twin entry

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Unsteady performance analysis of a twin-entry variable geometry turbocharger

turbine

Srithar Rajoo a, Alessandro Romagnoli b, Ricardo F. Martinez-Botas b,*

a Transportation Research Alliance, Universiti Teknologi Malaysia, 81310 UTM Skudai Johor, Malaysiab Department of Mechanical Engineering, Imperial College London, Exhibition Road, London SW7 2AZ, UK

a r t i c l e i n f o

Article history:

Received 10 May 2011

Received in revised form

13 December 2011

Accepted 15 December 2011

Available online 17 January 2012

Keywords:

Turbochargers

Aerodynamic efciency

Unsteady ows

Turbines

a b s t r a c t

This paper discusses the details of unsteady experimentation and analysis of a twin-entry variable

geometry turbine for an automotive turbocharger. The turbine in this study is the product of design

progression from a commercial nozzleless unit to a single-entry variable geometry and consequently to

a twin-entry unit. The main features of the turbine were kept similar across all congurations for

equivalent comparison basis. The unsteady curves of the twin-entry turbine exhibited the conventional

looping characteristics representing lling and emptying effects, which was also the case for the

nozzleless and single-entry nozzled turbine. The swallowing capacity of the twin-entry turbine, during

full admission testing, was recorded to be inconsistent between the two entries, in particular they were

at different pressure ratio levels e the shroud end entry was in most cases more pressurized compared

to the hub end entry, as much as 13%. Contrarily, during out-of-phase testing the swallowing capacity of

both the turbine entries was found to be similar. The cycle-averaged efciency of the nozzled turbine

either twin or single-entry was found to depart signicantly from the equivalent quasi-steady, in

comparison to the nozzleless single-entry turbine, this was as much as 32%.

2011 Elsevier Ltd. All rights reserved.

1. Introduction

A turbocharger turbine operates under unsteady conditions due

to the pulsating nature of the exhaust gases. In consequence, twin-

entry turbines are generally designed and used for better energy

extraction from the pulsating exhaust gases. Twin-entry turbine

allowsthe pulsationin theexhaust gasto be sustained,by separating

the banks of the manifold till it reaches the rotor/nozzle inlet. In

doing this, the higher isentropic energy in the pulse can made

available for the turbine to be extracted. However, pulsating ow

creates unique operating characteristics in a twin-entry turbine, in

comparison to a single-entry e this is further amplied with the

variable geometry congurations. The staggered pulsation in boththe entries, due to sequences in the exhaust valve opening, means

the turbine is operating in a combination mode of mostly unequal

and partial entry conditions. Thus, there is a need for experimental

workto understand the unsteady-state performance of a twin-entry

variable geometry turbine in various operating conditions, in

comparison to the single-entry and nozzlelessunit. This is necessary

for the understanding of turbine behaviour under various operating

conditions of full, unequal and partial admissions.

Wallace and Blair[1] and Benson and Scrimshaw[2] presented

the earliest systematic study on the unsteady performance of

a radial turbine. These were followed by Wallace et al. [3], Miya-

shita et al. [4], Benson[5] and Kosuge et al.[6]. All these investi-

gations were limited by the inability of the instruments to

instantaneously measure all the relevant performance parameters.

Thus, only the static pressure was measured instantaneously, while

the mass ow rate, temperature, speed and torque were measured

as time-mean values. This remains similar for the investigation

presented by Capobianco et al. [7,8]and Capobianco and Gambar-

otta[9]. Research work by Capobianco et al. [7,8]concentrated onestablishing correlations between the unsteady parameter and the

equivalent steady values. However, as only the pressure measured

instantaneously, all comparative parameters were assessed

through quasi-steady approach. Winterbone et al. [10]and Win-

terbone and Pearson[11]presented a good review of the experi-

mental techniques and understanding of the pulsating ow

performance of a turbocharger turbine.

Dale and Watson[12]and Dale[13]were the earliest to present

an unsteady performance data for a turbine with all the parameters

measured instantaneously (except instantaneous temperature). A

radial twin-entry nozzleless turbine was used to investigate the

* Corresponding author. Tel.:44 20 7594 7241; fax: 44 20 7823 8845.

E-mail addresses: [email protected] (S. Rajoo), [email protected]

(A. Romagnoli), [email protected](R.F. Martinez-Botas).

