (b)

Figure 5 The grid size independence validation

Wang and Zheng 2249

The CFD results and experiment data can matchwell with each other under steady inlet conditions andexperiments under pulsating inlet conditions will becarried out in the future to fine down the validation ofthe numerical model

In this paper first the performance and flow fieldsof the three types of turbines are compared understeady inlet conditions to determine the mechanism ofthe efficiency change and the geometric effects on theturbine performance Then the performance and flowfields under pulsating inlet conditions are investigatedto discover the pulse effects Finally based on theresults and analysis a criterion is established for theasymmetric twin-entry turbine design under steady andpulsating inlet conditions

Results under steady equal inlet conditions

The numerical results of the three simulated models areshown in Figure 7 The volute throat area was designedto maintain the same flow capacity in the three modelsas shown in the mass flow rate map It can be seen thatat low speeds the mass flow rates of the three turbine

types are nearly the same but at the highest speedthere was a drop in the mass flow rate of approximately02 which may cause a drop in the efficiency of thesymmetric turbine at high speeds putting it close to theasymmetric OP model Over the entire range of work-ing conditions the asymmetric turbine has the best tur-bine efficiency while the OP has the worst with thesymmetric turbine in the middle What is interesting isthat the efficiency drops between these three models atintervals of approximately 08 Based on this findingwe take the speed line 96000 rmin condition for fur-ther investigation to simplify the research process

Figure 8 shows the turbine and component perfor-mance comparison results It can be seen that the flowloss in the volute remains nearly constant with theexpansion ratio while the rotor efficiency shows thesame trend as turbine efficiency Although the sym-metric turbine has the lowest total pressure loss coeffi-cient this contributes little to the turbine performanceand considering that the test data used for the calcula-tions are measured at the interface the mixing flow atthe volute outlet domain may contribute to the rise inthe asymmetric and OPrsquos loss coefficients It is clear

Figure 6 Performance comparison between the numerical andexperimental results (a) efficiency and (b) mass flow rate

Figure 7 Performance comparison between the symmetricand the two different asymmetric turbines (a) efficiency and(b) mass flow rate

2250 Proc IMechE Part D J Automobile Engineering 233(8)

that the turbine efficiency is mainly determined by therotor efficiency and that the flow loss in the volute hasa negligible effect The main function of the volute is toprovide the rotor inlet flow field which may affect therotor efficiency Therefore even if the volute has differ-ent geometries the key point to determine the turbineefficiency is the flow field at the impeller inlet position

Figure 9 shows the shroud to hub spanwise distribu-tion for the rotor efficiency and radial velocity at theimpeller leading edge The rotor efficiency distributionis a hypothetical value that signifies the specific poweroutput that a unit mass of gas can provide if it entersthe impeller at a given spanwise position The valuesare calculated along the spanwise direction at the lead-ing edge and the corresponding trailing edge The radialvelocity represents the inlet mass flow and the rotorefficiency distribution and the match with the massflow distribution determine the impeller efficiencyThus if the high radial velocity position matches wellwith the high-efficiency region a senior rotor efficiencyis achieved

In Figure 9 the efficiency is higher at the hub sidethan at the shroud side and the asymmetric case hasthe highest efficiency while the OP has the lowest Forthe mass flow distribution for the asymmetric casesymmetric case and OP case the maximum mass flowposition moves from the hub side to the shroud side orfrom the high-efficiency region to the low-efficiencyregion These two reasons can explain the differences inthe rotor efficiency

Figure 10 shows the rotor inlet incidence angle ofthe three turbine models along the plane at 001mm

upstream of the rotor leading edge which is the criticalparameter to evaluate the rotor performance It isrecognized that optimum incidence is in the region of220 to 240 and that a deviation from this region cancause a drop in the efficiency Moustapha et al23 con-cluded that the radial turbine rotors are much moresensitive to positive incidence than to extremes of nega-tive incidence Then Baines24 put forward a postulatethat the rotor is more sensitive to the positive incidencenear the hub surface to explain the difference in the effi-ciencies under the hub and shroud side partial admis-sion condition Nevertheless he did not show anyexperimental or CFD results to support the postulate

It can be seen from Figure 10 that the positive inci-dence appears in the region that corresponds to theposition of the large volute scroll From Figure 10(a) to(c) the positive incidence region moves from the shroudside to the hub side of the impeller causing the effi-ciency to decrease These results correspond to Bainesrsquohypothesis

From the flow field at the rotor inlet region shown inFigure 11 a vortex exists caused by the flow separationat the hub side near the suction surface of the impellerCompared to the static entropy distribution it can beseen that there are two high entropy regions in the flowchannel one of which is near the hub and correspondsto the vortex The other is near the shroud which iscaused by the tip clearance flow Also the entropy gaincaused by the vortex is larger than the tip clearanceentropy gains

Figure 8 Turbine and component performance Figure 9 Rotor efficiency and radial velocity distribution at theimpeller leading edge

Wang and Zheng 2251

Moreover the vortex region is also where the highstatic entropy appears It is well known that a negative

incidence angle can weaken the flow separation whilea positive angle can aggravate it Therefore when thelarge volute scroll is placed on the shroud side the neg-ative incidence region is at the hub side of the impellerThis can weaken the flow separation fade or even erasethe vortex which may improve the turbine perfor-mance When the larger volute scroll is placed on thehub side the positive incidence region is at the hubside which may enhance the degree of the flow separa-tion and the vortex could become severe leading to thedrop in efficiency This is illustrated in the flow field inFigure 12

The static entropy distribution from Figure 13 at thevortex region shows the same conclusion It can be seenthat the asymmetric case only has an entropy gaincaused by tip clearance while the asymmetric OP modelhas the most severe entropy gain at the vortex regionThis can explain the efficiency differences between thesethree models

Figure 14 shows the velocity distribution at thevolute outlet plane at the impeller trailing edge It illus-trates that these three models have nearly the same out-let flow field and their mean velocities are 1403 1412and 1421ms respectively This means that the effi-ciency differences of the impellers are mainly caused bythe difference in the inlet flow condition

Results under the pulsating inlet condition

When the engine adopted the pulsating turbochargersystem the exhaust pressure is a pulse wave with a spe-cific amplitude and a specific main frequency and thefrequency is connected with the engine speed We canuse an ideal sinusoidal signal to represent the realengine exhaust pressure pulse wave as the turbine pul-sating inlet condition This paper has simplified the

Figure 10 Incidence angle distribution at the impeller leadingedge (a) asymmetric (b) symmetric and (c) asymmetric OP

Figure 11 Flow field and static entropy at the impeller inletregion

Figure 12 Flow field in the rotor passage (a) asymmetric (b) symmetric and (c) asymmetric OP

2252 Proc IMechE Part D J Automobile Engineering 233(8)

pulsating inlet condition to a sinusoidal signal whilekept the amplitude and frequency similar to the enginereal exhaust pressure and the prototype pressure waveis from a six-cylinder diesel engine in Mualler et al13

The pulsating frequency of the simulation is 25Hzcorresponding to the power output condition for this 11-L diesel engines This working point demand is morestringent to fuel consumption which means that a highturbine efficiency is required During pulsating inlet con-ditions the volute shroud side scroll and the hub sidescroll face two different inlet pulses that have a reversephase as shown in Figure 15 Considering the two isola-tion inlet conditions of the volute the isentropic efficiencyof the turbine at a specific time is defined as follows

h=m1s +m1h m1s +m1heth THORN T03=T01

m1s 1 (P3=P01s)g1

g

+m1h 1 (P3=P01h)

g1g

eth5THORN

Then the volute total pressure loss coefficient isnewly defined as follows

K=m1sP01s +m1hP01h (m1s +m1h)P02

(m1s +m1h)(P02 P2)eth6THORN

As for rotor efficiency considering that the form ofthe inlet and outlet remains the same as in the steadysituation the calculation remains unchanged

Figure 15 shows the performance of these three tur-bines under a 25-Hz pulsating inlet flow condition The

time-averaged efficiencies for these three turbines are7448 7469 and 7554 corresponding to theasymmetric symmetric and asymmetric OP modelrespectively The asymmetric type has the worst perfor-mance which is a contrary result of the steady perfor-mance where the asymmetric turbine shows the bestperformance This means that the performances of theturbines are different when under steady and pulsatinginlet flow conditions

Under the pulsating inlet conditions the highest effi-ciency points appear for all three models when the huband shroud inlet pressures are equal The efficiencydrops when facing the partial admission conditionespecially in the situation when the hub side inlet hasthe highest back-pressure

When the volute hub side inlet has the lower pres-sure corresponding to less mass flow the efficiencies ofthese three models are nearly the same and the asym-metric turbine is improved over the other twoHowever when the hub side has more pressure theefficiency drops and the asymmetric turbine is affectedthe most causing a decrease in the average efficiencyTherefore the inlet condition at the volute hub side isthe main factor to control the twin-entry turbine per-formance and the decrease in the mass flow may notcause a noticeable rise in the efficiency but the increasein mass flow will lead to an efficiency drop If the twin-entry turbine is symmetric the high efficiency appearswhen the shroud side inlet mass flow rate is larger thanthe hub side and drops when the mass flow is lesswhich agrees with previous conclusions under steadyinlet conditions from the literature

The asymmetric model that places the small scrollon the hub side is affected more by the hub side inletand the OP model that places the big scroll on the hubside is affected less Therefore the small scroll has moreinfluence on the turbine performance In summary thesmall scroll on the hub side is the main cause of deter-mining the turbine performance especially the effi-ciency drop when facing partial admission conditions

Comparing the component performance shown inFigure 16 with the turbine efficiency the same conclu-sion as the steady-state condition was found wherebythe turbine performances are mainly determined by the

Figure 13 Static entropy distribution in the rotor passage(a) asymmetric (b) symmetric and (c) asymmetric OP

Figure 14 Exit flow field at the impeller trailing edge (a) asymmetric (b) symmetric and (c) asymmetric OP

Wang and Zheng 2253

rotor performance and the contribution of the flow lossin the volute is marginally negligible

Figure 17 shows the rotor efficiency and the radialvelocity distribution at the impeller leading edge at timesteps A and B Time step A corresponds to the extremepartial admission condition when all of the turbineshave the efficiency drop It can be seen that the radialvelocity distribution is almost the same for the threemodels at this time The most important factor is theefficiency distribution and it is clear that the OP modelhas the highest efficiency and that the asymmetric onehas the lowest which influences the turbineperformance

Similar to the results under steady conditions theefficiency distribution for the pulsating inlet conditionis affected by the incidence angle at the impeller inletposition especially in the positive incidence regionFigure 18 illustrates that from the asymmetric case tothe OP case the positive incidence region moves fromthe hub side to the shroud side As mentioned beforethe positive incidence angle can sever the vortex in the

impeller flow path and cause the efficiency to dropThe impeller hub side leading edge is more sensitive toa positive incidence angle than the shroud sideTherefore as the positive incidence region moves fromthe hub to the shroud side the efficiency recoverswhich can explain the performance difference in thesethree models

Under equal steady inlet conditions the mass flow inthe big scroll is larger than in the small scroll resultingin a faster flow velocity and positive incidence angleCombined with the conclusion that the hub side of theimpeller is more sensitive to positive incidence if thebig scroll is on the hub side the efficiency drops Underpulsating inlet conditions the efficiency is different tothe steady-state situation especially for the asymmetricturbine when the hub side is charged more than theshroud side This is because the big volute has a highermass flow storage ability than the small one that is thesmall scroll response is more sensitive to the pressurerise and flow in the small scroll is accelerated leadingto a higher outlet velocity which causes a greater inci-dence angle When the pressure rises for the asymmetricturbine with the small scroll on the hub side a positiveincidence angle region occurs at the hub side and theefficiency drops

Conclusion

Simulations for a symmetric and two opposite asym-metric twin scroll turbines under steady and pulsatinginlet conditions have been conducted to study the dif-ferences in the flow field and performance betweenthese three types of turbines an experiment was carriedout to validate the numerical methodology under steadyinlet conditions and the conclusions are as follows

1 Under steady admission conditions the asym-metric turbine with the big scroll on the shroud sidehas the best efficiency which is nearly 16 higherthan the contrary type (OP) Under steady full inletconditions the incidence angle at the region out-side the big scroll is positive The impeller hub sideleading edge is more sensitive to the positive inci-dence than the shroud side due to the hub side vor-tex in the impeller flow path while the positiveincidence severs the vortex and causes the efficiencyto drop Therefore if the big scroll is on the hubside the efficiency is worse than on the shroudside

2 Under pulsating inlet conditions the results arecontrary to the steady state The asymmetric tur-bine with the big scroll on the hub side (OP) hasthe best efficiency which is nearly 11 higherthan on the shroud side Since the small scroll massflow capacity is smaller than the big scroll whenthe backpressure increases the flow in the smallscroll is faster than in the big scroll which maycause a positive incidence angle at the impeller

Figure 15 Turbine performance under the 25-Hz pulsatinginlet condition

Figure 16 Component performance under the 25-Hzpulsating inlet condition

2254 Proc IMechE Part D J Automobile Engineering 233(8)

leading edge Therefore if the small scroll is set onthe hub side the positive incidence appears at thehub side and the efficiency is degraded

3 The design criterion for the asymmetric twin-entryturbine state is as follows when under steady inletconditions (eg constant-pressure turbochargingsystem on the diesel engine) it is better to place thebig scroll on the shroud side When under pulsat-ing inlet conditions (eg pulse turbocharging sys-tem on the gasoline engine) it is better to place thebig scroll on the hub side

4 Simulation calculations were conducted for a givenrotational speed and a given simplified scroll inputpressure It must be controlled whether these con-clusions remain the same for real engine pressureconditions and other engines speeds Experimentswill have to be carried out to check these results

Declaration of conflicting interests

The author(s) declared no potential conflicts of interestwith respect to the research authorship andor publica-tion of this article

Funding

The author(s) received no financial support for theresearch authorship andor publication of this article

Figure 17 Rotor efficiency and radial velocity distribution at the impeller leading edge (a) A time step and (b) B time step

Figure 18 Incidence angle distribution at the impeller leadingedge (a) A time step and (b) B time step

Wang and Zheng 2255

References

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2 Lapuerta M Hernandez JJ and Gimenez F Evaluationof exhaust gas recirculation as a technique for reducingdiesel engine NOX emissions Proc IMechE Part D J

Automobile Engineering 2000 214 85ndash933 Zamboni G Moggia S and Capobianco M Hybrid EGR

and turbocharging systems control for low NOX and fuelconsumption in an automotive diesel engine Appl Energ2016 165 839ndash848

4 Moustapha H Zelesky MF Baines NC et al Axial andradial turbines vol 2 White River Junction VT Con-cepts NREC 2003

5 Wijetunge RS Hawley JG and Vaughan ND Anexhaust pressure control strategy for a diesel engine ProcIMechE Part D J Automobile Engineering 2004 218

449ndash4646 Zhu D and Zheng X Asymmetric twin-scroll turbochar-

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7 Schmidt S Rose MG Muller M et al Variable asym-metric turbine for heavy duty truck engines In ASME

turbo expo 2013 turbine technical conference and expositionSan Antonio TX 3ndash7 June 2013 ppV05AT23A013-1ndashV05AT23A013-11 New York ASME

8 Brinkert N Sumser S Weber S et al Understanding thetwin scroll turbine flow similarity J Turbomach 2013135 021039

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Mechanical Engineers 3rd International Conference on

Turbocharging and Turbochargers London May 6ndash81986 Paper No C11086 pp 65ndash76 Mechanical Engi-neering Publications

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gas turbine conference and products show WashingtonDC 28 Februaryndash4 March 1965 New York ASME

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flow turbines under unsteady flow conditions with full

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110517 Woods WA and Norbury JF Pulse flow performance of

axial turbines for marine turbochargers In ISME Tokyo

1973 Tokyo The Marine Engineering Society of Japan

pp1ndash518 Galindo J Fajardo P Navarro R et al Characterization

of a radial turbocharger turbine in pulsating flow by

means of CFD and its application to engine modeling

Appl Energ 2013 103 116ndash12719 Costall AW McDavid RM Martinez-Botas RF et al

Pulse performance modeling of a twin entry turbocharger

turbine under full and unequal admission J Turbomach

2011 133 02100520 Mamat AMB and Martinez-Botas RF Mean line flow

model of steady and pulsating flow of a mixed-flow tur-

bine turbocharger In ASME turbo expo 2010 power for

land sea and air Glasgow 14ndash18 June 2010 pp2393ndash

2404 New York ASME

21 Rajoo S Romagnoli A andMartinez-Botas RF Unsteady

performance analysis of a twin-entry variable geometry

turbocharger turbine Energy 2012 38 176ndash18922 De Bellis V Marelli S Bozza F et al 1D simulation and

experimental analysis of a turbocharger turbine for auto-

motive engines under steady and unsteady flow condi-

tions Energ Proced 2014 45 909ndash91823 Moustapha SH Kacker SC and Tremblay B An

improved incidence losses prediction method for turbine

airfoils In ASME international gas turbine and aeroengine

congress and exposition Toronto ON Canada 4ndash8 June

1989 ppV001T01A100-1ndashV001T01A100-10 New York

ASME

24 Baines N C (1998) A meanline prediction method for

radial turbine efficiency In Proceedings of the Institution

of Mechanical Engineers 6th International Conference on

Turbocharging and Air Management Systems London

November 3ndash5 Mechanical Engineering Publications

Paper No C554ndash6 pp 315ndash325

Appendix 1

Notation

K total pressure loss coefficientm mass flow rate (kgs)P pressure (Pa)T temperature (K)

g heat ratioh efficiency

Subscripts

0 total parameter1 volute inlet2 volute outlet3 impeller outleth hub sides shroud side

2256 Proc IMechE Part D J Automobile Engineering 233(8)

popular type of asymmetric turbine shown in Figure 1may not be suitable for pulsating turbocharger systemso it is important to investigate the effect of the asym-metric volute on turbine performance and find outwhich type of arrangement can better suit pulsatinginlet conditions Unfortunately there is relatively lessresearch on asymmetric turbine under pulsating inletcondition and even rare study on another type of asym-metric volute that places the small scroll on the shroudside

In the present paper the authors compared the per-formance and flow field of two different types of asym-metric turbines together with another symmetricturbine under steady and pulsating inlet conditionsBased on the different performances of the three tur-bines under different inlet conditions a brief explana-tion is given and a design criterion for the asymmetrictwin-entry turbine when under steady or pulsating inletconditions is described

