Mixed FLow Turbine

8/10/2019 Mixed FLow Turbine

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Hindawi Publishing CorporationInternational Journal of Rotating MachineryVolume 2012, Article ID 589720,14pagesdoi:10.1155/2012/589720

Research ArticleOn Mixed Flow Turbines for AutomotiveTurbocharger Applications

Bernhardt Luddecke, Dietmar Filsinger, and Jan Ehrhard

IHI Charging Systems International GmbH, Engineering Division, Haberstrae 24, D-69126 Heidelberg, Germany

Correspondence should be addressed to Bernhardt Luddecke,[email protected]

Received 19 December 2011; Accepted 8 June 2012

Academic Editor: Nick C. Baines

Copyright 2012 Bernhardt Luddecke et al. This is an open access article distributed under the Creative Commons AttributionLicense, which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properlycited.

Due to increased demands for improved fuel economy of passenger cars, low-end and part-load performance is of key importancefor the design of automotive turbocharger turbines. In an automotive drive cycle, a turbine which can extract more energy athigh pressure ratios and lower rotational speeds is desirable. In the literature it is typically found that radial turbines provide peakefficiency at speed ratios of 0.7, but at high pressure ratios and low rotational speeds the blade speed ratio will be low and the rotorwill experience high values of positive incidence at the inlet. Based on fundamental considerations, it is shown that mixed flowturbines offer substantial advantages for such applications. Moreover, to prove these considerations an experimental assessmentof mixed flow turbine efficiency and optimal blade speed ratio is presented. This has been achieved using a new semi-unsteadymeasurement approach. Finally, evidence of the benefits of mixed flow turbine behaviour in engine operation is given. Regardingturbocharged engine simulation, the benefit of wide-ranging turbine map measurement data as well as the need for reasonableturbine map extrapolation is illustrated.

1. Introduction

Due to emission legislation, turbocharging of the automotiveinternal combustion engine is becoming common practice.This is not only the case for Diesel engines but also forgasoline engines. Turbocharging the internal combustionengine helps to achieve the required emission levels whilemaintaining suitable driving characteristics. The key require-

ments for new turbochargers are improved performance overa wide operating range while meeting increasingly strictpackaging constraints.

To date radial flow turbines (RFTs) are mostly employedin turbocharger applications for automotive engines. Thispaper describes characteristics about mixed flow turbines(MFTs) leading to the conclusion that such turbines pro-vide considerable advantages for fulfilling the demands inautomotive turbocharger applications. In Figure 1, the mixedflow turbine geometry definition employed throughout thisarticle is illustrated.

A dominant role for the turbine performance is inci-dence, that is, the difference between rotor inlet flow angle

and blade angle at the rotor leading edge. According toJapikse and Baines [1], the optimum incidence for radial tur-bines is in the region of20 to 40. Off-design operationof the turbine is playing a very prominent role in currentturbocharger applications. Due to the intermittent exhaustgas pulse from the reciprocating engine, the turbochargerturbine operates under unsteady admittance. This is evenmore pronounced during engine transients. How the engine

responds during these transients is important in providingdriver satisfaction. Additionally, as combined downspeedingand downsizing concepts are introduced, the exhaust gaspulsation amplitudes increase, while the pulse frequency islowered [2]. Therefore, all technologies that improve theefficiency over a wide range, as well as transient turbinebehavior, are beneficial for turbocharger applications. In thisregard mixed flow turbines, which have been successfullyapplied to modern gasoline applications [36], offer advan-tages. These advantages include a flat efficiency characteristicover a wide range as well as reduced inertia. This studydescribes the principal properties of such kinds of turbines,illustrates a method for evaluating wide map operation,


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2 International Journal of Rotating Machinery

D3.5

s

D4s

D4h

D3.5

h

b3.5

(a)

4s

(b)

Figure 1: Turbine geometry definition.

and provides evidence of the advantages of mixed flowturbines for automotive turbocharger turbines. With regardto quasi-steady as well as transient engine simulation, theadvantage of wide-ranging measurement data is illustrated.Moreover, standard hot gas stand measurement data areusually not sufficient for turbocharged engine simulationand extrapolation is needed. The desire for reasonableturbine map extrapolation methods is pointed out. This isclosely linked to the correct knowledge of the blade speedratio for turbine optimum efficiency. A derivation of thisparameter is made inSection 3.

2. Turbocharger Turbine Design

It is common knowledge that radial turbines have theiroptimum efficiency at a blade speed ratio (u3.5/cs) of around0.7, whereas the peak efficiency for mixed flow turbinesis at lower blade speed ratios (e.g., [7]). For automotiveturbocharger applications, the highest power is available atlowu3.5/cs, and hence the turbine efficiency in this range hasa major influence on the performance of the turbochargedengine. Reducing incidence at low speed ratios can beachieved by reducing the relative rotor inlet flow angle and/orbackward sweeping of the blade leading edge [8].

Theoretically the rotor inlet flow angle can be reducedby increasing A/R of the turbine volute. However underpulsating flow conditions typical of automotive turbochargerapplications this would lead to an excessive dissipationof exhaust gas kinetic energy. This is caused due to thevolume of the turbine volute representing a substantialproportion of the overall volume of the exhaust manifold.Therefore, increasing A/R would have an adverse impacton the turbocharger performance. As a result, pulse tur-bocharging concepts in passenger car applications usuallyapply the smallest possible turbine volutes, hence acceptingan unfavourable inlet flow angle at low speed ratios in orderto exploit the pulse energy.