Contents lists available at SciVerse ScienceDirect

Energy

j o u r n a l h o m e p a g e : w w w . e l s e v i e r . c o m/ l o c a t e / e n e r g y

0360-5442/$e see front matter 2011 Elsevier Ltd. All rights reserved.

doi:10.1016/j.energy.2011.12.017

Energy 38 (2012) 176e189
mailto:[email protected]:[email protected]:[email protected]://www.sciencedirect.com/science/journal/03605442http://www.elsevier.com/locate/energyhttp://dx.doi.org/10.1016/j.energy.2011.12.017http://dx.doi.org/10.1016/j.energy.2011.12.017http://dx.doi.org/10.1016/j.energy.2011.12.017http://dx.doi.org/10.1016/j.energy.2011.12.017http://dx.doi.org/10.1016/j.energy.2011.12.017http://dx.doi.org/10.1016/j.energy.2011.12.017http://www.elsevier.com/locate/energyhttp://www.sciencedirect.com/science/journal/03605442mailto:[email protected]:[email protected]:[email protected]


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performance characteristics under steady, unsteady, equal admis-

sion andpartial admission conditions. Dale [13] presented results for

pulse frequency of 40 and 60 Hz, turbine speed 300e

700 RPS,a mean total inlet temperature of 400 K and a maximum instan-

taneous pressure ratio of 1.8. The turbine was coupled to an eddy-

current dynamometer with a loading capacity of 12 kW. Dale and

Watson [12] measured the instantaneous mass ow rate at the

turbine inlet with a hotwire anemometry and the instantaneous

inlet static pressure with a fast response pressure transducer. The

torque was deduced from the rotor rotational speed and accelera-

tion. Almost all researchers investigating the unsteady performance

of a radial/mixed ow turbine have used these measurement

techniques ever since. Based on the instantaneous measurement,

Dale and Watson [12] presented the typical hysteresis loop for

radial turbine unsteady performance curves. Nikpour [14] con-

ducted a study similar to Dale[13]but with a nozzleless turbine. A

hydraulic dynamometer was used to load the turbine with highermaximum speed compared to Dale[13]. A hysteresis loop was also

recorded for the unsteady performance curve, but with greater

deviation from the equivalent steady, compared to Dale [13].

Baines et al.[15]further developed the work by Dale and Wat-

son [12]with the study of a twin-entry nozzleless radial turbine

under pulsating ow conditions. The measuring instruments

remain largely similar to Dale and Watson [12]. The turbine

performances were evaluated for pulse frequency of 20, 40 and

60 Hz, turbine speed of 300e500 RPS, a mean total inlet temper-

ature of 400 K and a maximum instantaneous pressure ratio of 1.8.

Baines et al.[15]also documented the out-of-phase pulsating ow

results, where dual-loops in the instantaneous efciency trace were

found. These were attributed to the effect of reverse ow from the

open valve limb to the close valve limb in an out-of-phase testing.In overall, the similar hysteresis loops in the unsteady performance

curves were found as other researchers.The experimental facility in

the current study is an improved version from Dale and Watson

[12]and Baines et al.[15].

The more recent unsteady turbine performance evaluation was

given by Marelli and Capobianco [16], where they have utilized

largely similar measurement techniques as in the current study.

However, their investigation was focused on single-entry radial

turbine under waste gated conditions. In broad perspective, Marelli

and Capobianco [16] showed the signature unsteady curves and

quantied the turbine efciency depreciation in comparison to

quasi-steady assumptions.

This paper addresses the turbine unsteady performance under

thecombination of twin-entry andvariablegeometrycongurations,

which has not been shown previously. The presented experimental

data and performance characteristics will serve as an essential vali-

dation tools for computation models to address the accurate turbo-charger component assessments methods in an engine design

process.

2. Experimental facility

Fig. 1illustrates the turbine congurations used for the current

study, which was designed in progression from nozzleless single-

entry to nozzled twin-entry. The turbine was designed based on

conventional turbomachinery techniques as described in Japikse

and Baines[17].Table 1gives the basic features of the nozzless and

the nozzled turbine in the current study, while the further design

details can be seen in Romagnoli et al. [18]. The experimental

facility available in Imperial College London is a simulated recip-

rocating engine test-bed for turbocharger testing. The facility hasthe capability of conducting steady and unsteady ow testing of

single and twin-entry turbines. A schematic diagram of the turbine

test-rig is shown in Fig. 2, which includes the inlet and outlet

measurement locations for the turbine stage. The test-rig is

equipped with an eddy-current dynamometer, which enables

turbine testing within a large velocity ratio range[6,7]. The test-rig

is supplied by screw-type compressors, capable to delivering air up

to 1.2 kg/s mass ow rate at a maximum pressure of 5 bars (abso-

lute). The two separated streams of airow in each limb pass

through a rotary air pulse generator, which consists of two rotating

chopper plates as shown in Fig. 3. The chopper plates consist of

unique cut-out, which were originally designed by Dale and

Watson[12]to experimentally simulate the exhaust gas pulsation

ofan engineas shown in Fig. 3. A variable speed D.C. motorcontrolsthe rotating frequency of thechopper plates, hence the frequency of

the pulsation. The phase of the entry pulses in both the limbs can be

varied either to be in-phase or out-of-phase by changing the rela-

tive position of the two chopper plates. The simulated exhaust gas

pulsation of the test-rig has been the centre of numerous industrial

related turbocharger turbine investigations, which reiterates its

relevancy as an accurate representative of an engine exhaust gas.