Test facility

The turbine dynamometer test rig on which the turbinecan be tested in isolation from the rest of the turbochar-ger is shown in Figure 2 There are four pressure sensorsat the turbine inlet the arrangement is equispaced at thecircumferential direction and the measurement precisionis 004 The number of inlet temperature sensors is 3and they have the same arrangement the precision is025C At the turbine outlet the arrangement of thepressure sensors is same as inlet condition but there isonly one temperature sensor The hydraulic dynam-ometer is used to absorb the power output and the for-mula to calculate the turbine efficiency ht is as follows

ht =Pt

mf3Cp3DT0eth1THORN

where Pt is the turbine output power that is tested bydynamometer and mf is mass flow rate DT0 is the

difference between the total temperature of the inletT01 and the outlet T03

For further explanation the parameter that thedynamometer can measure directly is the torque so Pt

is defined as follows

Pt =Tq3n

9550eth2THORN

where Tq is the torque and n is the rotational speedWhatrsquos more considering the inaccuracy of the tem-perature measurement caused by the uniformity of theturbine outlet condition the value of the outlet totaltemperature T03 is calculated as follows

T03 =T01

P01

P03

g

g1eth3THORN

where P01 and P03 are the inlet and outlet total pres-sures respectively As the static inlet temperature of thetest rig is controlled under 40C heat exchange withthe ambient environment can be neglected and the flowcan be considered as isentropic

Numerical methodology

Numerical model

There are three types of models used to analyze thesymmetric and asymmetric twin-entry scroll turbinesThe turbine geometry models which are shown inFigure 3 consist of the volute and the impeller Thedetailed parameters of the geometry are shown inTable 1 The three models use the same impeller butare combined with different volutes The asymmetricdesign refers to that where the throat area of the voluteshroud side scroll is bigger than that of the hub sideand the asymmetric OP is the model with the oppositefeatures as the asymmetric design The total throat areaof these three models remains the same and they alsohave the same volute cross-sectional area distributionfeature to ensure a similar flow capacity

The prototype of the asymmetric turbine is a com-mercial turbine installed at a 11-L inline six-cylinderdiesel engine The six cylinders are connected in twogroups of three and each group feeds one inlet of the

Figure 2 Sketch map of the test benchFigure 3 Turbine geometry models (a) asymmetric(b) symmetric and (c) asymmetric OP

2248 Proc IMechE Part D J Automobile Engineering 233(8)

twin-entry turbine This paper focuses on the differentperformance of the asymmetric and asymmetric OP tur-bine the symmetric type works as a reference

Considering the nonuniform circumferential flowfield caused by the volute at the impeller inlet it is nec-essary to simulate the entire impeller flow domain asopposed to the single passage domain The mesh of theimpeller and volute are shown in Figure 4 The struc-tured mesh is used for the impeller with a total of4360760 grids points The computational domain ofthe volute adopted the unstructured mesh with a totalof 2037995 grids points Therefore the total gridnumber of the model is more than 6 million

Numerical method

The simulation was conducted with the ANSYS CFXsolver which is based on the three-dimensional com-pressible finite volume scheme to solve steady-stateReynolds-averaged NavierndashStokes equations with theconservative formulation Considering the calculationcost and time the kndashe turbulence model is selected andcan accurately simulate the flow separation and vor-tices The validation of the mesh independence is shownin Figure 5 and as seen 5 or 6 million grids points canprovide reasonably accurate numerical results

Simulation procedure

The absolute total pressure and total temperature areimposed as the inlet boundary conditions the absolute

static pressure is imposed as the outlet boundary condi-tion and the data are derived from the experimentalmeasurements Inlet1 and inlet2 employ the boundaryconditions separately to make different inlet conditionsfor each scroll when under the pulsating inflow condi-tion At solid boundaries as the turbine test rig adoptsthe cold inlet air and there is heat insulator outside thepipe adiabatic wall conditions are imposed for thesimulation which means no slip and no heat transferThe frozen rotor model is imposed at the interface

Simulation validation

The results comparing the simulated total-to-staticisentropic efficiency and the mass flow rate under fulladmission conditions with the experiments are shownin Figure 6 The simulated efficiency is defined asfollows

hts =T01 T03

T01 1 P3

P01

(g1)=g eth4THORN

where T01 and T03 are the inlet and outlet total tem-peratures respectively P3 and P01 are the static outletpressure and total inlet pressure respectively

The experiments concern the asymmetric turbine(a) and the four lines refer to the rotational speed64000 80000 96000 and 112000 rmin respectively(Figure 3) The results indicate that the simulation is inreasonable agreement with the experimental data bothfor the efficiency and for the flow capacity with a maxi-mum discrepancy of 09 The noticeable differencesare toward the lower speed and lower expansion ratiorange of the efficiency This may be a result of theexperiment deviation because it is difficult to maintainthe test speed at the required value at lower rotatingspeeds and the deviation would be more significantthan at higher rotating speeds The good agreement ofthe flow capacity can also prove the conclusion becausethe test of the mass flow rate is negligibly affected bythe change in speed

Table 1 Turbine geometric parameters

Parameter Value

Blade numbers 10Volute outlet radius 4280 mmVolute outlet width 1138 mmImpeller inlet radius 4175 mmImpeller outlet radius 2766 mmTip clearance 045 mmInlet blade angle 0Throat area 1168 mm2

Figure 4 Single passage of the impeller mesh (a) and thedomain and boundaries (b)

Figure 5 The grid size independence validation

Wang and Zheng 2249

The CFD results and experiment data can matchwell with each other under steady inlet conditions andexperiments under pulsating inlet conditions will becarried out in the future to fine down the validation ofthe numerical model

In this paper first the performance and flow fieldsof the three types of turbines are compared understeady inlet conditions to determine the mechanism ofthe efficiency change and the geometric effects on theturbine performance Then the performance and flowfields under pulsating inlet conditions are investigatedto discover the pulse effects Finally based on theresults and analysis a criterion is established for theasymmetric twin-entry turbine design under steady andpulsating inlet conditions

Results under steady equal inlet conditions

The numerical results of the three simulated models areshown in Figure 7 The volute throat area was designedto maintain the same flow capacity in the three modelsas shown in the mass flow rate map It can be seen thatat low speeds the mass flow rates of the three turbine

types are nearly the same but at the highest speedthere was a drop in the mass flow rate of approximately02 which may cause a drop in the efficiency of thesymmetric turbine at high speeds putting it close to theasymmetric OP model Over the entire range of work-ing conditions the asymmetric turbine has the best tur-bine efficiency while the OP has the worst with thesymmetric turbine in the middle What is interesting isthat the efficiency drops between these three models atintervals of approximately 08 Based on this findingwe take the speed line 96000 rmin condition for fur-ther investigation to simplify the research process

Figure 8 shows the turbine and component perfor-mance comparison results It can be seen that the flowloss in the volute remains nearly constant with theexpansion ratio while the rotor efficiency shows thesame trend as turbine efficiency Although the sym-metric turbine has the lowest total pressure loss coeffi-cient this contributes little to the turbine performanceand considering that the test data used for the calcula-tions are measured at the interface the mixing flow atthe volute outlet domain may contribute to the rise inthe asymmetric and OPrsquos loss coefficients It is clear

Figure 6 Performance comparison between the numerical andexperimental results (a) efficiency and (b) mass flow rate

Figure 7 Performance comparison between the symmetricand the two different asymmetric turbines (a) efficiency and(b) mass flow rate

2250 Proc IMechE Part D J Automobile Engineering 233(8)

that the turbine efficiency is mainly determined by therotor efficiency and that the flow loss in the volute hasa negligible effect The main function of the volute is toprovide the rotor inlet flow field which may affect therotor efficiency Therefore even if the volute has differ-ent geometries the key point to determine the turbineefficiency is the flow field at the impeller inlet position

Figure 9 shows the shroud to hub spanwise distribu-tion for the rotor efficiency and radial velocity at theimpeller leading edge The rotor efficiency distributionis a hypothetical value that signifies the specific poweroutput that a unit mass of gas can provide if it entersthe impeller at a given spanwise position The valuesare calculated along the spanwise direction at the lead-ing edge and the corresponding trailing edge The radialvelocity represents the inlet mass flow and the rotorefficiency distribution and the match with the massflow distribution determine the impeller efficiencyThus if the high radial velocity position matches wellwith the high-efficiency region a senior rotor efficiencyis achieved

In Figure 9 the efficiency is higher at the hub sidethan at the shroud side and the asymmetric case hasthe highest efficiency while the OP has the lowest Forthe mass flow distribution for the asymmetric casesymmetric case and OP case the maximum mass flowposition moves from the hub side to the shroud side orfrom the high-efficiency region to the low-efficiencyregion These two reasons can explain the differences inthe rotor efficiency

Figure 10 shows the rotor inlet incidence angle ofthe three turbine models along the plane at 001mm

upstream of the rotor leading edge which is the criticalparameter to evaluate the rotor performance It isrecognized that optimum incidence is in the region of220 to 240 and that a deviation from this region cancause a drop in the efficiency Moustapha et al23 con-cluded that the radial turbine rotors are much moresensitive to positive incidence than to extremes of nega-tive incidence Then Baines24 put forward a postulatethat the rotor is more sensitive to the positive incidencenear the hub surface to explain the difference in the effi-ciencies under the hub and shroud side partial admis-sion condition Nevertheless he did not show anyexperimental or CFD results to support the postulate

It can be seen from Figure 10 that the positive inci-dence appears in the region that corresponds to theposition of the large volute scroll From Figure 10(a) to(c) the positive incidence region moves from the shroudside to the hub side of the impeller causing the effi-ciency to decrease These results correspond to Bainesrsquohypothesis

From the flow field at the rotor inlet region shown inFigure 11 a vortex exists caused by the flow separationat the hub side near the suction surface of the impellerCompared to the static entropy distribution it can beseen that there are two high entropy regions in the flowchannel one of which is near the hub and correspondsto the vortex The other is near the shroud which iscaused by the tip clearance flow Also the entropy gaincaused by the vortex is larger than the tip clearanceentropy gains

Figure 8 Turbine and component performance Figure 9 Rotor efficiency and radial velocity distribution at theimpeller leading edge

Wang and Zheng 2251

Moreover the vortex region is also where the highstatic entropy appears It is well known that a negative

incidence angle can weaken the flow separation whilea positive angle can aggravate it Therefore when thelarge volute scroll is placed on the shroud side the neg-ative incidence region is at the hub side of the impellerThis can weaken the flow separation fade or even erasethe vortex which may improve the turbine perfor-mance When the larger volute scroll is placed on thehub side the positive incidence region is at the hubside which may enhance the degree of the flow separa-tion and the vortex could become severe leading to thedrop in efficiency This is illustrated in the flow field inFigure 12

The static entropy distribution from Figure 13 at thevortex region shows the same conclusion It can be seenthat the asymmetric case only has an entropy gaincaused by tip clearance while the asymmetric OP modelhas the most severe entropy gain at the vortex regionThis can explain the efficiency differences between thesethree models

Figure 14 shows the velocity distribution at thevolute outlet plane at the impeller trailing edge It illus-trates that these three models have nearly the same out-let flow field and their mean velocities are 1403 1412and 1421ms respectively This means that the effi-ciency differences of the impellers are mainly caused bythe difference in the inlet flow condition

Results under the pulsating inlet condition

When the engine adopted the pulsating turbochargersystem the exhaust pressure is a pulse wave with a spe-cific amplitude and a specific main frequency and thefrequency is connected with the engine speed We canuse an ideal sinusoidal signal to represent the realengine exhaust pressure pulse wave as the turbine pul-sating inlet condition This paper has simplified the

Figure 10 Incidence angle distribution at the impeller leadingedge (a) asymmetric (b) symmetric and (c) asymmetric OP

Figure 11 Flow field and static entropy at the impeller inletregion

Figure 12 Flow field in the rotor passage (a) asymmetric (b) symmetric and (c) asymmetric OP

2252 Proc IMechE Part D J Automobile Engineering 233(8)

pulsating inlet condition to a sinusoidal signal whilekept the amplitude and frequency similar to the enginereal exhaust pressure and the prototype pressure waveis from a six-cylinder diesel engine in Mualler et al13

The pulsating frequency of the simulation is 25Hzcorresponding to the power output condition for this 11-L diesel engines This working point demand is morestringent to fuel consumption which means that a highturbine efficiency is required During pulsating inlet con-ditions the volute shroud side scroll and the hub sidescroll face two different inlet pulses that have a reversephase as shown in Figure 15 Considering the two isola-tion inlet conditions of the volute the isentropic efficiencyof the turbine at a specific time is defined as follows

h=m1s +m1h m1s +m1heth THORN T03=T01

m1s 1 (P3=P01s)g1

g

+m1h 1 (P3=P01h)

g1g

eth5THORN

Then the volute total pressure loss coefficient isnewly defined as follows

K=m1sP01s +m1hP01h (m1s +m1h)P02

(m1s +m1h)(P02 P2)eth6THORN

As for rotor efficiency considering that the form ofthe inlet and outlet remains the same as in the steadysituation the calculation remains unchanged

Figure 15 shows the performance of these three tur-bines under a 25-Hz pulsating inlet flow condition The

time-averaged efficiencies for these three turbines are7448 7469 and 7554 corresponding to theasymmetric symmetric and asymmetric OP modelrespectively The asymmetric type has the worst perfor-mance which is a contrary result of the steady perfor-mance where the asymmetric turbine shows the bestperformance This means that the performances of theturbines are different when under steady and pulsatinginlet flow conditions

Under the pulsating inlet conditions the highest effi-ciency points appear for all three models when the huband shroud inlet pressures are equal The efficiencydrops when facing the partial admission conditionespecially in the situation when the hub side inlet hasthe highest back-pressure

When the volute hub side inlet has the lower pres-sure corresponding to less mass flow the efficiencies ofthese three models are nearly the same and the asym-metric turbine is improved over the other twoHowever when the hub side has more pressure theefficiency drops and the asymmetric turbine is affectedthe most causing a decrease in the average efficiencyTherefore the inlet condition at the volute hub side isthe main factor to control the twin-entry turbine per-formance and the decrease in the mass flow may notcause a noticeable rise in the efficiency but the increasein mass flow will lead to an efficiency drop If the twin-entry turbine is symmetric the high efficiency appearswhen the shroud side inlet mass flow rate is larger thanthe hub side and drops when the mass flow is lesswhich agrees with previous conclusions under steadyinlet conditions from the literature

The asymmetric model that places the small scrollon the hub side is affected more by the hub side inletand the OP model that places the big scroll on the hubside is affected less Therefore the small scroll has moreinfluence on the turbine performance In summary thesmall scroll on the hub side is the main cause of deter-mining the turbine performance especially the effi-ciency drop when facing partial admission conditions

Comparing the component performance shown inFigure 16 with the turbine efficiency the same conclu-sion as the steady-state condition was found wherebythe turbine performances are mainly determined by the

Figure 13 Static entropy distribution in the rotor passage(a) asymmetric (b) symmetric and (c) asymmetric OP

Figure 14 Exit flow field at the impeller trailing edge (a) asymmetric (b) symmetric and (c) asymmetric OP

Wang and Zheng 2253

rotor performance and the contribution of the flow lossin the volute is marginally negligible

Figure 17 shows the rotor efficiency and the radialvelocity distribution at the impeller leading edge at timesteps A and B Time step A corresponds to the extremepartial admission condition when all of the turbineshave the efficiency drop It can be seen that the radialvelocity distribution is almost the same for the threemodels at this time The most important factor is theefficiency distribution and it is clear that the OP modelhas the highest efficiency and that the asymmetric onehas the lowest which influences the turbineperformance

Similar to the results under steady conditions theefficiency distribution for the pulsating inlet conditionis affected by the incidence angle at the impeller inletposition especially in the positive incidence regionFigure 18 illustrates that from the asymmetric case tothe OP case the positive incidence region moves fromthe hub side to the shroud side As mentioned beforethe positive incidence angle can sever the vortex in the

impeller flow path and cause the efficiency to dropThe impeller hub side leading edge is more sensitive toa positive incidence angle than the shroud sideTherefore as the positive incidence region moves fromthe hub to the shroud side the efficiency recoverswhich can explain the performance difference in thesethree models

Under equal steady inlet conditions the mass flow inthe big scroll is larger than in the small scroll resultingin a faster flow velocity and positive incidence angleCombined with the conclusion that the hub side of theimpeller is more sensitive to positive incidence if thebig scroll is on the hub side the efficiency drops Underpulsating inlet conditions the efficiency is different tothe steady-state situation especially for the asymmetricturbine when the hub side is charged more than theshroud side This is because the big volute has a highermass flow storage ability than the small one that is thesmall scroll response is more sensitive to the pressurerise and flow in the small scroll is accelerated leadingto a higher outlet velocity which causes a greater inci-dence angle When the pressure rises for the asymmetricturbine with the small scroll on the hub side a positiveincidence angle region occurs at the hub side and theefficiency drops

Conclusion

Simulations for a symmetric and two opposite asym-metric twin scroll turbines under steady and pulsatinginlet conditions have been conducted to study the dif-ferences in the flow field and performance betweenthese three types of turbines an experiment was carriedout to validate the numerical methodology under steadyinlet conditions and the conclusions are as follows

1 Under steady admission conditions the asym-metric turbine with the big scroll on the shroud sidehas the best efficiency which is nearly 16 higherthan the contrary type (OP) Under steady full inletconditions the incidence angle at the region out-side the big scroll is positive The impeller hub sideleading edge is more sensitive to the positive inci-dence than the shroud side due to the hub side vor-tex in the impeller flow path while the positiveincidence severs the vortex and causes the efficiencyto drop Therefore if the big scroll is on the hubside the efficiency is worse than on the shroudside

2 Under pulsating inlet conditions the results arecontrary to the steady state The asymmetric tur-bine with the big scroll on the hub side (OP) hasthe best efficiency which is nearly 11 higherthan on the shroud side Since the small scroll massflow capacity is smaller than the big scroll whenthe backpressure increases the flow in the smallscroll is faster than in the big scroll which maycause a positive incidence angle at the impeller

Figure 15 Turbine performance under the 25-Hz pulsatinginlet condition

Figure 16 Component performance under the 25-Hzpulsating inlet condition

2254 Proc IMechE Part D J Automobile Engineering 233(8)

leading edge Therefore if the small scroll is set onthe hub side the positive incidence appears at thehub side and the efficiency is degraded

3 The design criterion for the asymmetric twin-entryturbine state is as follows when under steady inletconditions (eg constant-pressure turbochargingsystem on the diesel engine) it is better to place thebig scroll on the shroud side When under pulsat-ing inlet conditions (eg pulse turbocharging sys-tem on the gasoline engine) it is better to place thebig scroll on the hub side

4 Simulation calculations were conducted for a givenrotational speed and a given simplified scroll inputpressure It must be controlled whether these con-clusions remain the same for real engine pressureconditions and other engines speeds Experimentswill have to be carried out to check these results

Declaration of conflicting interests

The author(s) declared no potential conflicts of interestwith respect to the research authorship andor publica-tion of this article

Funding

The author(s) received no financial support for theresearch authorship andor publication of this article

Figure 17 Rotor efficiency and radial velocity distribution at the impeller leading edge (a) A time step and (b) B time step

Figure 18 Incidence angle distribution at the impeller leadingedge (a) A time step and (b) B time step