Most of the radial turbines currently used have radialfibres due to mechanical constraints. The blade inlet angle iszero and the combination with a volute with small A/R leadsto an unfavourable incidence angle. This disadvantage can beavoided by applying mixed flow turbines which allow varia-tion of the blade inlet angle while maintaining radial bladesections (compare Figures2 and8). Therefore the adverseinlet flow angle canto some extentbe compensated bybackward sweeping of the leading edge, avoiding detrimentalincidence.

Mixed flow turbines offer the advantage of additionaldegrees of freedom for aero design compared to radial inflowturbines which usually adopt a radial stacking because ofmechanical constraints. The blade inlet angle of mixed flowturbines can be nonzero even with radial blade sections.Therefore, with mixed flow turbines it is possible to realizemore favourable efficiency characteristics compared to radialturbines with respect to automotive turbocharger applica-tions. Mixed flow turbines can be designed having a lowerinertia which positively contributes to transient response, yetstill maintaining allowable stress limits. Stress levels in theturbine back disc are lower for a mixed flow design whichsupports higher allowable speeds.

Optimum incidence for mixed flow turbines occurs at

lower blade speed ratios u3.5/cs. A similar impact couldbe achieved by backward sweeping the leading edge of aradial turbine. However due to mechanical constraints theamount of backward sweep for radial turbines is very limited[8]. Therefore, until today almost all radial turbines arecharacterized by radial fibres.

The aforementioned considerations are valid for steadyflow conditions. However, turbines for automotive tur-bocharger applications are subject to highly pulsating inletflows. The concept of pulse turbocharging, which is becom-ing increasingly popular, is aiming at optimum utilizationof exhaust pulses through minimum manifold volume. As aconsequence, the instantaneous turbine inlet conditions vary


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International Journal of Rotating Machinery 3

Radial flow turbine

(RFT)

Mixed flow turbine

(MFT)

Figure 2: Radial versus mixed flow turbine.

over a wide range of flow rates. Therefore the developmentfocus is not on achieving optimum design point efficiencybut on achieving a turbine characteristic which offers a highefficiency over a wide range of flow conditions.

Looking at design point efficiencies would lead to theconclusion that radial turbines are superior compared tomixed flow turbines for the specific speeds relevant forautomotive turbocharger applications (e.g., [9]). For highperformance under pulsating operating conditions, theturbine efficiency has to be high for low blade speed ratios.For low blade speed ratios, the combination of high mass

flow with a high efficiency leads to a high power [10]. InSection 3basic considerations are given which support theaforementioned statements about mixed flow turbines.

3. Simple Turbomachinery Fundamentals

Throughout this work the nomenclature shown inFigure 3is adopted. In the automotive industry, index or subscript3 is commonly used for the turbine inlet position (stageinlet) while subscript 4 denotes turbine exit conditions. Forthe conditions at turbine wheel inlet, the subscript 3.5 isintroduced.

Figure 4displays velocity triangles at turbine wheel inletand exit. The general velocity triangle (with inlet swirl) atimpeller inlet is shown. Furthermore, two velocity trianglesat turbine exitone without exit swirl, while the other withexit swirlare shown.

The equivalent (inlet) diameter of a MFT is definedby (1). This value is also used for circumferential velocitycalculation of the mixed flow turbine wheel:

D3.5 =

D23.5h+D

23.5s

2 . (1)

Total-to-static turbine efficiency is defined as:

T,ts =h3t h4t

h3t h4s,is,is=

htt,stage

hts,stage,is. (2)

h

s

3t

3s

3.5s,is

4t,is,is

4s,is,is

3.5t

3.5s

4t,is

4s,is

p3t

p3.5t

p3s

p3.5s

4t

4s

p4t

p4s

Figure 3: Enthalpy-entropy diagram of an expansion process

within a turbine.

The isentropic spouting velocity, which could be achievedif the available total-to-static enthalpy drop would beconverted into kinetic energy by an isentropic process, canbe expressed as

cs2

2 = hts,stage,is. (3)

The blade loading factor is given by

=PT

mTu23.5

=

htt,stage

u23.5=

c3.5 u3.5

u23.5

c4 u4

u23.5. (4)


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Inlet: general velocity triangle

w3.5

3.5

3.5c3.5

u3.5

(a)

Exit:

velocity triangle

withoutexit swirl

w4

u4

c44

(b)

Exit:

velocity triangle

with exit swirl

w4

u4

c44

4

(c)

Figure 4: Velocity triangles at turbine wheel inlet and outlet.

The relationship between isentropic blade loading factor and

the real blade loading factor is derived by

=htt,stage

u23.5

hts,stage,is

hts,stage,is

=

hts,stage,is

u23.5

htt,stage

hts,stage,is= sT,ts.

(5)

If the exit swirl of a single radial or mixed flow turbineis small and circumferential blade velocity at rotor inlet iscomparably higher than at rotor exit, the last term in (4) canbe neglected, leading to

=c3.5

u3.5. (6)

The relationship between stage loading and velocity triangleat rotor inlet is given by

c3.5 = u3.5 cm3.5tan3.5

. (7)

Knowing this, the total to static turbine efficiency can bewritten as

T,ts =htt,stagehts,stage,is

=

u23.5cs2/2

= 2

u3.5cs

2. (8)

The term in brackets is commonly known as turbine bladespeed ratio which is most commonly used for assessing

turbine performance characteristics. Of key importance isthe blade speed ratio at which turbine peak efficiency occurs.