Some of the most recent ndings are documented by Costall et al.

(2011)[19]and Copeland et al. (2012) [20e22]. The focus of this

study, as with all the other related work in the past, is on charac-

terizing the turbine behaviour under a given pulse ow. The

understanding derived from the characterization can be an essen-

tial input towards the formulation of a generic assessment tool in

engine design process, for any forms of pulsating exhaust gases.

Nomenclature

A Area, mm2

C Absolute velocity, m/s

D Rotor Diameter, m

MFP Mass ow parameter

N Turbine speed, rps

P Pressure, Pa

PR Pressure ratio

r radius, mm

T Temperature, K

Tr Torque, Nm

U Rotor tip velocity, m/s_m Mass ow rate, Kg/s_W Power, kW

h Efciency

q Phase Angle

j Azimuth Angle

Cis isentropic velocity, m/s

Subscript

actual actual conditions

in inner limb (entry)

ins instantaneous conditions

isent isentropic

out outer limb (entry)

SE single-entry

TE twin-entry

tot total (inner outer limb entry)

ts total-to-static

0 Total conditions

1 Turbine stage inlet

5 Turbine stage exit

S. Rajoo et al. / Energy 38 (2012) 176e189 177


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3. Unsteady testing and analysis procedures

The conditions for the unsteady test points are set up to corre-

spond to the steady-state velocity ratio at peak efciency fora given

speed. This is achieved by comparing the energy averaged velocity

ratio at unsteady ow condition, with the steady-state velocity

ratio at peak efciency. Once the desired condition is achieved and

the system stabilizes, measurements of pressures, temperatures,

mass ow rate, speed and torque are acquired instantaneously with

reference to chopper plate frequency. A pulse signal produced once

per revolution by the chopper plate is used as trigger to concur-

rently measure all the parameters.

The instantaneously acquired data points are processed in fewstages before it can be used to calculate the performance parame-

ters. Hotwire anemometer used to measure mass ow rate is set to

traverse and measure at 36 locations in the pipe cross section.

Meanwhile the data logging system is set to acquire 50 cycles of

data for each location. Thus, a total of 1800 cycles are recorded for

every test point during unsteady experiments. For all but the hot-

wire measurement, the 1800 cycles are ensemble averaged to

reduce any random non-cyclic uctuation in the measurement.

Meanwhile the 36 points hotwire measurements are integrated as

specied in BS 1042:1983 and only 50 cycles are ensemble aver-

aged. The hotwire measurement is then corrected for temperature

before the nal mass ow rate can be deduced. The instantaneous

temperature of the ow is calculated based on the isentropic

compression assumption. The ensemble averaged signals are then

ltered with Finite Impulse Response (FIR) to further attenuate

noises in the reading.

All but the speed signals are recorded with an analogue-digital

card with constant sampling rate of 20,000 samples per second.The speed signal is logged through a counter card, thus the

recorded number of samples per cycle are not constant and not in

match with the analogue readings. For the purpose of instanta-

neous point-by-point analysis, the speed signal is resample at

a constant 20,000 samples per second with spline interpolation.

The speed signal is then derived to determine the uctuating

torque of the rotor. Consequently the uctuating torque is used to

calculate the total power of the turbine. The difference in the

measurement locations of the actual and isentropic properties, as

shown in Fig. 2, requires a degree of phase shifting to enable

evaluation on common time frame. This is achieved by phase

shifting all the measurement to a common reference location

using bulk sonic ow velocity as described by Szymko et al. [23]

(seeFig. 2).

Fig. 1. Turbine volutes used for the current study.

Table 1

Nozzleless vs. nozzled turbine volute dimensions[18].

Nozzleless Nozzled

Volute tongue position 50 30

Centroid throat radius at 0 (mm) 73.9 100.0

Stator throat area (mm2) 2568.6 3300.0

A/r 34.7 33.0

S. Rajoo et al. / Energy 38 (2012) 176e189178


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4. Unsteady performance parameters

4.1. Massow parameter vs. pressure ratio

The denition of the mass ow parameter and pressure ratio

follows the conventional turbomachinery method, with instanta-

neous time varying individual components for each property. Mass

ow parameter is a pseudo-dimensionless parameter derived by

relating the inlet ow velocity and the inlet Mach number, as given

in Eq. (1). For the twin-entry turbine the mass ow parameter is

calculated considering the contribution of each limb on the overall

ow capacity. The mass ow leaving the two entries mixes prior

entering to the rotor and it is therefore sensible to consider an

average of the ow properties. From the energy equation, the

stagnation temperature is calculated as a mass weighed average

value while for the total pressure at inlet, an area average value is

Fig. 2. Turbocharger test-rig schematic diagram, shown with relevant measurement locations.