Wang and Zheng 2255

References

1 Durnholz M Eifler G and Endres H Exhaust-gas recir-culation a measure to reduce exhaust emissions of DIdiesel engines SAE technical paper 920725 1992

2 Lapuerta M Hernandez JJ and Gimenez F Evaluationof exhaust gas recirculation as a technique for reducingdiesel engine NOX emissions Proc IMechE Part D J

Automobile Engineering 2000 214 85ndash933 Zamboni G Moggia S and Capobianco M Hybrid EGR

and turbocharging systems control for low NOX and fuelconsumption in an automotive diesel engine Appl Energ2016 165 839ndash848

4 Moustapha H Zelesky MF Baines NC et al Axial andradial turbines vol 2 White River Junction VT Con-cepts NREC 2003

5 Wijetunge RS Hawley JG and Vaughan ND Anexhaust pressure control strategy for a diesel engine ProcIMechE Part D J Automobile Engineering 2004 218

449ndash4646 Zhu D and Zheng X Asymmetric twin-scroll turbochar-

ging in diesel engines for energy and emission improve-ment Energy 2017 141 702ndash714

7 Schmidt S Rose MG Muller M et al Variable asym-metric turbine for heavy duty truck engines In ASME

turbo expo 2013 turbine technical conference and expositionSan Antonio TX 3ndash7 June 2013 ppV05AT23A013-1ndashV05AT23A013-11 New York ASME

8 Brinkert N Sumser S Weber S et al Understanding thetwin scroll turbine flow similarity J Turbomach 2013135 021039

9 Dale A and Watson N Vaneless radial turbochargerturbine performance In Proceedings of the Institution of

Mechanical Engineers 3rd International Conference on

Turbocharging and Turbochargers London May 6ndash81986 Paper No C11086 pp 65ndash76 Mechanical Engi-neering Publications

10 Baines NC and Lavy M Flows in vaned and vaneless sta-tors of radial inflow turbocharger turbines In IMechE

turbocharging and turbochargers conference LondonMay Paper No C405005 pp 7ndash12 Mechanical Engi-neering Publications

11 Baines NC and Yeo JH Flow in a radial turbine underequal and partial admission conditions In IMechE Eur-

opean conference on turbomachinery latest developments

in a changing scene London March 12ndash13 MechanicalEngineering Publications Paper No C423002 pp 103ndash112

12 Baines NC Hajilouy-Benisi A and Yeo JH The pulseflow performance and modelling of radial inflow tur-bines In IMechE conference publications 1994 6 209ndash220 London Medical Engineering Publications

13 Mualler M Streule T Sumser S et al The asymmetric

twin scroll turbine for exhaust gas turbochargers InASME turbo expo 2008 power for land sea and air Ber-lin 9ndash13 June 2008 pp1547ndash1554 New York ASME

14 Wallace FJ and Blair GP The pulsating-flow perfor-mance of inward radial-flow turbines In ASME 1965

gas turbine conference and products show WashingtonDC 28 Februaryndash4 March 1965 New York ASME

15 Wallace FJ Cave PR and Miles J Performance of inwardradial flow turbines under steady flow conditions withspecial reference to high pressure ratios and partialadmission Proc IMechE 1969 184 1027ndash1042

16 Wallace FJ and Miles J Performance of inward radial

flow turbines under unsteady flow conditions with full

and partial admission Proc IMechE 1970 185 1091ndash

110517 Woods WA and Norbury JF Pulse flow performance of

axial turbines for marine turbochargers In ISME Tokyo

1973 Tokyo The Marine Engineering Society of Japan

pp1ndash518 Galindo J Fajardo P Navarro R et al Characterization

of a radial turbocharger turbine in pulsating flow by

means of CFD and its application to engine modeling

Appl Energ 2013 103 116ndash12719 Costall AW McDavid RM Martinez-Botas RF et al

Pulse performance modeling of a twin entry turbocharger

turbine under full and unequal admission J Turbomach

2011 133 02100520 Mamat AMB and Martinez-Botas RF Mean line flow

model of steady and pulsating flow of a mixed-flow tur-

bine turbocharger In ASME turbo expo 2010 power for

land sea and air Glasgow 14ndash18 June 2010 pp2393ndash

2404 New York ASME

21 Rajoo S Romagnoli A andMartinez-Botas RF Unsteady

performance analysis of a twin-entry variable geometry

turbocharger turbine Energy 2012 38 176ndash18922 De Bellis V Marelli S Bozza F et al 1D simulation and

experimental analysis of a turbocharger turbine for auto-

motive engines under steady and unsteady flow condi-

tions Energ Proced 2014 45 909ndash91823 Moustapha SH Kacker SC and Tremblay B An

improved incidence losses prediction method for turbine

airfoils In ASME international gas turbine and aeroengine

congress and exposition Toronto ON Canada 4ndash8 June

1989 ppV001T01A100-1ndashV001T01A100-10 New York

ASME

24 Baines N C (1998) A meanline prediction method for

radial turbine efficiency In Proceedings of the Institution

of Mechanical Engineers 6th International Conference on

Turbocharging and Air Management Systems London

November 3ndash5 Mechanical Engineering Publications

Paper No C554ndash6 pp 315ndash325

Appendix 1

Notation

K total pressure loss coefficientm mass flow rate (kgs)P pressure (Pa)T temperature (K)

g heat ratioh efficiency

Subscripts

0 total parameter1 volute inlet2 volute outlet3 impeller outleth hub sides shroud side

2256 Proc IMechE Part D J Automobile Engineering 233(8)

twin-entry turbine This paper focuses on the differentperformance of the asymmetric and asymmetric OP tur-bine the symmetric type works as a reference

Considering the nonuniform circumferential flowfield caused by the volute at the impeller inlet it is nec-essary to simulate the entire impeller flow domain asopposed to the single passage domain The mesh of theimpeller and volute are shown in Figure 4 The struc-tured mesh is used for the impeller with a total of4360760 grids points The computational domain ofthe volute adopted the unstructured mesh with a totalof 2037995 grids points Therefore the total gridnumber of the model is more than 6 million

Numerical method

The simulation was conducted with the ANSYS CFXsolver which is based on the three-dimensional com-pressible finite volume scheme to solve steady-stateReynolds-averaged NavierndashStokes equations with theconservative formulation Considering the calculationcost and time the kndashe turbulence model is selected andcan accurately simulate the flow separation and vor-tices The validation of the mesh independence is shownin Figure 5 and as seen 5 or 6 million grids points canprovide reasonably accurate numerical results

Simulation procedure

The absolute total pressure and total temperature areimposed as the inlet boundary conditions the absolute

static pressure is imposed as the outlet boundary condi-tion and the data are derived from the experimentalmeasurements Inlet1 and inlet2 employ the boundaryconditions separately to make different inlet conditionsfor each scroll when under the pulsating inflow condi-tion At solid boundaries as the turbine test rig adoptsthe cold inlet air and there is heat insulator outside thepipe adiabatic wall conditions are imposed for thesimulation which means no slip and no heat transferThe frozen rotor model is imposed at the interface

Simulation validation

The results comparing the simulated total-to-staticisentropic efficiency and the mass flow rate under fulladmission conditions with the experiments are shownin Figure 6 The simulated efficiency is defined asfollows

hts =T01 T03

T01 1 P3

P01

(g1)=g eth4THORN

where T01 and T03 are the inlet and outlet total tem-peratures respectively P3 and P01 are the static outletpressure and total inlet pressure respectively

The experiments concern the asymmetric turbine(a) and the four lines refer to the rotational speed64000 80000 96000 and 112000 rmin respectively(Figure 3) The results indicate that the simulation is inreasonable agreement with the experimental data bothfor the efficiency and for the flow capacity with a maxi-mum discrepancy of 09 The noticeable differencesare toward the lower speed and lower expansion ratiorange of the efficiency This may be a result of theexperiment deviation because it is difficult to maintainthe test speed at the required value at lower rotatingspeeds and the deviation would be more significantthan at higher rotating speeds The good agreement ofthe flow capacity can also prove the conclusion becausethe test of the mass flow rate is negligibly affected bythe change in speed

Table 1 Turbine geometric parameters

Parameter Value

Blade numbers 10Volute outlet radius 4280 mmVolute outlet width 1138 mmImpeller inlet radius 4175 mmImpeller outlet radius 2766 mmTip clearance 045 mmInlet blade angle 0Throat area 1168 mm2

Figure 4 Single passage of the impeller mesh (a) and thedomain and boundaries (b)

Figure 5 The grid size independence validation

Wang and Zheng 2249

The CFD results and experiment data can matchwell with each other under steady inlet conditions andexperiments under pulsating inlet conditions will becarried out in the future to fine down the validation ofthe numerical model

In this paper first the performance and flow fieldsof the three types of turbines are compared understeady inlet conditions to determine the mechanism ofthe efficiency change and the geometric effects on theturbine performance Then the performance and flowfields under pulsating inlet conditions are investigatedto discover the pulse effects Finally based on theresults and analysis a criterion is established for theasymmetric twin-entry turbine design under steady andpulsating inlet conditions

Results under steady equal inlet conditions

The numerical results of the three simulated models areshown in Figure 7 The volute throat area was designedto maintain the same flow capacity in the three modelsas shown in the mass flow rate map It can be seen thatat low speeds the mass flow rates of the three turbine

types are nearly the same but at the highest speedthere was a drop in the mass flow rate of approximately02 which may cause a drop in the efficiency of thesymmetric turbine at high speeds putting it close to theasymmetric OP model Over the entire range of work-ing conditions the asymmetric turbine has the best tur-bine efficiency while the OP has the worst with thesymmetric turbine in the middle What is interesting isthat the efficiency drops between these three models atintervals of approximately 08 Based on this findingwe take the speed line 96000 rmin condition for fur-ther investigation to simplify the research process

Figure 8 shows the turbine and component perfor-mance comparison results It can be seen that the flowloss in the volute remains nearly constant with theexpansion ratio while the rotor efficiency shows thesame trend as turbine efficiency Although the sym-metric turbine has the lowest total pressure loss coeffi-cient this contributes little to the turbine performanceand considering that the test data used for the calcula-tions are measured at the interface the mixing flow atthe volute outlet domain may contribute to the rise inthe asymmetric and OPrsquos loss coefficients It is clear

Figure 6 Performance comparison between the numerical andexperimental results (a) efficiency and (b) mass flow rate

Figure 7 Performance comparison between the symmetricand the two different asymmetric turbines (a) efficiency and(b) mass flow rate

2250 Proc IMechE Part D J Automobile Engineering 233(8)

that the turbine efficiency is mainly determined by therotor efficiency and that the flow loss in the volute hasa negligible effect The main function of the volute is toprovide the rotor inlet flow field which may affect therotor efficiency Therefore even if the volute has differ-ent geometries the key point to determine the turbineefficiency is the flow field at the impeller inlet position

Figure 9 shows the shroud to hub spanwise distribu-tion for the rotor efficiency and radial velocity at theimpeller leading edge The rotor efficiency distributionis a hypothetical value that signifies the specific poweroutput that a unit mass of gas can provide if it entersthe impeller at a given spanwise position The valuesare calculated along the spanwise direction at the lead-ing edge and the corresponding trailing edge The radialvelocity represents the inlet mass flow and the rotorefficiency distribution and the match with the massflow distribution determine the impeller efficiencyThus if the high radial velocity position matches wellwith the high-efficiency region a senior rotor efficiencyis achieved

In Figure 9 the efficiency is higher at the hub sidethan at the shroud side and the asymmetric case hasthe highest efficiency while the OP has the lowest Forthe mass flow distribution for the asymmetric casesymmetric case and OP case the maximum mass flowposition moves from the hub side to the shroud side orfrom the high-efficiency region to the low-efficiencyregion These two reasons can explain the differences inthe rotor efficiency

Figure 10 shows the rotor inlet incidence angle ofthe three turbine models along the plane at 001mm

upstream of the rotor leading edge which is the criticalparameter to evaluate the rotor performance It isrecognized that optimum incidence is in the region of220 to 240 and that a deviation from this region cancause a drop in the efficiency Moustapha et al23 con-cluded that the radial turbine rotors are much moresensitive to positive incidence than to extremes of nega-tive incidence Then Baines24 put forward a postulatethat the rotor is more sensitive to the positive incidencenear the hub surface to explain the difference in the effi-ciencies under the hub and shroud side partial admis-sion condition Nevertheless he did not show anyexperimental or CFD results to support the postulate

It can be seen from Figure 10 that the positive inci-dence appears in the region that corresponds to theposition of the large volute scroll From Figure 10(a) to(c) the positive incidence region moves from the shroudside to the hub side of the impeller causing the effi-ciency to decrease These results correspond to Bainesrsquohypothesis

From the flow field at the rotor inlet region shown inFigure 11 a vortex exists caused by the flow separationat the hub side near the suction surface of the impellerCompared to the static entropy distribution it can beseen that there are two high entropy regions in the flowchannel one of which is near the hub and correspondsto the vortex The other is near the shroud which iscaused by the tip clearance flow Also the entropy gaincaused by the vortex is larger than the tip clearanceentropy gains

Figure 8 Turbine and component performance Figure 9 Rotor efficiency and radial velocity distribution at theimpeller leading edge

Wang and Zheng 2251

Moreover the vortex region is also where the highstatic entropy appears It is well known that a negative

incidence angle can weaken the flow separation whilea positive angle can aggravate it Therefore when thelarge volute scroll is placed on the shroud side the neg-ative incidence region is at the hub side of the impellerThis can weaken the flow separation fade or even erasethe vortex which may improve the turbine perfor-mance When the larger volute scroll is placed on thehub side the positive incidence region is at the hubside which may enhance the degree of the flow separa-tion and the vortex could become severe leading to thedrop in efficiency This is illustrated in the flow field inFigure 12

The static entropy distribution from Figure 13 at thevortex region shows the same conclusion It can be seenthat the asymmetric case only has an entropy gaincaused by tip clearance while the asymmetric OP modelhas the most severe entropy gain at the vortex regionThis can explain the efficiency differences between thesethree models

Figure 14 shows the velocity distribution at thevolute outlet plane at the impeller trailing edge It illus-trates that these three models have nearly the same out-let flow field and their mean velocities are 1403 1412and 1421ms respectively This means that the effi-ciency differences of the impellers are mainly caused bythe difference in the inlet flow condition

Results under the pulsating inlet condition

When the engine adopted the pulsating turbochargersystem the exhaust pressure is a pulse wave with a spe-cific amplitude and a specific main frequency and thefrequency is connected with the engine speed We canuse an ideal sinusoidal signal to represent the realengine exhaust pressure pulse wave as the turbine pul-sating inlet condition This paper has simplified the

Figure 10 Incidence angle distribution at the impeller leadingedge (a) asymmetric (b) symmetric and (c) asymmetric OP

Figure 11 Flow field and static entropy at the impeller inletregion

Figure 12 Flow field in the rotor passage (a) asymmetric (b) symmetric and (c) asymmetric OP

2252 Proc IMechE Part D J Automobile Engineering 233(8)

pulsating inlet condition to a sinusoidal signal whilekept the amplitude and frequency similar to the enginereal exhaust pressure and the prototype pressure waveis from a six-cylinder diesel engine in Mualler et al13

The pulsating frequency of the simulation is 25Hzcorresponding to the power output condition for this 11-L diesel engines This working point demand is morestringent to fuel consumption which means that a highturbine efficiency is required During pulsating inlet con-ditions the volute shroud side scroll and the hub sidescroll face two different inlet pulses that have a reversephase as shown in Figure 15 Considering the two isola-tion inlet conditions of the volute the isentropic efficiencyof the turbine at a specific time is defined as follows

h=m1s +m1h m1s +m1heth THORN T03=T01

m1s 1 (P3=P01s)g1

g

+m1h 1 (P3=P01h)

g1g

eth5THORN

Then the volute total pressure loss coefficient isnewly defined as follows

K=m1sP01s +m1hP01h (m1s +m1h)P02

(m1s +m1h)(P02 P2)eth6THORN

As for rotor efficiency considering that the form ofthe inlet and outlet remains the same as in the steadysituation the calculation remains unchanged

Figure 15 shows the performance of these three tur-bines under a 25-Hz pulsating inlet flow condition The

time-averaged efficiencies for these three turbines are7448 7469 and 7554 corresponding to theasymmetric symmetric and asymmetric OP modelrespectively The asymmetric type has the worst perfor-mance which is a contrary result of the steady perfor-mance where the asymmetric turbine shows the bestperformance This means that the performances of theturbines are different when under steady and pulsatinginlet flow conditions

Under the pulsating inlet conditions the highest effi-ciency points appear for all three models when the huband shroud inlet pressures are equal The efficiencydrops when facing the partial admission conditionespecially in the situation when the hub side inlet hasthe highest back-pressure

When the volute hub side inlet has the lower pres-sure corresponding to less mass flow the efficiencies ofthese three models are nearly the same and the asym-metric turbine is improved over the other twoHowever when the hub side has more pressure theefficiency drops and the asymmetric turbine is affectedthe most causing a decrease in the average efficiencyTherefore the inlet condition at the volute hub side isthe main factor to control the twin-entry turbine per-formance and the decrease in the mass flow may notcause a noticeable rise in the efficiency but the increasein mass flow will lead to an efficiency drop If the twin-entry turbine is symmetric the high efficiency appearswhen the shroud side inlet mass flow rate is larger thanthe hub side and drops when the mass flow is lesswhich agrees with previous conclusions under steadyinlet conditions from the literature

The asymmetric model that places the small scrollon the hub side is affected more by the hub side inletand the OP model that places the big scroll on the hubside is affected less Therefore the small scroll has moreinfluence on the turbine performance In summary thesmall scroll on the hub side is the main cause of deter-mining the turbine performance especially the effi-ciency drop when facing partial admission conditions

Comparing the component performance shown inFigure 16 with the turbine efficiency the same conclu-sion as the steady-state condition was found wherebythe turbine performances are mainly determined by the

Figure 13 Static entropy distribution in the rotor passage(a) asymmetric (b) symmetric and (c) asymmetric OP

Figure 14 Exit flow field at the impeller trailing edge (a) asymmetric (b) symmetric and (c) asymmetric OP

Wang and Zheng 2253

rotor performance and the contribution of the flow lossin the volute is marginally negligible

Figure 17 shows the rotor efficiency and the radialvelocity distribution at the impeller leading edge at timesteps A and B Time step A corresponds to the extremepartial admission condition when all of the turbineshave the efficiency drop It can be seen that the radialvelocity distribution is almost the same for the threemodels at this time The most important factor is theefficiency distribution and it is clear that the OP modelhas the highest efficiency and that the asymmetric onehas the lowest which influences the turbineperformance