For the ideal case, no losses, no incidence, and negligibleswirl at turbine outlet are assumed.

(i) In absence of losses, the turbine efficiency equalsunity:

T,ts,opt = 1. (9)

(ii) No incidence, negligible swirl at turbine exit (velocity triangle), and perfect flow turning lead to

opt

opt = 1. (10)

Then, the optimum value of blade speed ratio is given by

u3.5cs

opt

=

1

2 = 0.707. (11)

This value of 0.707 is often quoted as the blade speed ratiovalue for optimum efficiency of a radial turbine. In fact, ascan be seen by this derivation, the value of (u3.5/cs)opt isitself a function of maximum turbine efficiency, even if theblade loading factor is constant [11]. Additionally, optimumefficiency of a radial turbine occurs at loading factors belowone. This means that the flow is approaching the rotor bladeswith positive incidence as shown inFigure 5.

The definition of incidence is given by

i3.5 =3.5 b3.5. (12)

By combination of (6) and (7), the loading is defined by

=

1

cm3.5

u3.5 tan3.5

. (13)

Assuming a constant ratio ofcm3.5/u3.5 of 0.5, by changingthe incidence angle from 0 to 25, the blade loading foroptimum incidence reduces to

opt,RFT = 0.77. (14)

Furthermore assuming a peak efficiency of 0.7, the optimumblade speed ratio is calculated to

u3.5cs

opt,RFT,real

=

T,ts2

=

0.7

2 0.77 0.67. (15)

Hence for these parameters and also for a radial turbine, theactual optimum efficiency is occurring at a blade speed ratiobelow 0.707!

Considering mixed flow turbines, the consequences areas follows. For simplification, the characteristics of a MFTare explained assuming a RFT with back sweep as depictedinFigure 6.

The optimum flow turning within the backswept rotor isachieved for negative incidence, which in fact means almostradial inflow. Thus3.5 0. The corresponding loading foroptimum incidence is hence calculated to

opt,MFT = 1. (16)


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3.5 = 20

b3.5 = 0

i3.5 = 20

c3.5

u3.5

w3.5

SSSSPS

PS

0 >

3.5 = 0

b3.5 = 0

i3.5 = 0

c3.5

u3.5

w3.5

SSSSPS

PS

0 >

3.5 = 20

b3.5 = 0

i3.5 = 20

c3.5

u3.5

w3.5

SS

SSPS

PS

0 >

3.5 = 40

b3.5 = 0

i3.5= 40

c3.5

u3.5

w3.5

SS

PS SS

PS

0 >

Figure5: Radial flow turbine inflow characteristics for different incidence angles.

Assuming the same overall maximum turbine efficiency, theincreased blade loading for optimum turbine efficiency leadsto a shift in (u3.5/cs)optaccording to

u3.5cs

opt,MFT,real

=

T,ts2

=

0.7

2 1 0.59. (17)

This relationship between blade speed ratio and turbine effi-ciency is illustrated inFigure 7. The independent parameterfor the two plotted curves is blade loading. The two examplesderived in the previous section as well as the often quotedvalue of (u3.5/cs)opt = 0.707 are highlighted.

The fact that optimum incidence is achieved at higher

blade loading for mixed flow turbines has been reportedby several authors. Even optimum blade loading valuesexceeding unity have been reported (e.g., [12, 13]). Thereason why mixed flow turbines behave like radial flowturbines with back sweep is illustrated inFigure 8. When amixed flow turbine wheel is approached by a flow vectorperpendicular to its inlet edge (green vector), a trianglebetween the radial and the actual flow direction can be drawn(lines: green, white, red). In other words, this means thatthe flow is not purely radial but has an axial component. Byadditionally leaning the blade with a nonzero rake angle, thistriangle is rotated (lines: blue, white, red). The bottom lineis that the resulting flow vector that approaches the blade is

the one drawn in blue. While maintaining the radial stackingconstraint, the effective flow angle of a radial flow is changedto be nonradial by adding an axial component. If an MFT isapproached by a purely radial vector (red vector), the effectdescribed above is not occurring.

For the radial rotor this means that due to the absenceof an axial component in the vector approaching the rotor,the flow is by nature purely radial (red vector). The vectortriangle described above is not established, and thus evenwhen designing a nonzero rake angle, the mixed-flow-effect is not achievable.

It should be emphasized that with regard to mechanical

stress constraints, the distinct advantage of a mixed flowwheel over a radial flow wheel is that this nonzero blade inletblade angle is achieved without violation of radial stackingcondition. In addition to this, mixed flow turbine wheelsoffer the chance to design turbine wheels with reducedinertia. One key benefit is that the back disk is clearly reducedin diameter.

An analytical relationship between cone angle (), rakeangle (), and blade angle (b) is given by

tanb= cos() tan

. (18)

These theoretical considerations of turbine characteristicsare in agreement with measurements and supported by


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c3.5

u3.5

w3.5

SS SSPS

PS

0 >

3.5 = 20

b3.5 = 02

i3.5 = 40

c3.5

u3.5

w3.5

SSSSPS

PS

0 >

3.5 = 0

b3.5 = 02

i3.5 = 20

c3.5

u3.5

w3.5

SSSSPS

PS

0 >

3.5 = 20

b3.5 = 02

i3.5 = 0

c3.5

u3.5

w3.5

SSSSPS

PS

0 >

3.5 = 40

b3.5 = 02

02 i3.5 =

Figure 6: Backswept RFT/MFT inflow characteristics for diff

erent incidence angles.

several studies (e.g., [7, 13]). For example,Figure 9 showsa comparison of radial and mixed flow turbine efficiencyversus blade speed ratio. The shift to lower values of bladespeed ratio can clearly be recognized.