Fig. 3. Rotating chopper plate for the air pulse generator to experimentally simulate exhaust gas pulsation.



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more appropriate. The nal equation for the pseudo-dimensionless

mass ow parameterin a twin-entry turbine is therefore given in

Eq.(2), where _mins;tot _mins;in _mins;out.

MFPins;SE _mins

ffiffiffiffiffiffiffiffiffiffiffiT01ins

pP01ins

(1)

MFPins;TE _mins;tot,

ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi_mins;in_mins;tot

T01ins;in_mins;out_mins;tot

T01ins;out

s0:5

P01ins;in P01ins;out

(2)The instantaneous pressure ratio is given in Eq.(3). In a twin-

entry turbine, the pressure ratio can be dened using an area

averaged pressure ratio between the two entries as given in Eq.(4).

PRins;SE P01ins

P5ins(3)

PRins;TE 1

2

P01ins;in

P5ins

P01ins;outP5ins

(4)

4.2. Efciency vs. velocity ratio

The turbine instantaneous total-to-static efciency for a single-

entry turbine is given in Eq. (5). In the twin-entry cases instead,

since power is an extensive quantity, the instantaneous total-to-

static efciency can be calculated considering the sum of the

instantaneous isentropic power of the entries, as given in Eq. (6).

htsins;SE _Wactualins_Wisentins

(5)

htsins;TE

_Wactualins _Wisent;in _Wisent;out

ins

(6)

The instantaneous velocity ratio is the ratio between the rotor

tip speed and the inlet isentropic velocity. The formulation of the

instantaneous velocity ratios for the single and twin-entry turbine

is given in Eq.(7) and Eq.(8) respectively.U

Cis

ins;SE

pDNinsffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi

2 _Wisentins_mins

s (7)

U

Cis

ins;TE

pDNins

2 _Wisent;in

_Wisent;outins_mins;tot

(8)

5. Nozzleless and single-entry nozzled turbine:for

comparisons

The twin-entry turbine in the current study is the product of

progressive design stages from the commercial nozzleless followed

by single-entry nozzled turbine. Thus it would be benecial to look

into the unsteady characteristics of the preceding turbine congu-

rations, which will serve as a comparison basis to the twin-entry

variable geometry turbine performance in the following section.

Fig. 4 shows the plots of the turbine swallowing capacity with

nozzled single-entry and nozzlelessvolutes,for 40Hz and 60Hzow

and 80% equivalent speed. The nozzled turbine loops shown are for

50 vane angle setting, which has equivalent swallowing capacity as

with the nozzleless turbine. The equivalent quasi-steady curve of

the nozzleless turbine is also shown on each plot. It can be noticed

that the loop of the nozzleless turbine shows better encapsulation

of the equivalent quasi-steady curve. However, in both the nozzle-

less and nozzled cases, the range of mass ow parameter and

pressure ratio reduces at higher frequency. This shows that the

shorter time interval in between the pulses at higher frequency

enables a turbine to operate at a given point with narrower mass

ow rate and pressure ratio cycles. Furthermore, the hysteresis

loop of the nozzleless turbine shows a clear sign of increasing

wave action effect at higher ow frequency, since, if there was

a predominance of lling and emptying effects, one would have

seen a more regular shape in the parts of the hysteresis shape.Waveactioneffectcan bedenedas a continuous interaction between the

travelling pulses and the reected pulses in a turbine ow system.

This could be clearly observed as increasing perturbation in the

recorded instantaneous pressures as well as instantaneous mass

ow parameters. Concurrently, the travelling ow is also leavingthe

turbine as it enters, but at a different rate, which is dened aslling

andemptying effect.It hasbeen shown in many previous worksthat

there is a strong link between the pulse ow frequency and the

dominance of either wave action or lling emptying effect.

Compared to the nozzleless turbine, the hysteresis loop of the

nozzled turbine shows stronger sign oflling and emptying effect

even at higher ow frequency. This is possibly due to the bigger

volume and the existence of the nozzle ring, which effectively

creates an intermediate volume with its own

lling/emptying

0.00

1.00

2.00

3.00

4.00

5.00

6.00

7.00

8.00

1 1.4 1.8 2.2 2.6

s/gk(.raPwolFssaM

T10

P/

10)

Pressure Ratio (P01/P5)

x 1e-5

40Hz, 80% Speed

Single Entry Nozzled

Nozzeless

Steady Nozzeless

0.00

1.00

2.00

3.00

4.00

5.00

6.00

7.00

8.00

1 1.4 1.8 2.2 2.6

/

gk(.raPwolFssaM

T10

P/

10)

Pressure Ratio (P01/P5)

Nozzleless

Steady Nozzleless

x 1e-5

60Hz, 80% Speed

Single Entry Nozzled

s

Fig. 4. Comparison of the mass ow parameter vs. pressure ratio curves between the

single-entry nozzled and nozzleless turbine at 40 & 60 Hz, 80% equivalent speed

condition (nozzleless results from Szymko [24]).