Similar to the results under steady conditions theefficiency distribution for the pulsating inlet conditionis affected by the incidence angle at the impeller inletposition especially in the positive incidence regionFigure 18 illustrates that from the asymmetric case tothe OP case the positive incidence region moves fromthe hub side to the shroud side As mentioned beforethe positive incidence angle can sever the vortex in the

impeller flow path and cause the efficiency to dropThe impeller hub side leading edge is more sensitive toa positive incidence angle than the shroud sideTherefore as the positive incidence region moves fromthe hub to the shroud side the efficiency recoverswhich can explain the performance difference in thesethree models

Under equal steady inlet conditions the mass flow inthe big scroll is larger than in the small scroll resultingin a faster flow velocity and positive incidence angleCombined with the conclusion that the hub side of theimpeller is more sensitive to positive incidence if thebig scroll is on the hub side the efficiency drops Underpulsating inlet conditions the efficiency is different tothe steady-state situation especially for the asymmetricturbine when the hub side is charged more than theshroud side This is because the big volute has a highermass flow storage ability than the small one that is thesmall scroll response is more sensitive to the pressurerise and flow in the small scroll is accelerated leadingto a higher outlet velocity which causes a greater inci-dence angle When the pressure rises for the asymmetricturbine with the small scroll on the hub side a positiveincidence angle region occurs at the hub side and theefficiency drops

Conclusion

Simulations for a symmetric and two opposite asym-metric twin scroll turbines under steady and pulsatinginlet conditions have been conducted to study the dif-ferences in the flow field and performance betweenthese three types of turbines an experiment was carriedout to validate the numerical methodology under steadyinlet conditions and the conclusions are as follows

1 Under steady admission conditions the asym-metric turbine with the big scroll on the shroud sidehas the best efficiency which is nearly 16 higherthan the contrary type (OP) Under steady full inletconditions the incidence angle at the region out-side the big scroll is positive The impeller hub sideleading edge is more sensitive to the positive inci-dence than the shroud side due to the hub side vor-tex in the impeller flow path while the positiveincidence severs the vortex and causes the efficiencyto drop Therefore if the big scroll is on the hubside the efficiency is worse than on the shroudside

2 Under pulsating inlet conditions the results arecontrary to the steady state The asymmetric tur-bine with the big scroll on the hub side (OP) hasthe best efficiency which is nearly 11 higherthan on the shroud side Since the small scroll massflow capacity is smaller than the big scroll whenthe backpressure increases the flow in the smallscroll is faster than in the big scroll which maycause a positive incidence angle at the impeller

Figure 15 Turbine performance under the 25-Hz pulsatinginlet condition

Figure 16 Component performance under the 25-Hzpulsating inlet condition

2254 Proc IMechE Part D J Automobile Engineering 233(8)

leading edge Therefore if the small scroll is set onthe hub side the positive incidence appears at thehub side and the efficiency is degraded

3 The design criterion for the asymmetric twin-entryturbine state is as follows when under steady inletconditions (eg constant-pressure turbochargingsystem on the diesel engine) it is better to place thebig scroll on the shroud side When under pulsat-ing inlet conditions (eg pulse turbocharging sys-tem on the gasoline engine) it is better to place thebig scroll on the hub side

4 Simulation calculations were conducted for a givenrotational speed and a given simplified scroll inputpressure It must be controlled whether these con-clusions remain the same for real engine pressureconditions and other engines speeds Experimentswill have to be carried out to check these results

Declaration of conflicting interests

The author(s) declared no potential conflicts of interestwith respect to the research authorship andor publica-tion of this article

Funding

The author(s) received no financial support for theresearch authorship andor publication of this article

Figure 17 Rotor efficiency and radial velocity distribution at the impeller leading edge (a) A time step and (b) B time step

Figure 18 Incidence angle distribution at the impeller leadingedge (a) A time step and (b) B time step

Wang and Zheng 2255

References

1 Durnholz M Eifler G and Endres H Exhaust-gas recir-culation a measure to reduce exhaust emissions of DIdiesel engines SAE technical paper 920725 1992

2 Lapuerta M Hernandez JJ and Gimenez F Evaluationof exhaust gas recirculation as a technique for reducingdiesel engine NOX emissions Proc IMechE Part D J

Automobile Engineering 2000 214 85ndash933 Zamboni G Moggia S and Capobianco M Hybrid EGR

and turbocharging systems control for low NOX and fuelconsumption in an automotive diesel engine Appl Energ2016 165 839ndash848

4 Moustapha H Zelesky MF Baines NC et al Axial andradial turbines vol 2 White River Junction VT Con-cepts NREC 2003

5 Wijetunge RS Hawley JG and Vaughan ND Anexhaust pressure control strategy for a diesel engine ProcIMechE Part D J Automobile Engineering 2004 218

449ndash4646 Zhu D and Zheng X Asymmetric twin-scroll turbochar-

ging in diesel engines for energy and emission improve-ment Energy 2017 141 702ndash714

7 Schmidt S Rose MG Muller M et al Variable asym-metric turbine for heavy duty truck engines In ASME

turbo expo 2013 turbine technical conference and expositionSan Antonio TX 3ndash7 June 2013 ppV05AT23A013-1ndashV05AT23A013-11 New York ASME

8 Brinkert N Sumser S Weber S et al Understanding thetwin scroll turbine flow similarity J Turbomach 2013135 021039

9 Dale A and Watson N Vaneless radial turbochargerturbine performance In Proceedings of the Institution of

Mechanical Engineers 3rd International Conference on

Turbocharging and Turbochargers London May 6ndash81986 Paper No C11086 pp 65ndash76 Mechanical Engi-neering Publications

10 Baines NC and Lavy M Flows in vaned and vaneless sta-tors of radial inflow turbocharger turbines In IMechE

turbocharging and turbochargers conference LondonMay Paper No C405005 pp 7ndash12 Mechanical Engi-neering Publications

11 Baines NC and Yeo JH Flow in a radial turbine underequal and partial admission conditions In IMechE Eur-

opean conference on turbomachinery latest developments

in a changing scene London March 12ndash13 MechanicalEngineering Publications Paper No C423002 pp 103ndash112

12 Baines NC Hajilouy-Benisi A and Yeo JH The pulseflow performance and modelling of radial inflow tur-bines In IMechE conference publications 1994 6 209ndash220 London Medical Engineering Publications

13 Mualler M Streule T Sumser S et al The asymmetric

twin scroll turbine for exhaust gas turbochargers InASME turbo expo 2008 power for land sea and air Ber-lin 9ndash13 June 2008 pp1547ndash1554 New York ASME

14 Wallace FJ and Blair GP The pulsating-flow perfor-mance of inward radial-flow turbines In ASME 1965

gas turbine conference and products show WashingtonDC 28 Februaryndash4 March 1965 New York ASME

15 Wallace FJ Cave PR and Miles J Performance of inwardradial flow turbines under steady flow conditions withspecial reference to high pressure ratios and partialadmission Proc IMechE 1969 184 1027ndash1042

16 Wallace FJ and Miles J Performance of inward radial

flow turbines under unsteady flow conditions with full

and partial admission Proc IMechE 1970 185 1091ndash

110517 Woods WA and Norbury JF Pulse flow performance of

axial turbines for marine turbochargers In ISME Tokyo

1973 Tokyo The Marine Engineering Society of Japan

pp1ndash518 Galindo J Fajardo P Navarro R et al Characterization

of a radial turbocharger turbine in pulsating flow by

means of CFD and its application to engine modeling

Appl Energ 2013 103 116ndash12719 Costall AW McDavid RM Martinez-Botas RF et al

Pulse performance modeling of a twin entry turbocharger

turbine under full and unequal admission J Turbomach

2011 133 02100520 Mamat AMB and Martinez-Botas RF Mean line flow

model of steady and pulsating flow of a mixed-flow tur-

bine turbocharger In ASME turbo expo 2010 power for

land sea and air Glasgow 14ndash18 June 2010 pp2393ndash

2404 New York ASME

21 Rajoo S Romagnoli A andMartinez-Botas RF Unsteady

performance analysis of a twin-entry variable geometry

turbocharger turbine Energy 2012 38 176ndash18922 De Bellis V Marelli S Bozza F et al 1D simulation and

experimental analysis of a turbocharger turbine for auto-

motive engines under steady and unsteady flow condi-

tions Energ Proced 2014 45 909ndash91823 Moustapha SH Kacker SC and Tremblay B An

improved incidence losses prediction method for turbine

airfoils In ASME international gas turbine and aeroengine

congress and exposition Toronto ON Canada 4ndash8 June

1989 ppV001T01A100-1ndashV001T01A100-10 New York

ASME

24 Baines N C (1998) A meanline prediction method for

radial turbine efficiency In Proceedings of the Institution

of Mechanical Engineers 6th International Conference on

Turbocharging and Air Management Systems London

November 3ndash5 Mechanical Engineering Publications

Paper No C554ndash6 pp 315ndash325

Appendix 1

Notation

K total pressure loss coefficientm mass flow rate (kgs)P pressure (Pa)T temperature (K)

g heat ratioh efficiency

Subscripts

0 total parameter1 volute inlet2 volute outlet3 impeller outleth hub sides shroud side

2256 Proc IMechE Part D J Automobile Engineering 233(8)

The CFD results and experiment data can matchwell with each other under steady inlet conditions andexperiments under pulsating inlet conditions will becarried out in the future to fine down the validation ofthe numerical model

In this paper first the performance and flow fieldsof the three types of turbines are compared understeady inlet conditions to determine the mechanism ofthe efficiency change and the geometric effects on theturbine performance Then the performance and flowfields under pulsating inlet conditions are investigatedto discover the pulse effects Finally based on theresults and analysis a criterion is established for theasymmetric twin-entry turbine design under steady andpulsating inlet conditions

Results under steady equal inlet conditions

The numerical results of the three simulated models areshown in Figure 7 The volute throat area was designedto maintain the same flow capacity in the three modelsas shown in the mass flow rate map It can be seen thatat low speeds the mass flow rates of the three turbine

types are nearly the same but at the highest speedthere was a drop in the mass flow rate of approximately02 which may cause a drop in the efficiency of thesymmetric turbine at high speeds putting it close to theasymmetric OP model Over the entire range of work-ing conditions the asymmetric turbine has the best tur-bine efficiency while the OP has the worst with thesymmetric turbine in the middle What is interesting isthat the efficiency drops between these three models atintervals of approximately 08 Based on this findingwe take the speed line 96000 rmin condition for fur-ther investigation to simplify the research process

Figure 8 shows the turbine and component perfor-mance comparison results It can be seen that the flowloss in the volute remains nearly constant with theexpansion ratio while the rotor efficiency shows thesame trend as turbine efficiency Although the sym-metric turbine has the lowest total pressure loss coeffi-cient this contributes little to the turbine performanceand considering that the test data used for the calcula-tions are measured at the interface the mixing flow atthe volute outlet domain may contribute to the rise inthe asymmetric and OPrsquos loss coefficients It is clear

Figure 6 Performance comparison between the numerical andexperimental results (a) efficiency and (b) mass flow rate

Figure 7 Performance comparison between the symmetricand the two different asymmetric turbines (a) efficiency and(b) mass flow rate

2250 Proc IMechE Part D J Automobile Engineering 233(8)

that the turbine efficiency is mainly determined by therotor efficiency and that the flow loss in the volute hasa negligible effect The main function of the volute is toprovide the rotor inlet flow field which may affect therotor efficiency Therefore even if the volute has differ-ent geometries the key point to determine the turbineefficiency is the flow field at the impeller inlet position

Figure 9 shows the shroud to hub spanwise distribu-tion for the rotor efficiency and radial velocity at theimpeller leading edge The rotor efficiency distributionis a hypothetical value that signifies the specific poweroutput that a unit mass of gas can provide if it entersthe impeller at a given spanwise position The valuesare calculated along the spanwise direction at the lead-ing edge and the corresponding trailing edge The radialvelocity represents the inlet mass flow and the rotorefficiency distribution and the match with the massflow distribution determine the impeller efficiencyThus if the high radial velocity position matches wellwith the high-efficiency region a senior rotor efficiencyis achieved

In Figure 9 the efficiency is higher at the hub sidethan at the shroud side and the asymmetric case hasthe highest efficiency while the OP has the lowest Forthe mass flow distribution for the asymmetric casesymmetric case and OP case the maximum mass flowposition moves from the hub side to the shroud side orfrom the high-efficiency region to the low-efficiencyregion These two reasons can explain the differences inthe rotor efficiency

Figure 10 shows the rotor inlet incidence angle ofthe three turbine models along the plane at 001mm

upstream of the rotor leading edge which is the criticalparameter to evaluate the rotor performance It isrecognized that optimum incidence is in the region of220 to 240 and that a deviation from this region cancause a drop in the efficiency Moustapha et al23 con-cluded that the radial turbine rotors are much moresensitive to positive incidence than to extremes of nega-tive incidence Then Baines24 put forward a postulatethat the rotor is more sensitive to the positive incidencenear the hub surface to explain the difference in the effi-ciencies under the hub and shroud side partial admis-sion condition Nevertheless he did not show anyexperimental or CFD results to support the postulate

It can be seen from Figure 10 that the positive inci-dence appears in the region that corresponds to theposition of the large volute scroll From Figure 10(a) to(c) the positive incidence region moves from the shroudside to the hub side of the impeller causing the effi-ciency to decrease These results correspond to Bainesrsquohypothesis

From the flow field at the rotor inlet region shown inFigure 11 a vortex exists caused by the flow separationat the hub side near the suction surface of the impellerCompared to the static entropy distribution it can beseen that there are two high entropy regions in the flowchannel one of which is near the hub and correspondsto the vortex The other is near the shroud which iscaused by the tip clearance flow Also the entropy gaincaused by the vortex is larger than the tip clearanceentropy gains

Figure 8 Turbine and component performance Figure 9 Rotor efficiency and radial velocity distribution at theimpeller leading edge

Wang and Zheng 2251

Moreover the vortex region is also where the highstatic entropy appears It is well known that a negative

incidence angle can weaken the flow separation whilea positive angle can aggravate it Therefore when thelarge volute scroll is placed on the shroud side the neg-ative incidence region is at the hub side of the impellerThis can weaken the flow separation fade or even erasethe vortex which may improve the turbine perfor-mance When the larger volute scroll is placed on thehub side the positive incidence region is at the hubside which may enhance the degree of the flow separa-tion and the vortex could become severe leading to thedrop in efficiency This is illustrated in the flow field inFigure 12

The static entropy distribution from Figure 13 at thevortex region shows the same conclusion It can be seenthat the asymmetric case only has an entropy gaincaused by tip clearance while the asymmetric OP modelhas the most severe entropy gain at the vortex regionThis can explain the efficiency differences between thesethree models

Figure 14 shows the velocity distribution at thevolute outlet plane at the impeller trailing edge It illus-trates that these three models have nearly the same out-let flow field and their mean velocities are 1403 1412and 1421ms respectively This means that the effi-ciency differences of the impellers are mainly caused bythe difference in the inlet flow condition

Results under the pulsating inlet condition

When the engine adopted the pulsating turbochargersystem the exhaust pressure is a pulse wave with a spe-cific amplitude and a specific main frequency and thefrequency is connected with the engine speed We canuse an ideal sinusoidal signal to represent the realengine exhaust pressure pulse wave as the turbine pul-sating inlet condition This paper has simplified the

Figure 10 Incidence angle distribution at the impeller leadingedge (a) asymmetric (b) symmetric and (c) asymmetric OP

Figure 11 Flow field and static entropy at the impeller inletregion

Figure 12 Flow field in the rotor passage (a) asymmetric (b) symmetric and (c) asymmetric OP

2252 Proc IMechE Part D J Automobile Engineering 233(8)

pulsating inlet condition to a sinusoidal signal whilekept the amplitude and frequency similar to the enginereal exhaust pressure and the prototype pressure waveis from a six-cylinder diesel engine in Mualler et al13

The pulsating frequency of the simulation is 25Hzcorresponding to the power output condition for this 11-L diesel engines This working point demand is morestringent to fuel consumption which means that a highturbine efficiency is required During pulsating inlet con-ditions the volute shroud side scroll and the hub sidescroll face two different inlet pulses that have a reversephase as shown in Figure 15 Considering the two isola-tion inlet conditions of the volute the isentropic efficiencyof the turbine at a specific time is defined as follows

h=m1s +m1h m1s +m1heth THORN T03=T01

m1s 1 (P3=P01s)g1

g

+m1h 1 (P3=P01h)

g1g

eth5THORN

Then the volute total pressure loss coefficient isnewly defined as follows

K=m1sP01s +m1hP01h (m1s +m1h)P02

(m1s +m1h)(P02 P2)eth6THORN

As for rotor efficiency considering that the form ofthe inlet and outlet remains the same as in the steadysituation the calculation remains unchanged

Figure 15 shows the performance of these three tur-bines under a 25-Hz pulsating inlet flow condition The

time-averaged efficiencies for these three turbines are7448 7469 and 7554 corresponding to theasymmetric symmetric and asymmetric OP modelrespectively The asymmetric type has the worst perfor-mance which is a contrary result of the steady perfor-mance where the asymmetric turbine shows the bestperformance This means that the performances of theturbines are different when under steady and pulsatinginlet flow conditions

Under the pulsating inlet conditions the highest effi-ciency points appear for all three models when the huband shroud inlet pressures are equal The efficiencydrops when facing the partial admission conditionespecially in the situation when the hub side inlet hasthe highest back-pressure

When the volute hub side inlet has the lower pres-sure corresponding to less mass flow the efficiencies ofthese three models are nearly the same and the asym-metric turbine is improved over the other twoHowever when the hub side has more pressure theefficiency drops and the asymmetric turbine is affectedthe most causing a decrease in the average efficiencyTherefore the inlet condition at the volute hub side isthe main factor to control the twin-entry turbine per-formance and the decrease in the mass flow may notcause a noticeable rise in the efficiency but the increasein mass flow will lead to an efficiency drop If the twin-entry turbine is symmetric the high efficiency appearswhen the shroud side inlet mass flow rate is larger thanthe hub side and drops when the mass flow is lesswhich agrees with previous conclusions under steadyinlet conditions from the literature

The asymmetric model that places the small scrollon the hub side is affected more by the hub side inletand the OP model that places the big scroll on the hubside is affected less Therefore the small scroll has moreinfluence on the turbine performance In summary thesmall scroll on the hub side is the main cause of deter-mining the turbine performance especially the effi-ciency drop when facing partial admission conditions

Comparing the component performance shown inFigure 16 with the turbine efficiency the same conclu-sion as the steady-state condition was found wherebythe turbine performances are mainly determined by the

Figure 13 Static entropy distribution in the rotor passage(a) asymmetric (b) symmetric and (c) asymmetric OP

Figure 14 Exit flow field at the impeller trailing edge (a) asymmetric (b) symmetric and (c) asymmetric OP

Wang and Zheng 2253

rotor performance and the contribution of the flow lossin the volute is marginally negligible