This characteristic of mixed flow turbines is desirableespecially for the requirements of automotive turbochargerapplications. Under pulsation conditions, the maximumexhaust gas enthalpy is available for high pressure ratios,occurring directly after exhaust valve opening. As turbinespeed does only slightly changeif at allduring an engine

cycle, the high pressure ratio results in low values ofu3.5/cs.Therefore, it is important to have high efficiencies for theseconditions [10].

In the current work the enthalpy-based definition of thedegree of reaction is adopted:

R =hss,rotorhtt,stage

=h3.5s h4sh3t h4t

. (19)

Assuming that within the stator, that is, from station 3 tostation 3.5, no losses occur (e.g.,h3t h3.5t = 0), (19) canbe rearranged to

R =

h3.5t c

23.5/2

h4t c

24/2

h3t h4t

= 1 c23.5 c

24

2(h3t h4t).

(20)

Together with (4) the velocity triangles shown inFigure 4and the assumptions that the investigated turbine stage hasno losses (h3.5t = h4t), negligible exit swirl (c4 = 0), andcr3.5 = cz4 (no diffusion, no acceleration within a radialrotor), (20) can be rearranged to

R=

1

c23.5 c24

2c3.5 u3.5

= 1 c

23.5+ c

2r35 c

24+ c

2z4

2c3.5u3.5

= 1 c3.5

2u3.5= 1

2 .

(21)

From this, a relationship between (u3.5/cs)optand the degreeof reaction, depending on the turbine efficiency, is derived:

u3.5cs

opt

=

T,ts

4(1 R). (22)

In Figure 10 the interdependency of these parameters isplotted. The highlighted symbol in the graph shows again


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= 1

= 0.77

0.9

0.8

0.7

0.6

0.5

0.40.4 0.5 0.6 0.7 0.8 0.9 1 1.1

T,ts

(u3.5/cs

)opt

[]

[

]

Figure7: Optimumu3.5/cs versus assumed stage efficiency for twodifferent blade loadings.

the frequently quoted optimum value of 0.707 as describedbefore. Furthermore, the graph shows clearly that the opti-mum values ofu3.5/csversus the degree of reaction for non-ideal efficiencies can be expected to be lower in real turbineconfigurations. Depending on the degree of reaction of thechosen turbine stage design, the optimum blade speed ratiocan be modified. Moreover, in turbocharger applicationsincluding fixed turbine geometry configurationsit hasto be considered that the variation of inlet pressure overtime caused by the intermittent exhaust pulses from thereciprocating engine do cause a variation in degree ofreaction during operation. This means that there will not beone fixed value of optimum blade speed ratio that describesthe on-engine turbine behaviour (compare [14]).

Figure 11shows the measured mass flow parameter andefficiency versus expansion ratio of a mixed flow turbinein comparison with an equivalent radial turbine. The datawas obtained on a hot gas stand under quasi-steady flowconditions and is subject to heat flow effects typical formeasurements of small turbocharger turbines. This graph

already indicates the beneficial characteristics of mixed flowturbines.

From the theoretical analysis performed above, there issufficient reason for investigating the behaviour of turbinesover a very wide operating range in more detail.

A potential procedure for doing so is described in thefollowing Section.

4. Steady Wide Mapping Results fora Mixed Flow Turbine

This study was undertaken for a small mixed flow turbinefor automotive gasoline engine application. In [15], a new,

simple method for wide mapping by variation of turbineinlet temperatures has been presented. By minimising theinfluence of heat flows, a quasi-adiabatic turbine map wasevaluated from measured data. The simple heat transfermodel introduced in [15] showed good agreement with theapproaches of other authors [16]. The resulting contour map

after heat transfer correction is shown inFigure 12. This mapis not corrected for friction losses in the bearing system.The turbine pressure ratio is plotted versus blade speed

ratio. The colour and the isolines separate areas of same totalto static turbine efficiency. Optimum efficiency is achievedwithin a range ofu3.5/cs between 0.56 and 0.62, thus muchlower than usually cited in the literature [9].

Furthermore it can be seen that according to theory theoptimum blade speed ratio where the turbine offers bestefficiency is increasing with pressure ratio.

PIT,ts can (amongst others) be interpreted as a flowcoefficient and gives information about exit dynamic head.

PIT,ts and u3.5/cs are also related to several loss mech-anisms within the stator and rotor. Therefore it is justifiedto assess stage performance (more precisely: efficiency) usingthese parameters.

For a fixed geometry (wastegated) turbine, efficiencyis only depending on flow coefficient, loading coefficient,and Reynolds number. MFP and pressure ratio are stronglycoupled. As the turbine exit pressure is typically almostambient pressure for hot gas stand tests, PIT,ts also givesdirect information about exit dynamic head, when turbineinlet temperature is fixed. The blade speed ratio can beinterpreted as loading coefficient.