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characteristics. Thus the effect of different measurement locations

for the isentropic and actual conditions of the turbine, as shown in

the Fig. 2, in a pulsating ow test is more pronounced in the nozzled

turbine.

Fig. 5shows the plots of the turbine instantaneous efciency in

a pulse cycle for the nozzleless (plots (a)) and nozzled (plots (b) and

(c)) volutes. The curves are for different vane angle settings and 80%

equivalent speed. The efciency changes over a pulse cycle show

a trend of improvement with the increase in ow frequency. This is

observed in both the nozzled and nozzleless turbines, but

marginally more signicant in the latter. It can be noticed that the

nozzleless turbine exhibit more negative efciencies at the begin-

ning and end of a cycle, which is the low pressure ratio region. It is

more pronounced at lower ow frequency, which is also seen in the

nozzled setting but with lesser effect. In the nozzled settings, the

turbine did not exhibit negative efciency at close nozzle positions

(70 and 65). This is due to the nozzle constantly providing suf-

cient ow momentum to the rotor during the pulse cycle. But at

a more open positions especially at 40 and 50 vane angle settings,

the turbine gradually exhibit negative efciency. This reects the

decrease in the tangential ow velocity at open nozzle positions in

the low pressure ratio region. It is also noticed that during the rst

120 crank angles, where most of the isentropic power concen-

trated in the pulse, both the nozzleless and nozzled turbines shows

no signicant difference in most cases. The nozzled turbine shows

increasing departure from the nozzleless unit at closer nozzle

settings, which can be explained as the effect of choking and mass

accumulation. Table 2 shows the velocity ratio, cycle-averaged

efciency and the equivalent quasi-steady efciency for the

nozzleless and nozzled turbines as shown in Fig. 5. It can be seen

that the cycle-averaged efciency is generally lower than the quasi-

steady assumption in most cases. However, in specic, it is clear

that the degree of departure of the unsteady conditions from the

equivalent quasi-steady is more pronounced in the nozzled turbine,

especially at higher vane angle settings.

6. Results and discussions: twin-entry VGT

6.1. Swallowing capacity of twin-entry turbine

The nozzled single-entry turbine is converted into twin-entry

through partitioning as described in Romagnoli et al. [18]. The

twin-entry turbine is tested under pulsating ow with in-phase

and 180 out-of-phase ow admission conditions. The typical

inlet static pressure proles during in-phase and 180 out-of-phase

conditions for various ow frequencies and speed are given in

Fig. 6. The inner and outer limb notations in the gure refer to the

individual inlet of the twin-entry turbine, shown in Fig. 2. The inlet

conditions are monitored throughout the experimental process to

ensure a consistent phase between the limbs,as this affects the post

processing for performance analysis.

Figs. 7 and 8show the swallowing capacity of the twin-entry

turbine during in-phase full admission test at 50% and 80% equiv-

alent speed conditions respectively. In both the gures, results for

40 Hz and 60 Hz pulsating owconditions are shown, as well as the

equivalent quasi-steady curves for the partial admissions. The

swallowing capacity of the individual limbs is shown and the

overall swallowing capacity of the turbine calculated with mass

averaged inlet total temperatures and pressures. The common

looping curves are observed in all the cases, with some degree of

encapsulation around the quasi-steady curves; due to continuous

lling and emptying of the volume during pulsating ow condi-

tions. It can be seen that the outer limb swallowing capacity tends

towards steady characteristics during the 50% speed, butfalls below

the equivalent quasi-steady. However during 80% speed condition,

-2

-1.6

-1.2

-0.8

-0.4

0

0.4

0.8

1.21.6

2

0 90 180 270 360

Phase Angle (Deg)

40 Hz

60 Hz

Nozzleless80%Speed

-2

-1.6

-1.2

-0.8

-0.4

0

0.4

0.8

1.2

1.6

2

0 90 180 270 360

Phase Angle (Deg)

40deg

50deg

65deg

70deg

60Hz, 80%Speed

-2

-1.6

-1.2

-0.8

-0.4

0

0.4

0.8

1.2

1.6

2

0 90 180 270 360

Phase Angle (Deg)

40deg

50deg

65deg70deg

40Hz, 80%Speed

60deg

60Hz60deg

a

b

c

Efficiency(t-s

)

Efficiency(t-s

)

Efficiency(t-s

)

Fig. 5. Comparison of the efciency curves between the nozzled single-entry and

nozzleless turbine for different nozzle vane angles, different ow frequencies and 80%

equivalent speed condition (nozzleless results from Szymko [24]).