Figure 17 shows the rotor efficiency and the radialvelocity distribution at the impeller leading edge at timesteps A and B Time step A corresponds to the extremepartial admission condition when all of the turbineshave the efficiency drop It can be seen that the radialvelocity distribution is almost the same for the threemodels at this time The most important factor is theefficiency distribution and it is clear that the OP modelhas the highest efficiency and that the asymmetric onehas the lowest which influences the turbineperformance

Similar to the results under steady conditions theefficiency distribution for the pulsating inlet conditionis affected by the incidence angle at the impeller inletposition especially in the positive incidence regionFigure 18 illustrates that from the asymmetric case tothe OP case the positive incidence region moves fromthe hub side to the shroud side As mentioned beforethe positive incidence angle can sever the vortex in the

impeller flow path and cause the efficiency to dropThe impeller hub side leading edge is more sensitive toa positive incidence angle than the shroud sideTherefore as the positive incidence region moves fromthe hub to the shroud side the efficiency recoverswhich can explain the performance difference in thesethree models

Under equal steady inlet conditions the mass flow inthe big scroll is larger than in the small scroll resultingin a faster flow velocity and positive incidence angleCombined with the conclusion that the hub side of theimpeller is more sensitive to positive incidence if thebig scroll is on the hub side the efficiency drops Underpulsating inlet conditions the efficiency is different tothe steady-state situation especially for the asymmetricturbine when the hub side is charged more than theshroud side This is because the big volute has a highermass flow storage ability than the small one that is thesmall scroll response is more sensitive to the pressurerise and flow in the small scroll is accelerated leadingto a higher outlet velocity which causes a greater inci-dence angle When the pressure rises for the asymmetricturbine with the small scroll on the hub side a positiveincidence angle region occurs at the hub side and theefficiency drops

Conclusion

Simulations for a symmetric and two opposite asym-metric twin scroll turbines under steady and pulsatinginlet conditions have been conducted to study the dif-ferences in the flow field and performance betweenthese three types of turbines an experiment was carriedout to validate the numerical methodology under steadyinlet conditions and the conclusions are as follows

1 Under steady admission conditions the asym-metric turbine with the big scroll on the shroud sidehas the best efficiency which is nearly 16 higherthan the contrary type (OP) Under steady full inletconditions the incidence angle at the region out-side the big scroll is positive The impeller hub sideleading edge is more sensitive to the positive inci-dence than the shroud side due to the hub side vor-tex in the impeller flow path while the positiveincidence severs the vortex and causes the efficiencyto drop Therefore if the big scroll is on the hubside the efficiency is worse than on the shroudside

2 Under pulsating inlet conditions the results arecontrary to the steady state The asymmetric tur-bine with the big scroll on the hub side (OP) hasthe best efficiency which is nearly 11 higherthan on the shroud side Since the small scroll massflow capacity is smaller than the big scroll whenthe backpressure increases the flow in the smallscroll is faster than in the big scroll which maycause a positive incidence angle at the impeller

Figure 15 Turbine performance under the 25-Hz pulsatinginlet condition

Figure 16 Component performance under the 25-Hzpulsating inlet condition

2254 Proc IMechE Part D J Automobile Engineering 233(8)

leading edge Therefore if the small scroll is set onthe hub side the positive incidence appears at thehub side and the efficiency is degraded

3 The design criterion for the asymmetric twin-entryturbine state is as follows when under steady inletconditions (eg constant-pressure turbochargingsystem on the diesel engine) it is better to place thebig scroll on the shroud side When under pulsat-ing inlet conditions (eg pulse turbocharging sys-tem on the gasoline engine) it is better to place thebig scroll on the hub side

4 Simulation calculations were conducted for a givenrotational speed and a given simplified scroll inputpressure It must be controlled whether these con-clusions remain the same for real engine pressureconditions and other engines speeds Experimentswill have to be carried out to check these results

Declaration of conflicting interests

The author(s) declared no potential conflicts of interestwith respect to the research authorship andor publica-tion of this article

Funding

The author(s) received no financial support for theresearch authorship andor publication of this article

Figure 17 Rotor efficiency and radial velocity distribution at the impeller leading edge (a) A time step and (b) B time step

Figure 18 Incidence angle distribution at the impeller leadingedge (a) A time step and (b) B time step

Wang and Zheng 2255

References

1 Durnholz M Eifler G and Endres H Exhaust-gas recir-culation a measure to reduce exhaust emissions of DIdiesel engines SAE technical paper 920725 1992

2 Lapuerta M Hernandez JJ and Gimenez F Evaluationof exhaust gas recirculation as a technique for reducingdiesel engine NOX emissions Proc IMechE Part D J

Automobile Engineering 2000 214 85ndash933 Zamboni G Moggia S and Capobianco M Hybrid EGR

and turbocharging systems control for low NOX and fuelconsumption in an automotive diesel engine Appl Energ2016 165 839ndash848

4 Moustapha H Zelesky MF Baines NC et al Axial andradial turbines vol 2 White River Junction VT Con-cepts NREC 2003

5 Wijetunge RS Hawley JG and Vaughan ND Anexhaust pressure control strategy for a diesel engine ProcIMechE Part D J Automobile Engineering 2004 218

449ndash4646 Zhu D and Zheng X Asymmetric twin-scroll turbochar-

ging in diesel engines for energy and emission improve-ment Energy 2017 141 702ndash714

7 Schmidt S Rose MG Muller M et al Variable asym-metric turbine for heavy duty truck engines In ASME

turbo expo 2013 turbine technical conference and expositionSan Antonio TX 3ndash7 June 2013 ppV05AT23A013-1ndashV05AT23A013-11 New York ASME

8 Brinkert N Sumser S Weber S et al Understanding thetwin scroll turbine flow similarity J Turbomach 2013135 021039

9 Dale A and Watson N Vaneless radial turbochargerturbine performance In Proceedings of the Institution of

Mechanical Engineers 3rd International Conference on

Turbocharging and Turbochargers London May 6ndash81986 Paper No C11086 pp 65ndash76 Mechanical Engi-neering Publications

10 Baines NC and Lavy M Flows in vaned and vaneless sta-tors of radial inflow turbocharger turbines In IMechE

turbocharging and turbochargers conference LondonMay Paper No C405005 pp 7ndash12 Mechanical Engi-neering Publications

11 Baines NC and Yeo JH Flow in a radial turbine underequal and partial admission conditions In IMechE Eur-

opean conference on turbomachinery latest developments

in a changing scene London March 12ndash13 MechanicalEngineering Publications Paper No C423002 pp 103ndash112

12 Baines NC Hajilouy-Benisi A and Yeo JH The pulseflow performance and modelling of radial inflow tur-bines In IMechE conference publications 1994 6 209ndash220 London Medical Engineering Publications

13 Mualler M Streule T Sumser S et al The asymmetric

twin scroll turbine for exhaust gas turbochargers InASME turbo expo 2008 power for land sea and air Ber-lin 9ndash13 June 2008 pp1547ndash1554 New York ASME

14 Wallace FJ and Blair GP The pulsating-flow perfor-mance of inward radial-flow turbines In ASME 1965

gas turbine conference and products show WashingtonDC 28 Februaryndash4 March 1965 New York ASME

15 Wallace FJ Cave PR and Miles J Performance of inwardradial flow turbines under steady flow conditions withspecial reference to high pressure ratios and partialadmission Proc IMechE 1969 184 1027ndash1042

16 Wallace FJ and Miles J Performance of inward radial

flow turbines under unsteady flow conditions with full

and partial admission Proc IMechE 1970 185 1091ndash

110517 Woods WA and Norbury JF Pulse flow performance of

axial turbines for marine turbochargers In ISME Tokyo

1973 Tokyo The Marine Engineering Society of Japan

pp1ndash518 Galindo J Fajardo P Navarro R et al Characterization

of a radial turbocharger turbine in pulsating flow by

means of CFD and its application to engine modeling

Appl Energ 2013 103 116ndash12719 Costall AW McDavid RM Martinez-Botas RF et al

Pulse performance modeling of a twin entry turbocharger

turbine under full and unequal admission J Turbomach

2011 133 02100520 Mamat AMB and Martinez-Botas RF Mean line flow

model of steady and pulsating flow of a mixed-flow tur-

bine turbocharger In ASME turbo expo 2010 power for

land sea and air Glasgow 14ndash18 June 2010 pp2393ndash

2404 New York ASME

21 Rajoo S Romagnoli A andMartinez-Botas RF Unsteady

performance analysis of a twin-entry variable geometry

turbocharger turbine Energy 2012 38 176ndash18922 De Bellis V Marelli S Bozza F et al 1D simulation and

experimental analysis of a turbocharger turbine for auto-

motive engines under steady and unsteady flow condi-

tions Energ Proced 2014 45 909ndash91823 Moustapha SH Kacker SC and Tremblay B An

improved incidence losses prediction method for turbine

airfoils In ASME international gas turbine and aeroengine

congress and exposition Toronto ON Canada 4ndash8 June

1989 ppV001T01A100-1ndashV001T01A100-10 New York

ASME

24 Baines N C (1998) A meanline prediction method for

radial turbine efficiency In Proceedings of the Institution

of Mechanical Engineers 6th International Conference on

Turbocharging and Air Management Systems London

November 3ndash5 Mechanical Engineering Publications

Paper No C554ndash6 pp 315ndash325

Appendix 1

Notation

K total pressure loss coefficientm mass flow rate (kgs)P pressure (Pa)T temperature (K)

g heat ratioh efficiency

Subscripts

0 total parameter1 volute inlet2 volute outlet3 impeller outleth hub sides shroud side

2256 Proc IMechE Part D J Automobile Engineering 233(8)

that the turbine efficiency is mainly determined by therotor efficiency and that the flow loss in the volute hasa negligible effect The main function of the volute is toprovide the rotor inlet flow field which may affect therotor efficiency Therefore even if the volute has differ-ent geometries the key point to determine the turbineefficiency is the flow field at the impeller inlet position

Figure 9 shows the shroud to hub spanwise distribu-tion for the rotor efficiency and radial velocity at theimpeller leading edge The rotor efficiency distributionis a hypothetical value that signifies the specific poweroutput that a unit mass of gas can provide if it entersthe impeller at a given spanwise position The valuesare calculated along the spanwise direction at the lead-ing edge and the corresponding trailing edge The radialvelocity represents the inlet mass flow and the rotorefficiency distribution and the match with the massflow distribution determine the impeller efficiencyThus if the high radial velocity position matches wellwith the high-efficiency region a senior rotor efficiencyis achieved

In Figure 9 the efficiency is higher at the hub sidethan at the shroud side and the asymmetric case hasthe highest efficiency while the OP has the lowest Forthe mass flow distribution for the asymmetric casesymmetric case and OP case the maximum mass flowposition moves from the hub side to the shroud side orfrom the high-efficiency region to the low-efficiencyregion These two reasons can explain the differences inthe rotor efficiency

Figure 10 shows the rotor inlet incidence angle ofthe three turbine models along the plane at 001mm

upstream of the rotor leading edge which is the criticalparameter to evaluate the rotor performance It isrecognized that optimum incidence is in the region of220 to 240 and that a deviation from this region cancause a drop in the efficiency Moustapha et al23 con-cluded that the radial turbine rotors are much moresensitive to positive incidence than to extremes of nega-tive incidence Then Baines24 put forward a postulatethat the rotor is more sensitive to the positive incidencenear the hub surface to explain the difference in the effi-ciencies under the hub and shroud side partial admis-sion condition Nevertheless he did not show anyexperimental or CFD results to support the postulate

It can be seen from Figure 10 that the positive inci-dence appears in the region that corresponds to theposition of the large volute scroll From Figure 10(a) to(c) the positive incidence region moves from the shroudside to the hub side of the impeller causing the effi-ciency to decrease These results correspond to Bainesrsquohypothesis

From the flow field at the rotor inlet region shown inFigure 11 a vortex exists caused by the flow separationat the hub side near the suction surface of the impellerCompared to the static entropy distribution it can beseen that there are two high entropy regions in the flowchannel one of which is near the hub and correspondsto the vortex The other is near the shroud which iscaused by the tip clearance flow Also the entropy gaincaused by the vortex is larger than the tip clearanceentropy gains

Figure 8 Turbine and component performance Figure 9 Rotor efficiency and radial velocity distribution at theimpeller leading edge

Wang and Zheng 2251

Moreover the vortex region is also where the highstatic entropy appears It is well known that a negative

incidence angle can weaken the flow separation whilea positive angle can aggravate it Therefore when thelarge volute scroll is placed on the shroud side the neg-ative incidence region is at the hub side of the impellerThis can weaken the flow separation fade or even erasethe vortex which may improve the turbine perfor-mance When the larger volute scroll is placed on thehub side the positive incidence region is at the hubside which may enhance the degree of the flow separa-tion and the vortex could become severe leading to thedrop in efficiency This is illustrated in the flow field inFigure 12

The static entropy distribution from Figure 13 at thevortex region shows the same conclusion It can be seenthat the asymmetric case only has an entropy gaincaused by tip clearance while the asymmetric OP modelhas the most severe entropy gain at the vortex regionThis can explain the efficiency differences between thesethree models

Figure 14 shows the velocity distribution at thevolute outlet plane at the impeller trailing edge It illus-trates that these three models have nearly the same out-let flow field and their mean velocities are 1403 1412and 1421ms respectively This means that the effi-ciency differences of the impellers are mainly caused bythe difference in the inlet flow condition

Results under the pulsating inlet condition

When the engine adopted the pulsating turbochargersystem the exhaust pressure is a pulse wave with a spe-cific amplitude and a specific main frequency and thefrequency is connected with the engine speed We canuse an ideal sinusoidal signal to represent the realengine exhaust pressure pulse wave as the turbine pul-sating inlet condition This paper has simplified the

Figure 10 Incidence angle distribution at the impeller leadingedge (a) asymmetric (b) symmetric and (c) asymmetric OP

Figure 11 Flow field and static entropy at the impeller inletregion

Figure 12 Flow field in the rotor passage (a) asymmetric (b) symmetric and (c) asymmetric OP

2252 Proc IMechE Part D J Automobile Engineering 233(8)

pulsating inlet condition to a sinusoidal signal whilekept the amplitude and frequency similar to the enginereal exhaust pressure and the prototype pressure waveis from a six-cylinder diesel engine in Mualler et al13

The pulsating frequency of the simulation is 25Hzcorresponding to the power output condition for this 11-L diesel engines This working point demand is morestringent to fuel consumption which means that a highturbine efficiency is required During pulsating inlet con-ditions the volute shroud side scroll and the hub sidescroll face two different inlet pulses that have a reversephase as shown in Figure 15 Considering the two isola-tion inlet conditions of the volute the isentropic efficiencyof the turbine at a specific time is defined as follows

h=m1s +m1h m1s +m1heth THORN T03=T01

m1s 1 (P3=P01s)g1

g

+m1h 1 (P3=P01h)

g1g

eth5THORN

Then the volute total pressure loss coefficient isnewly defined as follows

K=m1sP01s +m1hP01h (m1s +m1h)P02

(m1s +m1h)(P02 P2)eth6THORN

As for rotor efficiency considering that the form ofthe inlet and outlet remains the same as in the steadysituation the calculation remains unchanged

Figure 15 shows the performance of these three tur-bines under a 25-Hz pulsating inlet flow condition The

time-averaged efficiencies for these three turbines are7448 7469 and 7554 corresponding to theasymmetric symmetric and asymmetric OP modelrespectively The asymmetric type has the worst perfor-mance which is a contrary result of the steady perfor-mance where the asymmetric turbine shows the bestperformance This means that the performances of theturbines are different when under steady and pulsatinginlet flow conditions

Under the pulsating inlet conditions the highest effi-ciency points appear for all three models when the huband shroud inlet pressures are equal The efficiencydrops when facing the partial admission conditionespecially in the situation when the hub side inlet hasthe highest back-pressure

When the volute hub side inlet has the lower pres-sure corresponding to less mass flow the efficiencies ofthese three models are nearly the same and the asym-metric turbine is improved over the other twoHowever when the hub side has more pressure theefficiency drops and the asymmetric turbine is affectedthe most causing a decrease in the average efficiencyTherefore the inlet condition at the volute hub side isthe main factor to control the twin-entry turbine per-formance and the decrease in the mass flow may notcause a noticeable rise in the efficiency but the increasein mass flow will lead to an efficiency drop If the twin-entry turbine is symmetric the high efficiency appearswhen the shroud side inlet mass flow rate is larger thanthe hub side and drops when the mass flow is lesswhich agrees with previous conclusions under steadyinlet conditions from the literature

The asymmetric model that places the small scrollon the hub side is affected more by the hub side inletand the OP model that places the big scroll on the hubside is affected less Therefore the small scroll has moreinfluence on the turbine performance In summary thesmall scroll on the hub side is the main cause of deter-mining the turbine performance especially the effi-ciency drop when facing partial admission conditions

Comparing the component performance shown inFigure 16 with the turbine efficiency the same conclu-sion as the steady-state condition was found wherebythe turbine performances are mainly determined by the

Figure 13 Static entropy distribution in the rotor passage(a) asymmetric (b) symmetric and (c) asymmetric OP

Figure 14 Exit flow field at the impeller trailing edge (a) asymmetric (b) symmetric and (c) asymmetric OP

Wang and Zheng 2253

rotor performance and the contribution of the flow lossin the volute is marginally negligible

Figure 17 shows the rotor efficiency and the radialvelocity distribution at the impeller leading edge at timesteps A and B Time step A corresponds to the extremepartial admission condition when all of the turbineshave the efficiency drop It can be seen that the radialvelocity distribution is almost the same for the threemodels at this time The most important factor is theefficiency distribution and it is clear that the OP modelhas the highest efficiency and that the asymmetric onehas the lowest which influences the turbineperformance

Similar to the results under steady conditions theefficiency distribution for the pulsating inlet conditionis affected by the incidence angle at the impeller inletposition especially in the positive incidence regionFigure 18 illustrates that from the asymmetric case tothe OP case the positive incidence region moves fromthe hub side to the shroud side As mentioned beforethe positive incidence angle can sever the vortex in the

impeller flow path and cause the efficiency to dropThe impeller hub side leading edge is more sensitive toa positive incidence angle than the shroud sideTherefore as the positive incidence region moves fromthe hub to the shroud side the efficiency recoverswhich can explain the performance difference in thesethree models

Under equal steady inlet conditions the mass flow inthe big scroll is larger than in the small scroll resultingin a faster flow velocity and positive incidence angleCombined with the conclusion that the hub side of theimpeller is more sensitive to positive incidence if thebig scroll is on the hub side the efficiency drops Underpulsating inlet conditions the efficiency is different tothe steady-state situation especially for the asymmetricturbine when the hub side is charged more than theshroud side This is because the big volute has a highermass flow storage ability than the small one that is thesmall scroll response is more sensitive to the pressurerise and flow in the small scroll is accelerated leadingto a higher outlet velocity which causes a greater inci-dence angle When the pressure rises for the asymmetricturbine with the small scroll on the hub side a positiveincidence angle region occurs at the hub side and theefficiency drops