5. Semi-Unsteady Turbine Efficiency

Measurement Approach for Wide Mapping ofa Mixed Flow Turbine

Based on the steady wide mapping results shown above, anew, instantaneous method for measurement of efficiency atvery low values ofu3.5/cs was developed. The so-called highinertia rotor (HIR) approach of the IHI Charging SystemsInternational (ICSI) is based on acceleration measurementof a rotor which has a significantly higher inertia than thestandard turbocharger rotor. To achieve this, the compressorwheel was replaced by a bladeless, zero work, high inertiaimpeller as shown inFigure 13.

By measuring the instantaneous speed and utilizing the

known rotor inertia, the instantaneous turbine accelerationpower is directly evaluated. This instantaneous accelerationpower is compared against an almost constant isentropicenthalpy difference, generated by the hot gas burner ofthe test bench. As mass flow was measured and heldconstant, temperature at turbine inlet was controlled bysetting the heating unit to constant power. This was re-checked by the temperature measurement. However, theapplied thermocouples were not fast response and could notdetermine accurately the temperature with respect to timedue to their thermal inertia. Pressure measurements weredone with fast response pressure transducers to controlwhether pressure ratio varies during acceleration and to


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b

Figure 8: Degrees of freedom for mixed flow and radial flow turbines.

0.9

0.8

0.7

0.6

0.5

0.2 0.4 0.6 0.8

MFT

RFTEfficiency

u/cs

PIT,ts = 1.5 to 3

Figure9: Turbine efficiency as a function of blade speed ratio forradial and mixed flow turbines [7].

allow for a phase correction between the pressure beforeand after the turbine. This was done with calibrated piezo-resistive absolute pressure transducers [17]. The absolutepressure value of the fast transducers was cross-checked, withthe signal of the standard slow response pressure sensorsignal before the HIR acceleration was started. A picture ofthe test-bench setup is shown inFigure 14.

Turbine housing as well as all hot gas and measurementpipes as insulated to minimize heat transfer between the

turbocharger and the test cell environment. Furthermore,turbine inlet temperature, oil conditioning, and water cool-ing were set to constant low values to minimize heat transfer.Initially, the HIR is locked and the desired turbine pressureratio and turbine inlet temperature are set. After starting ofthe transient measurement system, the rotor is released andaccelerates. An automatic shutdown procedure is applied toprevent overspeed of the HIR.

Regarding maximum rotational speed, two main aspectshave to be considered.

(i) Firstly maximum allowable speed must not beexceeded, to avoid any damage of the HIR itself aswell as the bearing system. It is clear that the rotordynamics of such a HIR system are very differentfrom a conventional turbocharger rotor.

(ii) Secondly, the speed has to be low enough; that is,elastic deformation of the rotor does not change themoment of inertia. Otherwise, the speed signal couldnot be used for turbine net power measurementduring acceleration of the rotor.

A typical result of the instantaneous measurements, for aconstantPIT,ts = 1.4, is shown inFigure 15.

The calculated values for torque, power, and efficiency forvery low values ofu3.5/cs are not reliable. This is indicated


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1.2

1.1

1

0.9

0.8

0.7

0.6

0.5

0.4

0.3

0.20 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1

Degree of reaction

= 1

(u3.5/cs

)opt

T,ts = 0.4T,ts = 0.5

T,ts = 0.6

T,ts = 0.7T,ts = 0.8

T,ts = 1

()

()

Figure 10: Optimum u3.5/cs versus degree of reaction in depen-dency on turbine efficiency.

on the very left side of the graph. The journal bearingsare starting to rotate and the oil film is developing. Thusvalues of u3.5/cs lower than 0.08 have to be omitted. Thehighest possible value of blade speed ratio is limited by themaximum rotational speed of the HIR and depends on the

applied turbine pressure ratio. The higher the desiredPIT,ts,the lower the maximum of the u3.5/cs that can be achieveddue to stress limitations.

The evaluation procedure for instantaneous torque,power, and efficiency is given as follows.

The rotor acceleration is calculated by

(t) =(t)

t = 2

nT(t) nT(t 1)

t

1

60. (23)

The instantaneous torque can then be calculated, if rotorinertia is known. The rotor inertia of the HIR is about 28times higher compared to a conventional turbocharger rotor:

T(t) =J (t). (24)

Instantaneous turbine power can then be calculated accord-ing to

PT,tm(t) = T(t) (t). (25)

The instantaneous power has to be compared with the almostconstant burner power or ideal total-to-static enthalpy flow:

Pts,is(t) = mgas(t) cp,gas(t)T3(t)

1

1

T,ts(t)

(g(t)1)/g(t)

.

(26)

0.05

0.2

0.1

Expansion ratio

MFP/e5msk0

.

5

Turbineefficiency

Radial flow turbine (RFT)

Mixed flow turbine (MFT)

()

()

Figure 11: Turbine performance comparison: mixed flow versusradial flow turbine.

4.5

4

3.5

3

2.5

2

1.5

1

0.4 0.45 0.5 0.55 0.6 0.65

=

0.0

1

u/cs

PIT,ts

()

()

Figure12: Mixed flow turbine efficiency contour plot [15].

The instantaneous thermomechanical total-to-static turbineefficiency is then defined by

T,ts,tm(t) =PT,tm(t)

Pts,is(t). (27)

These semi-unsteady results are compared with the resultsof the steady wide mapping results. To do this, the contour


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Figure13: High inertia rotor (HIR) assembly.