Table 2

Energy weighed cycle-averaged values of velocity ratio and ef

ciency for single-entry nozzled and nozzleless turbine, 80% equivalent speed conditions ( nozzleless

results from Szymko,[24]).

Frequency & conguration U/Cis hcycle-avg hquasi-steady

40 Hz 40deg 0.65 0.604 (2.8%) 0.576

40 Hz 50deg 0.662 0.607 (3.1%) 0.638

40 Hz 60deg 0.652 0.613 (12.6%) 0.739

40 Hz 65deg 0.645 0.603 (15.4%) 0.757

40 Hz 70deg 0.619 0.588 (14.6%) 0.734

40 Hznozzleless 0.624 0.626 (4.6%) 0.672

60 Hz 40deg 0.676 0.657 (7.3%) 0.584

60 Hz 50deg 0.684 0.69 (4.2%) 0.648

60 Hz 60deg 0.682 0.686 (7.1%) 0.757

60 Hz 65deg 0.662 0.64 (13.2%) 0.772

60 Hz 70deg 0.62 0.556 (19.4%) 0.75

60 Hz nozzleless 0.653 0.691 (0.8%) 0.683



7/14

100

110

120

130

140

150

160

170

180

190

0 45 90 135 180 225 270 315 360

)aPk(erusserPcitatStelnI

Phase Angle (deg)

Outer Limb

Inner Limb

40Hz 50% Speed 180 Out- of -Phase Admission

100

110

120

130

140

150

160

170

180

190

0 45 90 135 180 225 270 315 360

)aPk(erusserPcitatSteln

I

Phase Angle (deg)

Outer Limb

Inner Limb

60Hz 50% Speed 180 Out-of-Phase Admission

100

120

140

160

180

200

220

240

0 45 90 135 180 225 270 315 360

)aPk(erusserPcitatS

telnI

Phase Angle (deg)

Outer Limb

Inner Limb

40Hz 80% Speed In-Phase Full Admission

100

120

140

160

180

200

220

240

0 45 90 135 180 225 270 315 360

)aPk(erusserPcitat

StelnI

Phase Angle (deg)

Outer Limb

Inner Limb

60Hz 80% Speed In-Phase Full Admission

Fig. 6. Twin-entry turbine typical inlet static pressure proles for in-phase and 180 out-of-phase admission: given for individual entries, inner and outer limbs.

Fig. 7. Swallowing capacity of twin-entry turbine at 50% equivalent speed and in-

phase full admission: given for individual entries, inner and outer limbs as well as

mass averaged combination.

Fig. 8. Swallowing capacity of twin-entry turbine at 80% equivalent speed and in-

phase full admission: given for individual entries, inner and outer limbs as well as




8/14

it shows clear looping characteristics around the steady curve.

Comparing the swallowing capacity between 50% and 80% speeds,

it can be seen that the turbine swallows more compared to the

equivalent steady at higher speed conditions. The cycle-averaged

unsteady mass ow parameter at 40 Hz 80% speed is around 38%

higher than the cycle-averaged steady condition. As for 60 Hz 80%

condition the number is approximately 60%.

Another observation fromFigs. 7 and 8is that the pressure ratio

range during a cycle in pulsating ow emulates the steady curves

more in the 80% compared to the 50% speed. Interesting as it is,

further scrutiny in comparing the two speed cases, one can notice

that for the 80% speed, the swallowing capacity exhibits plateauing

characteristics during the lling period, which is pronounced in the

40 Hz ow. This suggests the turbine is in a choking state, broadly

speaking beyond the pressure ratio of 1.8, which in some ways

explains the situation of the turbine experiencing high pressure

ratio and the strong lling and emptying behaviour in the 80%

speed cases. Furthermore, comparing the individual limbs under

unsteady ow partial admission, the swallowing capacity of the

outer limb is consistently at the higher pressure ratio, this is more

apparent in the 50% speed conditions. However this is not observed

in the steady ow partial admission operation. So the question here

is why the swallowing capacity of the individual limb under partialadmission is different between each other only in the unsteady

ow operation.