Conclusion

Simulations for a symmetric and two opposite asym-metric twin scroll turbines under steady and pulsatinginlet conditions have been conducted to study the dif-ferences in the flow field and performance betweenthese three types of turbines an experiment was carriedout to validate the numerical methodology under steadyinlet conditions and the conclusions are as follows

1 Under steady admission conditions the asym-metric turbine with the big scroll on the shroud sidehas the best efficiency which is nearly 16 higherthan the contrary type (OP) Under steady full inletconditions the incidence angle at the region out-side the big scroll is positive The impeller hub sideleading edge is more sensitive to the positive inci-dence than the shroud side due to the hub side vor-tex in the impeller flow path while the positiveincidence severs the vortex and causes the efficiencyto drop Therefore if the big scroll is on the hubside the efficiency is worse than on the shroudside

2 Under pulsating inlet conditions the results arecontrary to the steady state The asymmetric tur-bine with the big scroll on the hub side (OP) hasthe best efficiency which is nearly 11 higherthan on the shroud side Since the small scroll massflow capacity is smaller than the big scroll whenthe backpressure increases the flow in the smallscroll is faster than in the big scroll which maycause a positive incidence angle at the impeller

Figure 15 Turbine performance under the 25-Hz pulsatinginlet condition

Figure 16 Component performance under the 25-Hzpulsating inlet condition

2254 Proc IMechE Part D J Automobile Engineering 233(8)

leading edge Therefore if the small scroll is set onthe hub side the positive incidence appears at thehub side and the efficiency is degraded

3 The design criterion for the asymmetric twin-entryturbine state is as follows when under steady inletconditions (eg constant-pressure turbochargingsystem on the diesel engine) it is better to place thebig scroll on the shroud side When under pulsat-ing inlet conditions (eg pulse turbocharging sys-tem on the gasoline engine) it is better to place thebig scroll on the hub side

4 Simulation calculations were conducted for a givenrotational speed and a given simplified scroll inputpressure It must be controlled whether these con-clusions remain the same for real engine pressureconditions and other engines speeds Experimentswill have to be carried out to check these results

Declaration of conflicting interests

The author(s) declared no potential conflicts of interestwith respect to the research authorship andor publica-tion of this article

Funding

The author(s) received no financial support for theresearch authorship andor publication of this article

Figure 17 Rotor efficiency and radial velocity distribution at the impeller leading edge (a) A time step and (b) B time step

Figure 18 Incidence angle distribution at the impeller leadingedge (a) A time step and (b) B time step

Wang and Zheng 2255

References

1 Durnholz M Eifler G and Endres H Exhaust-gas recir-culation a measure to reduce exhaust emissions of DIdiesel engines SAE technical paper 920725 1992

2 Lapuerta M Hernandez JJ and Gimenez F Evaluationof exhaust gas recirculation as a technique for reducingdiesel engine NOX emissions Proc IMechE Part D J

Automobile Engineering 2000 214 85ndash933 Zamboni G Moggia S and Capobianco M Hybrid EGR

and turbocharging systems control for low NOX and fuelconsumption in an automotive diesel engine Appl Energ2016 165 839ndash848

4 Moustapha H Zelesky MF Baines NC et al Axial andradial turbines vol 2 White River Junction VT Con-cepts NREC 2003

5 Wijetunge RS Hawley JG and Vaughan ND Anexhaust pressure control strategy for a diesel engine ProcIMechE Part D J Automobile Engineering 2004 218

449ndash4646 Zhu D and Zheng X Asymmetric twin-scroll turbochar-

ging in diesel engines for energy and emission improve-ment Energy 2017 141 702ndash714

7 Schmidt S Rose MG Muller M et al Variable asym-metric turbine for heavy duty truck engines In ASME

turbo expo 2013 turbine technical conference and expositionSan Antonio TX 3ndash7 June 2013 ppV05AT23A013-1ndashV05AT23A013-11 New York ASME

8 Brinkert N Sumser S Weber S et al Understanding thetwin scroll turbine flow similarity J Turbomach 2013135 021039

9 Dale A and Watson N Vaneless radial turbochargerturbine performance In Proceedings of the Institution of

Mechanical Engineers 3rd International Conference on

Turbocharging and Turbochargers London May 6ndash81986 Paper No C11086 pp 65ndash76 Mechanical Engi-neering Publications

10 Baines NC and Lavy M Flows in vaned and vaneless sta-tors of radial inflow turbocharger turbines In IMechE

turbocharging and turbochargers conference LondonMay Paper No C405005 pp 7ndash12 Mechanical Engi-neering Publications

11 Baines NC and Yeo JH Flow in a radial turbine underequal and partial admission conditions In IMechE Eur-

opean conference on turbomachinery latest developments

in a changing scene London March 12ndash13 MechanicalEngineering Publications Paper No C423002 pp 103ndash112

12 Baines NC Hajilouy-Benisi A and Yeo JH The pulseflow performance and modelling of radial inflow tur-bines In IMechE conference publications 1994 6 209ndash220 London Medical Engineering Publications

13 Mualler M Streule T Sumser S et al The asymmetric

twin scroll turbine for exhaust gas turbochargers InASME turbo expo 2008 power for land sea and air Ber-lin 9ndash13 June 2008 pp1547ndash1554 New York ASME

14 Wallace FJ and Blair GP The pulsating-flow perfor-mance of inward radial-flow turbines In ASME 1965

gas turbine conference and products show WashingtonDC 28 Februaryndash4 March 1965 New York ASME

15 Wallace FJ Cave PR and Miles J Performance of inwardradial flow turbines under steady flow conditions withspecial reference to high pressure ratios and partialadmission Proc IMechE 1969 184 1027ndash1042

16 Wallace FJ and Miles J Performance of inward radial

flow turbines under unsteady flow conditions with full

and partial admission Proc IMechE 1970 185 1091ndash

110517 Woods WA and Norbury JF Pulse flow performance of

axial turbines for marine turbochargers In ISME Tokyo

1973 Tokyo The Marine Engineering Society of Japan

pp1ndash518 Galindo J Fajardo P Navarro R et al Characterization

of a radial turbocharger turbine in pulsating flow by

means of CFD and its application to engine modeling

Appl Energ 2013 103 116ndash12719 Costall AW McDavid RM Martinez-Botas RF et al

Pulse performance modeling of a twin entry turbocharger

turbine under full and unequal admission J Turbomach

2011 133 02100520 Mamat AMB and Martinez-Botas RF Mean line flow

model of steady and pulsating flow of a mixed-flow tur-

bine turbocharger In ASME turbo expo 2010 power for

land sea and air Glasgow 14ndash18 June 2010 pp2393ndash

2404 New York ASME

21 Rajoo S Romagnoli A andMartinez-Botas RF Unsteady

performance analysis of a twin-entry variable geometry

turbocharger turbine Energy 2012 38 176ndash18922 De Bellis V Marelli S Bozza F et al 1D simulation and

experimental analysis of a turbocharger turbine for auto-

motive engines under steady and unsteady flow condi-

tions Energ Proced 2014 45 909ndash91823 Moustapha SH Kacker SC and Tremblay B An

improved incidence losses prediction method for turbine

airfoils In ASME international gas turbine and aeroengine

congress and exposition Toronto ON Canada 4ndash8 June

1989 ppV001T01A100-1ndashV001T01A100-10 New York

ASME

24 Baines N C (1998) A meanline prediction method for

radial turbine efficiency In Proceedings of the Institution

of Mechanical Engineers 6th International Conference on

Turbocharging and Air Management Systems London

November 3ndash5 Mechanical Engineering Publications

Paper No C554ndash6 pp 315ndash325

Appendix 1

Notation

K total pressure loss coefficientm mass flow rate (kgs)P pressure (Pa)T temperature (K)

g heat ratioh efficiency

Subscripts

0 total parameter1 volute inlet2 volute outlet3 impeller outleth hub sides shroud side

2256 Proc IMechE Part D J Automobile Engineering 233(8)

Moreover the vortex region is also where the highstatic entropy appears It is well known that a negative

incidence angle can weaken the flow separation whilea positive angle can aggravate it Therefore when thelarge volute scroll is placed on the shroud side the neg-ative incidence region is at the hub side of the impellerThis can weaken the flow separation fade or even erasethe vortex which may improve the turbine perfor-mance When the larger volute scroll is placed on thehub side the positive incidence region is at the hubside which may enhance the degree of the flow separa-tion and the vortex could become severe leading to thedrop in efficiency This is illustrated in the flow field inFigure 12

The static entropy distribution from Figure 13 at thevortex region shows the same conclusion It can be seenthat the asymmetric case only has an entropy gaincaused by tip clearance while the asymmetric OP modelhas the most severe entropy gain at the vortex regionThis can explain the efficiency differences between thesethree models

Figure 14 shows the velocity distribution at thevolute outlet plane at the impeller trailing edge It illus-trates that these three models have nearly the same out-let flow field and their mean velocities are 1403 1412and 1421ms respectively This means that the effi-ciency differences of the impellers are mainly caused bythe difference in the inlet flow condition

Results under the pulsating inlet condition

When the engine adopted the pulsating turbochargersystem the exhaust pressure is a pulse wave with a spe-cific amplitude and a specific main frequency and thefrequency is connected with the engine speed We canuse an ideal sinusoidal signal to represent the realengine exhaust pressure pulse wave as the turbine pul-sating inlet condition This paper has simplified the

Figure 10 Incidence angle distribution at the impeller leadingedge (a) asymmetric (b) symmetric and (c) asymmetric OP

Figure 11 Flow field and static entropy at the impeller inletregion

Figure 12 Flow field in the rotor passage (a) asymmetric (b) symmetric and (c) asymmetric OP

2252 Proc IMechE Part D J Automobile Engineering 233(8)

pulsating inlet condition to a sinusoidal signal whilekept the amplitude and frequency similar to the enginereal exhaust pressure and the prototype pressure waveis from a six-cylinder diesel engine in Mualler et al13

The pulsating frequency of the simulation is 25Hzcorresponding to the power output condition for this 11-L diesel engines This working point demand is morestringent to fuel consumption which means that a highturbine efficiency is required During pulsating inlet con-ditions the volute shroud side scroll and the hub sidescroll face two different inlet pulses that have a reversephase as shown in Figure 15 Considering the two isola-tion inlet conditions of the volute the isentropic efficiencyof the turbine at a specific time is defined as follows

h=m1s +m1h m1s +m1heth THORN T03=T01

m1s 1 (P3=P01s)g1

g

+m1h 1 (P3=P01h)

g1g

eth5THORN

Then the volute total pressure loss coefficient isnewly defined as follows

K=m1sP01s +m1hP01h (m1s +m1h)P02

(m1s +m1h)(P02 P2)eth6THORN

As for rotor efficiency considering that the form ofthe inlet and outlet remains the same as in the steadysituation the calculation remains unchanged

Figure 15 shows the performance of these three tur-bines under a 25-Hz pulsating inlet flow condition The

time-averaged efficiencies for these three turbines are7448 7469 and 7554 corresponding to theasymmetric symmetric and asymmetric OP modelrespectively The asymmetric type has the worst perfor-mance which is a contrary result of the steady perfor-mance where the asymmetric turbine shows the bestperformance This means that the performances of theturbines are different when under steady and pulsatinginlet flow conditions

Under the pulsating inlet conditions the highest effi-ciency points appear for all three models when the huband shroud inlet pressures are equal The efficiencydrops when facing the partial admission conditionespecially in the situation when the hub side inlet hasthe highest back-pressure

When the volute hub side inlet has the lower pres-sure corresponding to less mass flow the efficiencies ofthese three models are nearly the same and the asym-metric turbine is improved over the other twoHowever when the hub side has more pressure theefficiency drops and the asymmetric turbine is affectedthe most causing a decrease in the average efficiencyTherefore the inlet condition at the volute hub side isthe main factor to control the twin-entry turbine per-formance and the decrease in the mass flow may notcause a noticeable rise in the efficiency but the increasein mass flow will lead to an efficiency drop If the twin-entry turbine is symmetric the high efficiency appearswhen the shroud side inlet mass flow rate is larger thanthe hub side and drops when the mass flow is lesswhich agrees with previous conclusions under steadyinlet conditions from the literature

The asymmetric model that places the small scrollon the hub side is affected more by the hub side inletand the OP model that places the big scroll on the hubside is affected less Therefore the small scroll has moreinfluence on the turbine performance In summary thesmall scroll on the hub side is the main cause of deter-mining the turbine performance especially the effi-ciency drop when facing partial admission conditions

Comparing the component performance shown inFigure 16 with the turbine efficiency the same conclu-sion as the steady-state condition was found wherebythe turbine performances are mainly determined by the

Figure 13 Static entropy distribution in the rotor passage(a) asymmetric (b) symmetric and (c) asymmetric OP

Figure 14 Exit flow field at the impeller trailing edge (a) asymmetric (b) symmetric and (c) asymmetric OP

Wang and Zheng 2253

rotor performance and the contribution of the flow lossin the volute is marginally negligible

Figure 17 shows the rotor efficiency and the radialvelocity distribution at the impeller leading edge at timesteps A and B Time step A corresponds to the extremepartial admission condition when all of the turbineshave the efficiency drop It can be seen that the radialvelocity distribution is almost the same for the threemodels at this time The most important factor is theefficiency distribution and it is clear that the OP modelhas the highest efficiency and that the asymmetric onehas the lowest which influences the turbineperformance

Similar to the results under steady conditions theefficiency distribution for the pulsating inlet conditionis affected by the incidence angle at the impeller inletposition especially in the positive incidence regionFigure 18 illustrates that from the asymmetric case tothe OP case the positive incidence region moves fromthe hub side to the shroud side As mentioned beforethe positive incidence angle can sever the vortex in the

impeller flow path and cause the efficiency to dropThe impeller hub side leading edge is more sensitive toa positive incidence angle than the shroud sideTherefore as the positive incidence region moves fromthe hub to the shroud side the efficiency recoverswhich can explain the performance difference in thesethree models

Under equal steady inlet conditions the mass flow inthe big scroll is larger than in the small scroll resultingin a faster flow velocity and positive incidence angleCombined with the conclusion that the hub side of theimpeller is more sensitive to positive incidence if thebig scroll is on the hub side the efficiency drops Underpulsating inlet conditions the efficiency is different tothe steady-state situation especially for the asymmetricturbine when the hub side is charged more than theshroud side This is because the big volute has a highermass flow storage ability than the small one that is thesmall scroll response is more sensitive to the pressurerise and flow in the small scroll is accelerated leadingto a higher outlet velocity which causes a greater inci-dence angle When the pressure rises for the asymmetricturbine with the small scroll on the hub side a positiveincidence angle region occurs at the hub side and theefficiency drops

Conclusion

Simulations for a symmetric and two opposite asym-metric twin scroll turbines under steady and pulsatinginlet conditions have been conducted to study the dif-ferences in the flow field and performance betweenthese three types of turbines an experiment was carriedout to validate the numerical methodology under steadyinlet conditions and the conclusions are as follows

1 Under steady admission conditions the asym-metric turbine with the big scroll on the shroud sidehas the best efficiency which is nearly 16 higherthan the contrary type (OP) Under steady full inletconditions the incidence angle at the region out-side the big scroll is positive The impeller hub sideleading edge is more sensitive to the positive inci-dence than the shroud side due to the hub side vor-tex in the impeller flow path while the positiveincidence severs the vortex and causes the efficiencyto drop Therefore if the big scroll is on the hubside the efficiency is worse than on the shroudside

2 Under pulsating inlet conditions the results arecontrary to the steady state The asymmetric tur-bine with the big scroll on the hub side (OP) hasthe best efficiency which is nearly 11 higherthan on the shroud side Since the small scroll massflow capacity is smaller than the big scroll whenthe backpressure increases the flow in the smallscroll is faster than in the big scroll which maycause a positive incidence angle at the impeller

Figure 15 Turbine performance under the 25-Hz pulsatinginlet condition

Figure 16 Component performance under the 25-Hzpulsating inlet condition

2254 Proc IMechE Part D J Automobile Engineering 233(8)

leading edge Therefore if the small scroll is set onthe hub side the positive incidence appears at thehub side and the efficiency is degraded

3 The design criterion for the asymmetric twin-entryturbine state is as follows when under steady inletconditions (eg constant-pressure turbochargingsystem on the diesel engine) it is better to place thebig scroll on the shroud side When under pulsat-ing inlet conditions (eg pulse turbocharging sys-tem on the gasoline engine) it is better to place thebig scroll on the hub side

4 Simulation calculations were conducted for a givenrotational speed and a given simplified scroll inputpressure It must be controlled whether these con-clusions remain the same for real engine pressureconditions and other engines speeds Experimentswill have to be carried out to check these results

Declaration of conflicting interests

The author(s) declared no potential conflicts of interestwith respect to the research authorship andor publica-tion of this article

Funding

The author(s) received no financial support for theresearch authorship andor publication of this article

Figure 17 Rotor efficiency and radial velocity distribution at the impeller leading edge (a) A time step and (b) B time step

Figure 18 Incidence angle distribution at the impeller leadingedge (a) A time step and (b) B time step

Wang and Zheng 2255

References

1 Durnholz M Eifler G and Endres H Exhaust-gas recir-culation a measure to reduce exhaust emissions of DIdiesel engines SAE technical paper 920725 1992

2 Lapuerta M Hernandez JJ and Gimenez F Evaluationof exhaust gas recirculation as a technique for reducingdiesel engine NOX emissions Proc IMechE Part D J

Automobile Engineering 2000 214 85ndash933 Zamboni G Moggia S and Capobianco M Hybrid EGR

and turbocharging systems control for low NOX and fuelconsumption in an automotive diesel engine Appl Energ2016 165 839ndash848

4 Moustapha H Zelesky MF Baines NC et al Axial andradial turbines vol 2 White River Junction VT Con-cepts NREC 2003

5 Wijetunge RS Hawley JG and Vaughan ND Anexhaust pressure control strategy for a diesel engine ProcIMechE Part D J Automobile Engineering 2004 218

449ndash4646 Zhu D and Zheng X Asymmetric twin-scroll turbochar-

ging in diesel engines for energy and emission improve-ment Energy 2017 141 702ndash714

7 Schmidt S Rose MG Muller M et al Variable asym-metric turbine for heavy duty truck engines In ASME

turbo expo 2013 turbine technical conference and expositionSan Antonio TX 3ndash7 June 2013 ppV05AT23A013-1ndashV05AT23A013-11 New York ASME

8 Brinkert N Sumser S Weber S et al Understanding thetwin scroll turbine flow similarity J Turbomach 2013135 021039

9 Dale A and Watson N Vaneless radial turbochargerturbine performance In Proceedings of the Institution of

Mechanical Engineers 3rd International Conference on

Turbocharging and Turbochargers London May 6ndash81986 Paper No C11086 pp 65ndash76 Mechanical Engi-neering Publications