Figure14: Test setup for unsteady turbine performance measure-ment.

plot ofFigure 12 is intersected at a pressure ratio of 1.4.Then a single curve of efficiency as a function of blade speedratio obtained. This curve is compared to the results already

shown inFigure 15. Additionally, to judge the quality of theinstantaneous measurement and of the extrapolation, the so-called runaway speed has also been measured (Figure 16).This was done as described by Smiljanovski et al. [18],and the corresponding measured value is also includedin Figure 17. A zero-friction impeller (ZFI) replaces thecompressor wheel, and the resulting speed that is measuredfor different turbine pressure ratios is the speed, whereturbine power and friction power are equal. The results ofthe runaway speed measurements for two different turbineinlet temperatures and several pressure ratios are presentedinFigure 16. It can be seen that for pressure ratios higherthan 1.6 the runaway speed remains constant.

300

Poweracceleration/(W)

0.1

torque/(Nm)

T,tm,ts

()/

0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1 1.1

u/cs (blade speed ratio) ()

HIRm

ax.speedlimit

Lowspee

dlimit

Acceleration torque

Power acceleration

Instantaneous turbine efficiencyExtrapolation of instantaneous turbine efficiency

Limits due to min. and max. speed

Figure15: Typical result of an unsteady HIR measurement.

1.2

1

0.8

0.6

0.4

0.2

0

1 1.25 1.5 1.75 2 2.25 2.5

PIT,ts

u/cs

(bladespee

dratio

)

(u/cs)runaway@T3 = 20C

(u/cs)runaway @T3 = 400C

()

Figure 16: Runaway speed measurements for two turbine inlettemperatures.

For a pressure ratio of 1.4 and a turbine inlet temperature

of 20C, a runaway blade speed ratio of about 1.04 wasrecorded.

From analysingFigure 17, it can be stated that the steadyand unsteady results give a consistent picture of turbineefficiency characteristics. It also proves the values of bladespeed ratio, where optimal efficiency is reached.

However, some deviation exists which can be explained.The steady results have been collected by the so-calledturbine net efficiency approach [19], using the measuredcompressor power to calculate thermo-mechanical turbineefficiency. Heat transfer effects have been corrected by asimple heat transfer model [15], which in general givesvery reasonable results in turbine efficiency trend. Opposed


11/15


0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1 1.1

u/cs (blade speed ratio)

Acceleration torque

Power acceleration

Instantaneous turbine efficiencyExtrapolation of instantaneous turbine efficiencyInterpolated static measurement dataMeasured runaway point

()

Poweraccele

ration/(W)

300

0.1

torque/(Nm

)

T,tm,ts

()/

Figure17: Comparison of steady and unsteady test results.

to this, the unsteady approach does not need a heattransfer correction, as measurements have been carried outat very low turbine temperatures and the turbine powermeasurement is done by measuring acceleration power. Butas already mentioned, for this approach, no compressorwheel exists, and hence the axial thrust is different comparedto the steady-state wide map efficiency measurements. Thishas an impact on bearing losses and thus thermomechanicalturbine efficiency.

Regarding unsteady turbine operation, it is to note thatdue to the almost constant turbine pressure ratio duringacceleration, no filling and emptying effects within theturbine scroll have to be expected. Thus, although this is anunsteady measurement, the problems that are encounteredduring efficiency measurement under pulsed conditions areovercome. Thus, the experimental approach presented hereis labelled semi-unsteady.

6. Turbine Performance under PulsatingFlow Conditions

Due to the exhaust gas pulse from the intermittent opera-

tion of the reciprocating engine, the turbocharger turbineoperates under unsteady admittance. In Figure 18 the turbineoperation during a typical engine cycle is shown. The greenfilled diamonds represent the available measured data points.The solid blue lines show the data fitted extrapolation,and the red line gives the unsteady turbine operationduring engine cycle. The figure illustrates that usually thelimited measured data has to be extrapolated far beyond theavailable range and underlines the importance of accurateextrapolation techniques, what is consistent with findings in[20,21].

Thus, the correct prediction of engine steady operation,as well as unsteady operation or even vehicle acceleration

0.05

0.2

T,ts

PIT,ts

Standard hot gas test bench data

Extrapolated hot gas test bench data

Turbine operation during engine cycle simulation

()

()

Figure18: On-engine turbine operation.

behavior, is strongly depending on sensible and correctextrapolation.

A wide map measurement of turbine data can help toavoid the need for extrapolation. However, usually a widemap measurement is not available. A wide map measurementcan be used to develop improved extrapolation algorithms.In general the turbine efficiency data to be extrapolatedshould not contain friction influence from the bearingsystem. Friction is not related to aerodynamic parameters butto real shaft speed as well as to thrust load.

However, standard hot gas stand turbine efficiency datausually contains friction data due to the measurementmethod. Thus, modelling of bearing friction depending onshaft speed, and thrust load can also be a source of error forextrapolation.

In this section it is investigated how various turbinecharacteristics and designs affect the on-engine operation.From the above sections it is known that the efficiencycharacteristics of mixed flow turbines can be advantageousfor automotive turbocharger applications.

Figure 19 illustrates the calculated increase of boostpressure versus time using the in-house engine simulationcode ITES (IHI Turbocharged Engine Simulation) by Ikeyaet al. [22].