The swallowing capacity of a turbine is inuenced by factors

such as, the volute and rotor inlet area, centrifugal forces on the

ow, boundary layer growth and possible ow separations. The

experiment results suggest that one or a combination of these

factors is different between the limbs in the unsteady ow opera-

tion. Even though the current experimental results do not allow

a complete understanding of such a complex behaviour, one can

indicate that the ow has insufcient time to stabilize in the

unsteadyow, thus causing a departure from the steady as well as

a difference between limbs. In this respect, it can be said that ow

near the rotor shroud (fed mainly by the outer limb) has a higher

blockage, which is the result of an increase ow curvature and maylead to a ow separation. Higher blockage leads to higher pressure

ratio operation as per the experimental observations. Overall, the

unsteady swallowing capacity curves at different conditions in

Figs. 7 and 8 show that the twin-entry turbine exhibits more of

a lling and emptying characteristics, consistent with the observed

characteristics of the single-entry nozzled turbine (Fig. 4).

Fig. 9shows the swallowing capacity of the twin-entry turbine

under unequal 180 out-of-phase ow condition, for 40 Hz and

60 Hz pulsations. The swallowing capacity of the individual entry

shows clear lling and emptying effect for all conditions, with

almost full encapsulation of the steady curves. The calculation of

the overall swallowing capacity of the twin-entry turbine is not

straightforward in an unequal ow conditions. Mass averaging the

pressure ratios of the two entries will reduce the range coveredduring a cycle, compared to the individual entry. This can be seen

clearly inFig. 9, as well as the double looping in the overall swal-

lowing capacity representing both entries.

6.2. Power and efciency of twin-entry turbine

Fig. 10 shows the isentropic and actual power curves of the

twin-entry turbine under different conditions. The actual power

is calculated from the measured rotor torque and speed, while the

isentropic power calculated with measured upstream parameters.

In general, the turbine power is found to be similar for both the

entries. A good correlation is observed between the isentropic

and actual power, given the separate deriving technique for both.

It can be seen that at some instances the actual power of the

turbine goes negative; this is explained in Szymko et al. [24] as

the conditions where the turbine rotor imparts momentum on

the ow due to a very low energy content of the free stream.

Interestingly such condition does not occur for out-of-phase ow,

indicating better ow momentum with 2 pulses per cycle. In

addition, Fig. 10 also shows that the isentropic power at 60 Hz

and 80% speed under full admissiondoes not show a clear peak as

expected. Instead, the isentropic power remains fairly at over

the whole pulsation period, which results in poor turbine

instantaneous efciency representation for majority of the pulse

period. However, the instantaneous pressure prole, as shown in

Fig. 6, does not show a similar plateauing trend. This seems to

suggest mass accumulation effect due to the existence of nozzle

between the volute and rotor. This effect is clearly dominating atthe higher frequency, as it was observed even at 80 Hz (not dis-

cussed here). At higher frequency, the emptying rate of the noz-

zled turbine system is not capable to cope with the lling rate,

thus resulting in mass accumulations. However this is not

observed at out-of-phase conditions, as the amount of mass

entering the turbine system is staggered, resulting in better

emptying process.

The comparison of the actual and isentropic power of the

turbine will enable the derivation of efciency however instanta-

neous comparison creates a level of complexity. This is due to

factors affecting the comparison mainly the phase shifting due to

different measurement locations. There are still debates on the

effective method of dening the efciency of a turbine under

unsteady

ow conditions, Rajoo and Martinez-Botas[25].

Fig. 9. Swallowing capacity of twin-entry turbine at 50% equivalent speed and 180

out-of-phase admission: given for individual entries, inner and outer limbs as well as




9/14

6.3. Cycle averaging and comparison to quasi-steady

Given the uncertainty associated with the meaning of the

instantaneous turbine efciency, an alternative method to evaluate

energy conversion is by cycle averaging the parameters. The main

advantage of the cycle averaging method is that it is not affected by

phase shifting. The cycle-averaged efciency represents the ratio ofenergy extracted by the turbine per pulse cycle divided by isen-

tropic energy owing into the system, as given in Eq.(9).

htsenergyavg

Zq0

hts

_Wtisent

dt

Zq0

_Wtisentdt

Zq0

_Wtact

dt

Zq0

_Wtisentdt

htscycleavg (9)

The velocity ratio is calculated with energy weighed averaging

as shown in Eq. (10),1

U

Cis

energyavg

Zq0

U

Cist, _Wtisent

dt

Zq0

_Wtisentdt

(10)

The calculated energy weighed cycle-averaged velocity ratio

is then used to read the corresponding efciency from the steady

map for that given speed. The cycle-averaged efciency is then

compared with the equivalent quasi-steady efciency. By doing this

an appropriate comparison between parameters obtained at

different conditions can therefore be performed.

Table 3reports a comparison between the cycle-averaged and

the quasi-steady efciency obtained for 60 vane angle at 50% and

80% equivalent speeds under in-phase and out-of-phase ow, for40 Hz, 60 Hz and 80 Hz frequencies. It is worth noting that the full

admission assumption has been considered here for the quasi-

steady value. This is appropriate when dealing with in-phase ow

(since both entries ow concurrently) while further analysis would

be required for the out-of-phase ow condition (since each limb is

pressurized at staggered intervals).