10 Baines NC and Lavy M Flows in vaned and vaneless sta-tors of radial inflow turbocharger turbines In IMechE

turbocharging and turbochargers conference LondonMay Paper No C405005 pp 7ndash12 Mechanical Engi-neering Publications

11 Baines NC and Yeo JH Flow in a radial turbine underequal and partial admission conditions In IMechE Eur-

opean conference on turbomachinery latest developments

in a changing scene London March 12ndash13 MechanicalEngineering Publications Paper No C423002 pp 103ndash112

12 Baines NC Hajilouy-Benisi A and Yeo JH The pulseflow performance and modelling of radial inflow tur-bines In IMechE conference publications 1994 6 209ndash220 London Medical Engineering Publications

13 Mualler M Streule T Sumser S et al The asymmetric

twin scroll turbine for exhaust gas turbochargers InASME turbo expo 2008 power for land sea and air Ber-lin 9ndash13 June 2008 pp1547ndash1554 New York ASME

14 Wallace FJ and Blair GP The pulsating-flow perfor-mance of inward radial-flow turbines In ASME 1965

gas turbine conference and products show WashingtonDC 28 Februaryndash4 March 1965 New York ASME

15 Wallace FJ Cave PR and Miles J Performance of inwardradial flow turbines under steady flow conditions withspecial reference to high pressure ratios and partialadmission Proc IMechE 1969 184 1027ndash1042

16 Wallace FJ and Miles J Performance of inward radial

flow turbines under unsteady flow conditions with full

and partial admission Proc IMechE 1970 185 1091ndash

110517 Woods WA and Norbury JF Pulse flow performance of

axial turbines for marine turbochargers In ISME Tokyo

1973 Tokyo The Marine Engineering Society of Japan

pp1ndash518 Galindo J Fajardo P Navarro R et al Characterization

of a radial turbocharger turbine in pulsating flow by

means of CFD and its application to engine modeling

Appl Energ 2013 103 116ndash12719 Costall AW McDavid RM Martinez-Botas RF et al

Pulse performance modeling of a twin entry turbocharger

turbine under full and unequal admission J Turbomach

2011 133 02100520 Mamat AMB and Martinez-Botas RF Mean line flow

model of steady and pulsating flow of a mixed-flow tur-

bine turbocharger In ASME turbo expo 2010 power for

land sea and air Glasgow 14ndash18 June 2010 pp2393ndash

2404 New York ASME

21 Rajoo S Romagnoli A andMartinez-Botas RF Unsteady

performance analysis of a twin-entry variable geometry

turbocharger turbine Energy 2012 38 176ndash18922 De Bellis V Marelli S Bozza F et al 1D simulation and

experimental analysis of a turbocharger turbine for auto-

motive engines under steady and unsteady flow condi-

tions Energ Proced 2014 45 909ndash91823 Moustapha SH Kacker SC and Tremblay B An

improved incidence losses prediction method for turbine

airfoils In ASME international gas turbine and aeroengine

congress and exposition Toronto ON Canada 4ndash8 June

1989 ppV001T01A100-1ndashV001T01A100-10 New York

ASME

24 Baines N C (1998) A meanline prediction method for

radial turbine efficiency In Proceedings of the Institution

of Mechanical Engineers 6th International Conference on

Turbocharging and Air Management Systems London

November 3ndash5 Mechanical Engineering Publications

Paper No C554ndash6 pp 315ndash325

Appendix 1

Notation

K total pressure loss coefficientm mass flow rate (kgs)P pressure (Pa)T temperature (K)

g heat ratioh efficiency

Subscripts

0 total parameter1 volute inlet2 volute outlet3 impeller outleth hub sides shroud side

2256 Proc IMechE Part D J Automobile Engineering 233(8)

pulsating inlet condition to a sinusoidal signal whilekept the amplitude and frequency similar to the enginereal exhaust pressure and the prototype pressure waveis from a six-cylinder diesel engine in Mualler et al13

The pulsating frequency of the simulation is 25Hzcorresponding to the power output condition for this 11-L diesel engines This working point demand is morestringent to fuel consumption which means that a highturbine efficiency is required During pulsating inlet con-ditions the volute shroud side scroll and the hub sidescroll face two different inlet pulses that have a reversephase as shown in Figure 15 Considering the two isola-tion inlet conditions of the volute the isentropic efficiencyof the turbine at a specific time is defined as follows

h=m1s +m1h m1s +m1heth THORN T03=T01

m1s 1 (P3=P01s)g1

g

+m1h 1 (P3=P01h)

g1g

eth5THORN

Then the volute total pressure loss coefficient isnewly defined as follows

K=m1sP01s +m1hP01h (m1s +m1h)P02

(m1s +m1h)(P02 P2)eth6THORN

As for rotor efficiency considering that the form ofthe inlet and outlet remains the same as in the steadysituation the calculation remains unchanged

Figure 15 shows the performance of these three tur-bines under a 25-Hz pulsating inlet flow condition The

time-averaged efficiencies for these three turbines are7448 7469 and 7554 corresponding to theasymmetric symmetric and asymmetric OP modelrespectively The asymmetric type has the worst perfor-mance which is a contrary result of the steady perfor-mance where the asymmetric turbine shows the bestperformance This means that the performances of theturbines are different when under steady and pulsatinginlet flow conditions

Under the pulsating inlet conditions the highest effi-ciency points appear for all three models when the huband shroud inlet pressures are equal The efficiencydrops when facing the partial admission conditionespecially in the situation when the hub side inlet hasthe highest back-pressure

When the volute hub side inlet has the lower pres-sure corresponding to less mass flow the efficiencies ofthese three models are nearly the same and the asym-metric turbine is improved over the other twoHowever when the hub side has more pressure theefficiency drops and the asymmetric turbine is affectedthe most causing a decrease in the average efficiencyTherefore the inlet condition at the volute hub side isthe main factor to control the twin-entry turbine per-formance and the decrease in the mass flow may notcause a noticeable rise in the efficiency but the increasein mass flow will lead to an efficiency drop If the twin-entry turbine is symmetric the high efficiency appearswhen the shroud side inlet mass flow rate is larger thanthe hub side and drops when the mass flow is lesswhich agrees with previous conclusions under steadyinlet conditions from the literature

The asymmetric model that places the small scrollon the hub side is affected more by the hub side inletand the OP model that places the big scroll on the hubside is affected less Therefore the small scroll has moreinfluence on the turbine performance In summary thesmall scroll on the hub side is the main cause of deter-mining the turbine performance especially the effi-ciency drop when facing partial admission conditions

Comparing the component performance shown inFigure 16 with the turbine efficiency the same conclu-sion as the steady-state condition was found wherebythe turbine performances are mainly determined by the

Figure 13 Static entropy distribution in the rotor passage(a) asymmetric (b) symmetric and (c) asymmetric OP

Figure 14 Exit flow field at the impeller trailing edge (a) asymmetric (b) symmetric and (c) asymmetric OP

Wang and Zheng 2253

rotor performance and the contribution of the flow lossin the volute is marginally negligible

Figure 17 shows the rotor efficiency and the radialvelocity distribution at the impeller leading edge at timesteps A and B Time step A corresponds to the extremepartial admission condition when all of the turbineshave the efficiency drop It can be seen that the radialvelocity distribution is almost the same for the threemodels at this time The most important factor is theefficiency distribution and it is clear that the OP modelhas the highest efficiency and that the asymmetric onehas the lowest which influences the turbineperformance

Similar to the results under steady conditions theefficiency distribution for the pulsating inlet conditionis affected by the incidence angle at the impeller inletposition especially in the positive incidence regionFigure 18 illustrates that from the asymmetric case tothe OP case the positive incidence region moves fromthe hub side to the shroud side As mentioned beforethe positive incidence angle can sever the vortex in the

impeller flow path and cause the efficiency to dropThe impeller hub side leading edge is more sensitive toa positive incidence angle than the shroud sideTherefore as the positive incidence region moves fromthe hub to the shroud side the efficiency recoverswhich can explain the performance difference in thesethree models

Under equal steady inlet conditions the mass flow inthe big scroll is larger than in the small scroll resultingin a faster flow velocity and positive incidence angleCombined with the conclusion that the hub side of theimpeller is more sensitive to positive incidence if thebig scroll is on the hub side the efficiency drops Underpulsating inlet conditions the efficiency is different tothe steady-state situation especially for the asymmetricturbine when the hub side is charged more than theshroud side This is because the big volute has a highermass flow storage ability than the small one that is thesmall scroll response is more sensitive to the pressurerise and flow in the small scroll is accelerated leadingto a higher outlet velocity which causes a greater inci-dence angle When the pressure rises for the asymmetricturbine with the small scroll on the hub side a positiveincidence angle region occurs at the hub side and theefficiency drops

Conclusion

Simulations for a symmetric and two opposite asym-metric twin scroll turbines under steady and pulsatinginlet conditions have been conducted to study the dif-ferences in the flow field and performance betweenthese three types of turbines an experiment was carriedout to validate the numerical methodology under steadyinlet conditions and the conclusions are as follows

1 Under steady admission conditions the asym-metric turbine with the big scroll on the shroud sidehas the best efficiency which is nearly 16 higherthan the contrary type (OP) Under steady full inletconditions the incidence angle at the region out-side the big scroll is positive The impeller hub sideleading edge is more sensitive to the positive inci-dence than the shroud side due to the hub side vor-tex in the impeller flow path while the positiveincidence severs the vortex and causes the efficiencyto drop Therefore if the big scroll is on the hubside the efficiency is worse than on the shroudside

2 Under pulsating inlet conditions the results arecontrary to the steady state The asymmetric tur-bine with the big scroll on the hub side (OP) hasthe best efficiency which is nearly 11 higherthan on the shroud side Since the small scroll massflow capacity is smaller than the big scroll whenthe backpressure increases the flow in the smallscroll is faster than in the big scroll which maycause a positive incidence angle at the impeller

Figure 15 Turbine performance under the 25-Hz pulsatinginlet condition

Figure 16 Component performance under the 25-Hzpulsating inlet condition

2254 Proc IMechE Part D J Automobile Engineering 233(8)

leading edge Therefore if the small scroll is set onthe hub side the positive incidence appears at thehub side and the efficiency is degraded

3 The design criterion for the asymmetric twin-entryturbine state is as follows when under steady inletconditions (eg constant-pressure turbochargingsystem on the diesel engine) it is better to place thebig scroll on the shroud side When under pulsat-ing inlet conditions (eg pulse turbocharging sys-tem on the gasoline engine) it is better to place thebig scroll on the hub side

4 Simulation calculations were conducted for a givenrotational speed and a given simplified scroll inputpressure It must be controlled whether these con-clusions remain the same for real engine pressureconditions and other engines speeds Experimentswill have to be carried out to check these results

Declaration of conflicting interests

The author(s) declared no potential conflicts of interestwith respect to the research authorship andor publica-tion of this article

Funding

The author(s) received no financial support for theresearch authorship andor publication of this article

Figure 17 Rotor efficiency and radial velocity distribution at the impeller leading edge (a) A time step and (b) B time step

Figure 18 Incidence angle distribution at the impeller leadingedge (a) A time step and (b) B time step

Wang and Zheng 2255

References

1 Durnholz M Eifler G and Endres H Exhaust-gas recir-culation a measure to reduce exhaust emissions of DIdiesel engines SAE technical paper 920725 1992

2 Lapuerta M Hernandez JJ and Gimenez F Evaluationof exhaust gas recirculation as a technique for reducingdiesel engine NOX emissions Proc IMechE Part D J

Automobile Engineering 2000 214 85ndash933 Zamboni G Moggia S and Capobianco M Hybrid EGR

and turbocharging systems control for low NOX and fuelconsumption in an automotive diesel engine Appl Energ2016 165 839ndash848

4 Moustapha H Zelesky MF Baines NC et al Axial andradial turbines vol 2 White River Junction VT Con-cepts NREC 2003

5 Wijetunge RS Hawley JG and Vaughan ND Anexhaust pressure control strategy for a diesel engine ProcIMechE Part D J Automobile Engineering 2004 218

449ndash4646 Zhu D and Zheng X Asymmetric twin-scroll turbochar-

ging in diesel engines for energy and emission improve-ment Energy 2017 141 702ndash714

7 Schmidt S Rose MG Muller M et al Variable asym-metric turbine for heavy duty truck engines In ASME

turbo expo 2013 turbine technical conference and expositionSan Antonio TX 3ndash7 June 2013 ppV05AT23A013-1ndashV05AT23A013-11 New York ASME

8 Brinkert N Sumser S Weber S et al Understanding thetwin scroll turbine flow similarity J Turbomach 2013135 021039

9 Dale A and Watson N Vaneless radial turbochargerturbine performance In Proceedings of the Institution of

Mechanical Engineers 3rd International Conference on

Turbocharging and Turbochargers London May 6ndash81986 Paper No C11086 pp 65ndash76 Mechanical Engi-neering Publications

10 Baines NC and Lavy M Flows in vaned and vaneless sta-tors of radial inflow turbocharger turbines In IMechE

turbocharging and turbochargers conference LondonMay Paper No C405005 pp 7ndash12 Mechanical Engi-neering Publications

11 Baines NC and Yeo JH Flow in a radial turbine underequal and partial admission conditions In IMechE Eur-

opean conference on turbomachinery latest developments

in a changing scene London March 12ndash13 MechanicalEngineering Publications Paper No C423002 pp 103ndash112

12 Baines NC Hajilouy-Benisi A and Yeo JH The pulseflow performance and modelling of radial inflow tur-bines In IMechE conference publications 1994 6 209ndash220 London Medical Engineering Publications

13 Mualler M Streule T Sumser S et al The asymmetric

twin scroll turbine for exhaust gas turbochargers InASME turbo expo 2008 power for land sea and air Ber-lin 9ndash13 June 2008 pp1547ndash1554 New York ASME

14 Wallace FJ and Blair GP The pulsating-flow perfor-mance of inward radial-flow turbines In ASME 1965

gas turbine conference and products show WashingtonDC 28 Februaryndash4 March 1965 New York ASME

15 Wallace FJ Cave PR and Miles J Performance of inwardradial flow turbines under steady flow conditions withspecial reference to high pressure ratios and partialadmission Proc IMechE 1969 184 1027ndash1042

16 Wallace FJ and Miles J Performance of inward radial

flow turbines under unsteady flow conditions with full

and partial admission Proc IMechE 1970 185 1091ndash

110517 Woods WA and Norbury JF Pulse flow performance of

axial turbines for marine turbochargers In ISME Tokyo

1973 Tokyo The Marine Engineering Society of Japan

pp1ndash518 Galindo J Fajardo P Navarro R et al Characterization

of a radial turbocharger turbine in pulsating flow by

means of CFD and its application to engine modeling

Appl Energ 2013 103 116ndash12719 Costall AW McDavid RM Martinez-Botas RF et al

Pulse performance modeling of a twin entry turbocharger

turbine under full and unequal admission J Turbomach

2011 133 02100520 Mamat AMB and Martinez-Botas RF Mean line flow

model of steady and pulsating flow of a mixed-flow tur-

bine turbocharger In ASME turbo expo 2010 power for

land sea and air Glasgow 14ndash18 June 2010 pp2393ndash

2404 New York ASME

21 Rajoo S Romagnoli A andMartinez-Botas RF Unsteady

performance analysis of a twin-entry variable geometry

turbocharger turbine Energy 2012 38 176ndash18922 De Bellis V Marelli S Bozza F et al 1D simulation and

experimental analysis of a turbocharger turbine for auto-

motive engines under steady and unsteady flow condi-

tions Energ Proced 2014 45 909ndash91823 Moustapha SH Kacker SC and Tremblay B An

improved incidence losses prediction method for turbine

airfoils In ASME international gas turbine and aeroengine

congress and exposition Toronto ON Canada 4ndash8 June

1989 ppV001T01A100-1ndashV001T01A100-10 New York

ASME

24 Baines N C (1998) A meanline prediction method for

radial turbine efficiency In Proceedings of the Institution

of Mechanical Engineers 6th International Conference on

Turbocharging and Air Management Systems London

November 3ndash5 Mechanical Engineering Publications

Paper No C554ndash6 pp 315ndash325

Appendix 1

Notation

K total pressure loss coefficientm mass flow rate (kgs)P pressure (Pa)T temperature (K)

g heat ratioh efficiency

Subscripts

0 total parameter1 volute inlet2 volute outlet3 impeller outleth hub sides shroud side

2256 Proc IMechE Part D J Automobile Engineering 233(8)

rotor performance and the contribution of the flow lossin the volute is marginally negligible

Figure 17 shows the rotor efficiency and the radialvelocity distribution at the impeller leading edge at timesteps A and B Time step A corresponds to the extremepartial admission condition when all of the turbineshave the efficiency drop It can be seen that the radialvelocity distribution is almost the same for the threemodels at this time The most important factor is theefficiency distribution and it is clear that the OP modelhas the highest efficiency and that the asymmetric onehas the lowest which influences the turbineperformance

Similar to the results under steady conditions theefficiency distribution for the pulsating inlet conditionis affected by the incidence angle at the impeller inletposition especially in the positive incidence regionFigure 18 illustrates that from the asymmetric case tothe OP case the positive incidence region moves fromthe hub side to the shroud side As mentioned beforethe positive incidence angle can sever the vortex in the

impeller flow path and cause the efficiency to dropThe impeller hub side leading edge is more sensitive toa positive incidence angle than the shroud sideTherefore as the positive incidence region moves fromthe hub to the shroud side the efficiency recoverswhich can explain the performance difference in thesethree models

Under equal steady inlet conditions the mass flow inthe big scroll is larger than in the small scroll resultingin a faster flow velocity and positive incidence angleCombined with the conclusion that the hub side of theimpeller is more sensitive to positive incidence if thebig scroll is on the hub side the efficiency drops Underpulsating inlet conditions the efficiency is different tothe steady-state situation especially for the asymmetricturbine when the hub side is charged more than theshroud side This is because the big volute has a highermass flow storage ability than the small one that is thesmall scroll response is more sensitive to the pressurerise and flow in the small scroll is accelerated leadingto a higher outlet velocity which causes a greater inci-dence angle When the pressure rises for the asymmetricturbine with the small scroll on the hub side a positiveincidence angle region occurs at the hub side and theefficiency drops

Conclusion

Simulations for a symmetric and two opposite asym-metric twin scroll turbines under steady and pulsatinginlet conditions have been conducted to study the dif-ferences in the flow field and performance betweenthese three types of turbines an experiment was carriedout to validate the numerical methodology under steadyinlet conditions and the conclusions are as follows

1 Under steady admission conditions the asym-metric turbine with the big scroll on the shroud sidehas the best efficiency which is nearly 16 higherthan the contrary type (OP) Under steady full inletconditions the incidence angle at the region out-side the big scroll is positive The impeller hub sideleading edge is more sensitive to the positive inci-dence than the shroud side due to the hub side vor-tex in the impeller flow path while the positiveincidence severs the vortex and causes the efficiencyto drop Therefore if the big scroll is on the hubside the efficiency is worse than on the shroudside