This simulation program is capable of predicting tur-

bocharged engine performance under steady and transientconditions. ITES is focused on the detailed modelling andnumerical description of the turbocharger. The study showninFigure 19aims to identify turbine configurations whichare advantageous for transient operation. The study wasperformed on a representative four-cylinder gasoline enginefor passenger cars with a load step at 1500 rpm. As alreadymentioned, transient operation is of major importance sincesteady-state operation of the turbocharger is basically notexistent. This is even more pronounced since speeding upa vehicle from stand still as well as satisfactory accelerationduring vehicle drives are the most important events inevaluating drivers satisfaction. Therefore, all technologies


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that improve the transient turbine behaviour are beneficialfor turbocharger applications.

All results are compared to a mixed flow turbine rotormade from conventional nickel-base alloy. This is referred toas the base configuration.

The effects that have been investigated are inertia and

turbine effi

ciency. A clear advantage can be seen whencomparing the base configuration with a turbine variantmade from gamma titanium aluminide (-TiAl). Since thedensity of this material is much lower compared to nickel-base alloys, the rotor inertia is reduced and hence theacceleration of such a turbine is notably improved and helpsto increase the boost pressure rise. It has to be mentionedthat this advantage might be compromised due to moresevere manufacturing constraints. Due to the unfavourablecastability of-TiAl and its lower ductility, it is susceptible toforeign object damage. This almost inevitably compromisesthe aerodynamic design. For example, inFigure 20two tur-bine wheel designsone for nickel-base alloy and one for -TiAlwith identical swallowing capacity are compared. Thegrey wheel represents the standard nickel-base alloy design,while the superimposed red single blade shows the -TiAldesign that requires higher material thickness. To achieve thesame flow capacity, the blade angle distribution needs to bemodified for readjustment of the throat area of the wheel.

The -TiAl turbine wheel from this comparison still has a46% lower polar moment of inertia compared to the nickel-base alloy wheel. Regarding the rotor assembly (turbinewheel and shaft plus compressor wheel), this advantagereduces but still is about 30%. In the predicted values showninFigure 20, because of the aforementioned -TiAl designconstraints a turbine efficiency penalty of 5% is assumed.Any benefit achieved by the reduced inertia is offset by thedrop in efficiency. Additionally, steady-state engine brakespecific fuel consumption will deteriorate as a result ofdecreased turbine efficiency. Consequently, depending onindividual design requirements, using this material might beneither favourable nor desirable for certain applications.

Please note that the simulation results also depend onthe investigated engine load step and especially the startingconditions of the load step for example, acceleration fromstand still profits more from reduced inertia.

The effect of modifying the characteristics of the turbinemap, the main topic of the current work, was also investi-gated. Compared to the base configuration, a variant withreduced (u3.5/cs)optvalues was simulated. The peak efficiency

of such a (mixed flow) turbine is not necessarily increased asindicated byFigure 7, but the map area where peak efficiencyis reached is modified. It can clearly be noted that thismodification of the turbine stage behaviour is of benefitfor providing quick boost pressure rise. Similar findings arereported in [23]. Of course, a turbine variant offering bothoptimized map and low inertiashows the best transientresponse.

7. Summary and Conclusion

Due to increasingly higher demands for improved fuel econ-omy of passenger cars, low-end and part-load performance is

0.5 s

Boostp

ressure

Time

Base configuration

Base + TiAl

Base + TiAl, 95% T-effy

(u3.5/cs)opt


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positive incidence at the inlet. The present study shows thateven for radial turbines the blade speed ratio where optimumefficiency is reached is usually lower than the commonlyquoted blade speed ratio of 0.7. The present work givestheoretical justification and experimental evidence that formixed flow turbines optimum efficiency can be obtained at

even lower blade speed ratios. This can be attributed to amore favourable inlet blade angle, swallowing capacity andinertia when compared to a radial design.

The present study shows that mixed flow turbines havekey advantages for automotive turbocharger applications asthey have improved performance at low blade speed ratios.This means that a significant portion of the pulse energyavailable in the exhaust gas can be utilized. The behaviourof a mixed flow turbocharger turbine was investigated bysteady-state wide mapping and also by employing a new,semi-unsteady measurement approach. It was found thatthe unsteady approach shows very good agreement withthe steady and runaway measurements. It was theoreticallyderived that the blade speed ratio for optimum efficiencyof a mixed flow turbine is far below the commonly citedvalue of 0.7. This was also proven experimentally. Finally aninvestigation of how this could improve on-engine behaviourwas described. The benefit of low-inertia mixed flow tur-bocharger turbine wheels has been clearly demonstrated.

Nomenclature

p: Static pressure (Pa)ptot: Stagnation pressure (Pa)u3.5/cs: Blade speed ratio ()PI(also): Pressure ratio ()c: Velocity in stationary frame (m/s)w: Velocity in rotating, relative frame (m/s)u: Blade speed (m/s)h: Enthalpy difference (J/kg)C: CompressorD: Diametercs: Isentropic spouting velocity (m/s)R: Degree of reaction (-)i: Incidence (deg,)m: Mass flow rate (kg/s)n: Rotational velocity (1/s)t: Time difference (s)t: Time (s)

P: Power (W)T: Torque (Nm)h: Enthalpy (J/(kg K))s: Entropy (J/(kg K)).

Abbreviations

CFD: Computational fluid dynamicsMFP: Mass flow parameterRFT: Radial flow turbineMFT: Mixed flow turbine.

Greek Symbols

: Cone angle (deg) : Rake or camber angle (deg)b: Blade angle (deg) : Relative flow angle (deg)

: Absolute flow angle (deg) : Loading coefficient (): Simplified loading coefficient () : Rotational speed (rad/s).

Indices

opt: Optimums: Statict: Totalz: Axial, inz-directionr: Radial

is: Isentropicss: Static to statictt: Total to totalts: Total to static3: Turbine stage inlet3.5: Turbine wheel inlet4: Turbine (wheel) exitm: MeridionalT: Turbine: Circumferential: Acceleration.

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Concepts ETI, 1997.[2] R. Golloch, Downsizing bei Verbrennungsmotoren, Springer,

2005.[3] P. Luckert, F. Kreitmann, N. Merdes et al., The new 1. 8-

litre 4-cylinder petrol engine with direct injection and turbocharging for all passenger cars with standard drive trains fromMercedes-Benz, inWiener Motorensymposium, 2009.

[4] D. Hagelstein, L. Hentschel, S. Strobel et al., Die Aufladeen-twicklung fur den Neuen 1.2l TSI Motor von Volkswagen,Aufladetechnische Konferenz, Dresden, Germany, 2009.

[5] F. Baumel, J. Jedro, C. Weber, A. H. Hinkelmann, A. Mayer,and C. Kirschner,The Turbocharger of the Third Generation of

the AUDI R4-TFSI Engines Using the Example of the New 1. 8lTFSI, Aufladetechnische Konferenz, Dresden, Germany, 2011.

[6] N. Merdes, C. Enderle, G. Vent, and R. Weller, Der neuevierzylinder-ottomotor mit turbo-aufladung von Mercedes-Benz, Motortechnische Zeitschrift 12/2011, 72. Jahrgang, pp.942-951, 2011.

[7] N. Watson and M. S. Janota, Turbocharging the InternalCombustion Engine, The MacMillan Press, 1982.

[8] J. Walkingshaw, S. Spence, J. Ehrhard, and D. Thornhill,An investigation into improving off-design performancein a turbocharger turbine utilizing non-radial blading, inProceedings of the ASME Turbo Expo Conference, 2011.

[9] N. C. Baines, Fundamentals of turbocharging, ConceptsNREC, Edward Brother Incorporated, 2005.


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[10] K. Zinner, Aufladung von Verbrennungsmotoren, Springer,Auflage, Germany, 1985.

[11] M. Abidat, H. Chen, N. C. Baines, and M. R. Firth, Designof a highly loaded mixed flow turbine, Proceedings of theInstitution of Mechanical Engineers Part A, vol. 206, no. 2, pp.95107, 1992.

[12] H. Chen and N. C. Baines, The aerodynamic loading of radial

and mixed-flow turbines,International Journal of MechanicalSciences, vol. 36, no. 1, pp. 6379, 1994.

[13] S. Rajoo and R. Martinez-Botas, Mixed flow turbine research:a review,Journal of Turbomachinery, vol. 130, no. 4, Article ID044001, 12 pages, 2008.

[14] D. Filsinger, G. Fitzky, and B. Phillipsen, Flexible tur-bocharger turbine test rig MONA VI, inProceedings of the 8thInternational Conference on Turbochargers and Turbocharging(IMechE 06), pp. 207222, May 2006.

[15] B. Luddecke, D. Filsinger, and M. Bargende, On wide map-ping of a mixed flow turbine with regard to compressor heatflows during turbocharger testing, in Proceedings of the 10thInternational Conference on Turbochargers and Turbocharging(IMechE 11), pp. 185202, 2011.

[16] M. V. Casey and T. M. Fesich, On the efficiency of compres-sors with diabatic flows, in Proceedings of the ASME TurboExpo Conference, pp. 785797, Orlando, Fla, USA, June 2009.

[17] Kistler Instrumente AG Winterthur, CH-8408 Winterthur,Switzerland, Pressure Sensors Type 4045 and 4075,http://www.kistler.com/mediaaccess/4045A10 000-064m-10.92.pdf.

[18] V. Smiljanovski, J. Scharf, N. Schorn, S. Pischinger, and B.Funken, Messung des Turbinen-Wirkungsgrads bei NiedrigenDrehzahlen, Aufladetechnische Konferenz, Dresden, Germany,2008.

[19] S. Scharf, Extended turbocharger mapping and engine simula-tion [Dissertation], RWTH, Aachen, Germany, 2010.

[20] A. Pesiridis, W. S. I. W. Salim, and R. F. Martinez-Botas,Turbocharger matching methodology for improved exhaustgas energy recovery, in Proceedings of the 10th InternationalConference on Turbochargers and Turbocharging (IMechE 12),pp. 203218, 2012.

[21] N. Baines and C. Fredriksson, The simulation of tur-bocharger performance for engine matching, Motorprozess-simulation und Aufladung, 2, pp. 101-111, 2007.

[22] N. Ikeya, H. Yamaguchi, K. Mitsubori, and N. Kondoh,Development of advanced model of turbocharger for auto-motive engines, SAE Paper 920047, 1992.

[23] P. Scheller, C. Schnuckel, C. Jordens, D. Hagelstein, and J.Theobald,Evaluation Study of Different Turbine Wheel Geome-tries for Turbocharged Gasoline Engines, AufladetechnischeKonferenz, Dresden, Germany, 2011.


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Mixed FLow Turbine