2 Under pulsating inlet conditions the results arecontrary to the steady state The asymmetric tur-bine with the big scroll on the hub side (OP) hasthe best efficiency which is nearly 11 higherthan on the shroud side Since the small scroll massflow capacity is smaller than the big scroll whenthe backpressure increases the flow in the smallscroll is faster than in the big scroll which maycause a positive incidence angle at the impeller

Figure 15 Turbine performance under the 25-Hz pulsatinginlet condition

Figure 16 Component performance under the 25-Hzpulsating inlet condition

2254 Proc IMechE Part D J Automobile Engineering 233(8)

leading edge Therefore if the small scroll is set onthe hub side the positive incidence appears at thehub side and the efficiency is degraded

3 The design criterion for the asymmetric twin-entryturbine state is as follows when under steady inletconditions (eg constant-pressure turbochargingsystem on the diesel engine) it is better to place thebig scroll on the shroud side When under pulsat-ing inlet conditions (eg pulse turbocharging sys-tem on the gasoline engine) it is better to place thebig scroll on the hub side

4 Simulation calculations were conducted for a givenrotational speed and a given simplified scroll inputpressure It must be controlled whether these con-clusions remain the same for real engine pressureconditions and other engines speeds Experimentswill have to be carried out to check these results

Declaration of conflicting interests

The author(s) declared no potential conflicts of interestwith respect to the research authorship andor publica-tion of this article

Funding

The author(s) received no financial support for theresearch authorship andor publication of this article

Figure 17 Rotor efficiency and radial velocity distribution at the impeller leading edge (a) A time step and (b) B time step

Figure 18 Incidence angle distribution at the impeller leadingedge (a) A time step and (b) B time step

Wang and Zheng 2255

References

1 Durnholz M Eifler G and Endres H Exhaust-gas recir-culation a measure to reduce exhaust emissions of DIdiesel engines SAE technical paper 920725 1992

2 Lapuerta M Hernandez JJ and Gimenez F Evaluationof exhaust gas recirculation as a technique for reducingdiesel engine NOX emissions Proc IMechE Part D J

Automobile Engineering 2000 214 85ndash933 Zamboni G Moggia S and Capobianco M Hybrid EGR

and turbocharging systems control for low NOX and fuelconsumption in an automotive diesel engine Appl Energ2016 165 839ndash848

4 Moustapha H Zelesky MF Baines NC et al Axial andradial turbines vol 2 White River Junction VT Con-cepts NREC 2003

5 Wijetunge RS Hawley JG and Vaughan ND Anexhaust pressure control strategy for a diesel engine ProcIMechE Part D J Automobile Engineering 2004 218

449ndash4646 Zhu D and Zheng X Asymmetric twin-scroll turbochar-

ging in diesel engines for energy and emission improve-ment Energy 2017 141 702ndash714

7 Schmidt S Rose MG Muller M et al Variable asym-metric turbine for heavy duty truck engines In ASME

turbo expo 2013 turbine technical conference and expositionSan Antonio TX 3ndash7 June 2013 ppV05AT23A013-1ndashV05AT23A013-11 New York ASME

8 Brinkert N Sumser S Weber S et al Understanding thetwin scroll turbine flow similarity J Turbomach 2013135 021039

9 Dale A and Watson N Vaneless radial turbochargerturbine performance In Proceedings of the Institution of

Mechanical Engineers 3rd International Conference on

Turbocharging and Turbochargers London May 6ndash81986 Paper No C11086 pp 65ndash76 Mechanical Engi-neering Publications

10 Baines NC and Lavy M Flows in vaned and vaneless sta-tors of radial inflow turbocharger turbines In IMechE

turbocharging and turbochargers conference LondonMay Paper No C405005 pp 7ndash12 Mechanical Engi-neering Publications

11 Baines NC and Yeo JH Flow in a radial turbine underequal and partial admission conditions In IMechE Eur-

opean conference on turbomachinery latest developments

in a changing scene London March 12ndash13 MechanicalEngineering Publications Paper No C423002 pp 103ndash112

12 Baines NC Hajilouy-Benisi A and Yeo JH The pulseflow performance and modelling of radial inflow tur-bines In IMechE conference publications 1994 6 209ndash220 London Medical Engineering Publications

13 Mualler M Streule T Sumser S et al The asymmetric

twin scroll turbine for exhaust gas turbochargers InASME turbo expo 2008 power for land sea and air Ber-lin 9ndash13 June 2008 pp1547ndash1554 New York ASME

14 Wallace FJ and Blair GP The pulsating-flow perfor-mance of inward radial-flow turbines In ASME 1965

gas turbine conference and products show WashingtonDC 28 Februaryndash4 March 1965 New York ASME

15 Wallace FJ Cave PR and Miles J Performance of inwardradial flow turbines under steady flow conditions withspecial reference to high pressure ratios and partialadmission Proc IMechE 1969 184 1027ndash1042

16 Wallace FJ and Miles J Performance of inward radial

flow turbines under unsteady flow conditions with full

and partial admission Proc IMechE 1970 185 1091ndash

110517 Woods WA and Norbury JF Pulse flow performance of

axial turbines for marine turbochargers In ISME Tokyo

1973 Tokyo The Marine Engineering Society of Japan

pp1ndash518 Galindo J Fajardo P Navarro R et al Characterization

of a radial turbocharger turbine in pulsating flow by

means of CFD and its application to engine modeling

Appl Energ 2013 103 116ndash12719 Costall AW McDavid RM Martinez-Botas RF et al

Pulse performance modeling of a twin entry turbocharger

turbine under full and unequal admission J Turbomach

2011 133 02100520 Mamat AMB and Martinez-Botas RF Mean line flow

model of steady and pulsating flow of a mixed-flow tur-

bine turbocharger In ASME turbo expo 2010 power for

land sea and air Glasgow 14ndash18 June 2010 pp2393ndash

2404 New York ASME

21 Rajoo S Romagnoli A andMartinez-Botas RF Unsteady

performance analysis of a twin-entry variable geometry

turbocharger turbine Energy 2012 38 176ndash18922 De Bellis V Marelli S Bozza F et al 1D simulation and

experimental analysis of a turbocharger turbine for auto-

motive engines under steady and unsteady flow condi-

tions Energ Proced 2014 45 909ndash91823 Moustapha SH Kacker SC and Tremblay B An

improved incidence losses prediction method for turbine

airfoils In ASME international gas turbine and aeroengine

congress and exposition Toronto ON Canada 4ndash8 June

1989 ppV001T01A100-1ndashV001T01A100-10 New York

ASME

24 Baines N C (1998) A meanline prediction method for

radial turbine efficiency In Proceedings of the Institution

of Mechanical Engineers 6th International Conference on

Turbocharging and Air Management Systems London

November 3ndash5 Mechanical Engineering Publications

Paper No C554ndash6 pp 315ndash325

Appendix 1

Notation

K total pressure loss coefficientm mass flow rate (kgs)P pressure (Pa)T temperature (K)

g heat ratioh efficiency

Subscripts

0 total parameter1 volute inlet2 volute outlet3 impeller outleth hub sides shroud side

2256 Proc IMechE Part D J Automobile Engineering 233(8)

leading edge Therefore if the small scroll is set onthe hub side the positive incidence appears at thehub side and the efficiency is degraded

3 The design criterion for the asymmetric twin-entryturbine state is as follows when under steady inletconditions (eg constant-pressure turbochargingsystem on the diesel engine) it is better to place thebig scroll on the shroud side When under pulsat-ing inlet conditions (eg pulse turbocharging sys-tem on the gasoline engine) it is better to place thebig scroll on the hub side

4 Simulation calculations were conducted for a givenrotational speed and a given simplified scroll inputpressure It must be controlled whether these con-clusions remain the same for real engine pressureconditions and other engines speeds Experimentswill have to be carried out to check these results

Declaration of conflicting interests

The author(s) declared no potential conflicts of interestwith respect to the research authorship andor publica-tion of this article

Funding

The author(s) received no financial support for theresearch authorship andor publication of this article

Figure 17 Rotor efficiency and radial velocity distribution at the impeller leading edge (a) A time step and (b) B time step

Figure 18 Incidence angle distribution at the impeller leadingedge (a) A time step and (b) B time step

Wang and Zheng 2255

References

1 Durnholz M Eifler G and Endres H Exhaust-gas recir-culation a measure to reduce exhaust emissions of DIdiesel engines SAE technical paper 920725 1992

2 Lapuerta M Hernandez JJ and Gimenez F Evaluationof exhaust gas recirculation as a technique for reducingdiesel engine NOX emissions Proc IMechE Part D J

Automobile Engineering 2000 214 85ndash933 Zamboni G Moggia S and Capobianco M Hybrid EGR

and turbocharging systems control for low NOX and fuelconsumption in an automotive diesel engine Appl Energ2016 165 839ndash848

4 Moustapha H Zelesky MF Baines NC et al Axial andradial turbines vol 2 White River Junction VT Con-cepts NREC 2003

5 Wijetunge RS Hawley JG and Vaughan ND Anexhaust pressure control strategy for a diesel engine ProcIMechE Part D J Automobile Engineering 2004 218

449ndash4646 Zhu D and Zheng X Asymmetric twin-scroll turbochar-

ging in diesel engines for energy and emission improve-ment Energy 2017 141 702ndash714

7 Schmidt S Rose MG Muller M et al Variable asym-metric turbine for heavy duty truck engines In ASME

turbo expo 2013 turbine technical conference and expositionSan Antonio TX 3ndash7 June 2013 ppV05AT23A013-1ndashV05AT23A013-11 New York ASME

8 Brinkert N Sumser S Weber S et al Understanding thetwin scroll turbine flow similarity J Turbomach 2013135 021039

9 Dale A and Watson N Vaneless radial turbochargerturbine performance In Proceedings of the Institution of

Mechanical Engineers 3rd International Conference on

Turbocharging and Turbochargers London May 6ndash81986 Paper No C11086 pp 65ndash76 Mechanical Engi-neering Publications

10 Baines NC and Lavy M Flows in vaned and vaneless sta-tors of radial inflow turbocharger turbines In IMechE

turbocharging and turbochargers conference LondonMay Paper No C405005 pp 7ndash12 Mechanical Engi-neering Publications

11 Baines NC and Yeo JH Flow in a radial turbine underequal and partial admission conditions In IMechE Eur-

opean conference on turbomachinery latest developments

in a changing scene London March 12ndash13 MechanicalEngineering Publications Paper No C423002 pp 103ndash112

12 Baines NC Hajilouy-Benisi A and Yeo JH The pulseflow performance and modelling of radial inflow tur-bines In IMechE conference publications 1994 6 209ndash220 London Medical Engineering Publications

13 Mualler M Streule T Sumser S et al The asymmetric

twin scroll turbine for exhaust gas turbochargers InASME turbo expo 2008 power for land sea and air Ber-lin 9ndash13 June 2008 pp1547ndash1554 New York ASME

14 Wallace FJ and Blair GP The pulsating-flow perfor-mance of inward radial-flow turbines In ASME 1965

gas turbine conference and products show WashingtonDC 28 Februaryndash4 March 1965 New York ASME

15 Wallace FJ Cave PR and Miles J Performance of inwardradial flow turbines under steady flow conditions withspecial reference to high pressure ratios and partialadmission Proc IMechE 1969 184 1027ndash1042

16 Wallace FJ and Miles J Performance of inward radial

flow turbines under unsteady flow conditions with full

and partial admission Proc IMechE 1970 185 1091ndash

110517 Woods WA and Norbury JF Pulse flow performance of

axial turbines for marine turbochargers In ISME Tokyo

1973 Tokyo The Marine Engineering Society of Japan

pp1ndash518 Galindo J Fajardo P Navarro R et al Characterization

of a radial turbocharger turbine in pulsating flow by

means of CFD and its application to engine modeling

Appl Energ 2013 103 116ndash12719 Costall AW McDavid RM Martinez-Botas RF et al

Pulse performance modeling of a twin entry turbocharger

turbine under full and unequal admission J Turbomach

2011 133 02100520 Mamat AMB and Martinez-Botas RF Mean line flow

model of steady and pulsating flow of a mixed-flow tur-

bine turbocharger In ASME turbo expo 2010 power for

land sea and air Glasgow 14ndash18 June 2010 pp2393ndash

2404 New York ASME

21 Rajoo S Romagnoli A andMartinez-Botas RF Unsteady

performance analysis of a twin-entry variable geometry

turbocharger turbine Energy 2012 38 176ndash18922 De Bellis V Marelli S Bozza F et al 1D simulation and

experimental analysis of a turbocharger turbine for auto-

motive engines under steady and unsteady flow condi-

tions Energ Proced 2014 45 909ndash91823 Moustapha SH Kacker SC and Tremblay B An

improved incidence losses prediction method for turbine

airfoils In ASME international gas turbine and aeroengine

congress and exposition Toronto ON Canada 4ndash8 June

1989 ppV001T01A100-1ndashV001T01A100-10 New York

ASME

24 Baines N C (1998) A meanline prediction method for

radial turbine efficiency In Proceedings of the Institution

of Mechanical Engineers 6th International Conference on

Turbocharging and Air Management Systems London

November 3ndash5 Mechanical Engineering Publications

Paper No C554ndash6 pp 315ndash325

Appendix 1

Notation

K total pressure loss coefficientm mass flow rate (kgs)P pressure (Pa)T temperature (K)

g heat ratioh efficiency

Subscripts

0 total parameter1 volute inlet2 volute outlet3 impeller outleth hub sides shroud side

2256 Proc IMechE Part D J Automobile Engineering 233(8)

References

1 Durnholz M Eifler G and Endres H Exhaust-gas recir-culation a measure to reduce exhaust emissions of DIdiesel engines SAE technical paper 920725 1992

2 Lapuerta M Hernandez JJ and Gimenez F Evaluationof exhaust gas recirculation as a technique for reducingdiesel engine NOX emissions Proc IMechE Part D J

Automobile Engineering 2000 214 85ndash933 Zamboni G Moggia S and Capobianco M Hybrid EGR

and turbocharging systems control for low NOX and fuelconsumption in an automotive diesel engine Appl Energ2016 165 839ndash848

4 Moustapha H Zelesky MF Baines NC et al Axial andradial turbines vol 2 White River Junction VT Con-cepts NREC 2003

5 Wijetunge RS Hawley JG and Vaughan ND Anexhaust pressure control strategy for a diesel engine ProcIMechE Part D J Automobile Engineering 2004 218

449ndash4646 Zhu D and Zheng X Asymmetric twin-scroll turbochar-

ging in diesel engines for energy and emission improve-ment Energy 2017 141 702ndash714

7 Schmidt S Rose MG Muller M et al Variable asym-metric turbine for heavy duty truck engines In ASME

turbo expo 2013 turbine technical conference and expositionSan Antonio TX 3ndash7 June 2013 ppV05AT23A013-1ndashV05AT23A013-11 New York ASME

8 Brinkert N Sumser S Weber S et al Understanding thetwin scroll turbine flow similarity J Turbomach 2013135 021039

9 Dale A and Watson N Vaneless radial turbochargerturbine performance In Proceedings of the Institution of

Mechanical Engineers 3rd International Conference on

Turbocharging and Turbochargers London May 6ndash81986 Paper No C11086 pp 65ndash76 Mechanical Engi-neering Publications

10 Baines NC and Lavy M Flows in vaned and vaneless sta-tors of radial inflow turbocharger turbines In IMechE

turbocharging and turbochargers conference LondonMay Paper No C405005 pp 7ndash12 Mechanical Engi-neering Publications

11 Baines NC and Yeo JH Flow in a radial turbine underequal and partial admission conditions In IMechE Eur-

opean conference on turbomachinery latest developments

in a changing scene London March 12ndash13 MechanicalEngineering Publications Paper No C423002 pp 103ndash112

12 Baines NC Hajilouy-Benisi A and Yeo JH The pulseflow performance and modelling of radial inflow tur-bines In IMechE conference publications 1994 6 209ndash220 London Medical Engineering Publications

13 Mualler M Streule T Sumser S et al The asymmetric

twin scroll turbine for exhaust gas turbochargers InASME turbo expo 2008 power for land sea and air Ber-lin 9ndash13 June 2008 pp1547ndash1554 New York ASME

14 Wallace FJ and Blair GP The pulsating-flow perfor-mance of inward radial-flow turbines In ASME 1965

gas turbine conference and products show WashingtonDC 28 Februaryndash4 March 1965 New York ASME

15 Wallace FJ Cave PR and Miles J Performance of inwardradial flow turbines under steady flow conditions withspecial reference to high pressure ratios and partialadmission Proc IMechE 1969 184 1027ndash1042

16 Wallace FJ and Miles J Performance of inward radial

flow turbines under unsteady flow conditions with full

and partial admission Proc IMechE 1970 185 1091ndash

110517 Woods WA and Norbury JF Pulse flow performance of

axial turbines for marine turbochargers In ISME Tokyo

1973 Tokyo The Marine Engineering Society of Japan

pp1ndash518 Galindo J Fajardo P Navarro R et al Characterization

of a radial turbocharger turbine in pulsating flow by

means of CFD and its application to engine modeling

Appl Energ 2013 103 116ndash12719 Costall AW McDavid RM Martinez-Botas RF et al

Pulse performance modeling of a twin entry turbocharger

turbine under full and unequal admission J Turbomach

2011 133 02100520 Mamat AMB and Martinez-Botas RF Mean line flow

model of steady and pulsating flow of a mixed-flow tur-

bine turbocharger In ASME turbo expo 2010 power for

land sea and air Glasgow 14ndash18 June 2010 pp2393ndash

2404 New York ASME

21 Rajoo S Romagnoli A andMartinez-Botas RF Unsteady

performance analysis of a twin-entry variable geometry

turbocharger turbine Energy 2012 38 176ndash18922 De Bellis V Marelli S Bozza F et al 1D simulation and

experimental analysis of a turbocharger turbine for auto-

motive engines under steady and unsteady flow condi-

tions Energ Proced 2014 45 909ndash91823 Moustapha SH Kacker SC and Tremblay B An

improved incidence losses prediction method for turbine

airfoils In ASME international gas turbine and aeroengine

congress and exposition Toronto ON Canada 4ndash8 June

1989 ppV001T01A100-1ndashV001T01A100-10 New York

ASME

24 Baines N C (1998) A meanline prediction method for

radial turbine efficiency In Proceedings of the Institution

of Mechanical Engineers 6th International Conference on

Turbocharging and Air Management Systems London

November 3ndash5 Mechanical Engineering Publications

Paper No C554ndash6 pp 315ndash325

Appendix 1

Notation

K total pressure loss coefficientm mass flow rate (kgs)P pressure (Pa)T temperature (K)

g heat ratioh efficiency

Subscripts

0 total parameter1 volute inlet2 volute outlet3 impeller outleth hub sides shroud side

2256 Proc IMechE Part D J Automobile Engineering 233(8)