Figs. 11 and 12show the efciency comparison for the in-phase

and out-of-phase ow respectively (refer Table 3). Observing the

comparison, one can notice that the cycle-averaged unsteady ef-

ciency drops substantially, especially for the in-phase ow and

lower speed conditions. The cycle-averaged efciency remains

below the quasi-steady efciency with a difference of 25.4%, 25.4%

and 0.8% for 40 Hz, 60 Hz and 80 Hz respectively. At higher speed

instead, the discrepancy is less signicant and in some cases thecycle-averaged efciency is higher than the quasi-steady efciency.

This occurs at 80 Hz and 40 Hz where the cycle-averaged efciency

is 5.4% and 6.5% higher than the quasi-steady. At high frequency the

similarity between the cycle-averaged and the quasi-steady ef-

ciency can be explained considering that the pulse amplitude is

smallerwhich makes the rotorexposedto more continuousow. At

lower frequency instead, the ow follows a lling and emptying

behaviour, thus signicant swallowing capacity changes are expe-

rienced leading to the differences between the cycle-averaged and

quasi-steady efciencies. Similar results were also shown by Kar-

amanis and Martinez-Botas [26], Szymko et al. [23,24]and Rajoo

and Martinez-Botas [25], who measured higher cycle-averaged

efciency in a single-entry turbine for higher speeds and

frequencies. However, it must be noted that no consistent trend

0

15

30

45

60

75

90

0 45 90 135 180 225 270 315 360

)W

k(rewoP

Phase Angle (deg)

Wisnt

Wact

40Hz 80% SpeedIn-Phase Full Admission

0

15

30

45

60

75

90

0 45 90 135 180 225 270 315 360

)W

k(rewoP

Phase Angle (deg)

Wisnt

Wact

60Hz 80% SpeedIn-Phase Full Admission

0

5

10

15

20

25

0 45 90 135 180 225 270 315 360

)W

k(rewoP

Phase Angle (deg)

Wisnt

Wact

40Hz 50% Speed

180 Out-of-Phase Admission

0

5

10

15

20

25

0 45 90 135 180 225 270 315 360

)W

k(rewoP

Phase Angle (deg)

Wisnt

Wact

60Hz 50% Speed

180 Out-of-Phase Admission

Fig. 10. Twin-entry turbine isentropic and actual power proles for in-phase and 180 out-of-phase admissions: given for individual entries, inner and outer limbs.

1 In a twin-entry conguration _Wisent _Wisent;in

_Wisent;out , while the velocity

ratioU/Cis is equal to that provided in Eq. (8).



10/14

0.3

0.35

0.4

0.45

0.5

0.550.6

0.65

0.7

0.75

0.8

40 50 60 70 80

ycneiciffE

Frequency (Hz)

0.3

0.35

0.4

0.45

0.5

0.55

0.6

0.65

0.7

0.75

0.8

0.4 0.45 0.5 0.55 0.6 0.65

y

cneiciffE

U/Cis

Out-of-Phase Flow

Cycle avrg eff - 27.9 revs/K^0.5 Quasi-steady - 27.9 revs/K^0.5


50% Equiv Speed 50% Equiv Speed


Fig. 12. Comparison cycle-averaged vs. quasi-steady ef

ciency for the out-of-phase

ow at 50% and 80 equivalent speed for 60

vane angle.

Table 3

Comparison of energy weighted cycle-averaged and quasi-steady efciency.

60 vane angle

Dh hcycleavg=hquasisteady 1 100% Dh hcycleavg=hquasisteady 1 100%

50% equivalent speed 80% equivalent speed

U/Cis hcycle-avg hquasi-steady h U/Cis hcycle-avg hquasi-steady h

In-phase

40 Hz 0.542 0.472 0.631 25.4% 0.640 0.835 0.784 6.5%

60 Hz 0.503 0.456 0.611 25.4% 0.596 0.616 0.768 19.8%

80 Hz 0.533 0.631 0.626 0.8% 0.629 0.823 0.781 5.4%

Out-of-phase

40Hz 0.587 0.434 0.647 32.9% 0.595 0.631 0.768 17.8%

60Hz 0.464 0.415 0.586 29.2% 0.587 0.520 0.763 31.8%

80Hz 0.567 0.526 0.641 17.9% 0.577 0.560 0.757 26.0%

0.3

0.4

0.5

0.6

0.7

0.8

0.9

40 50 60 70 80

ycneiciffE

Frequency (Hz)

0.3

0.4

0.5

0.6

0.7

0.8

0.9

0.4 0.45 0.5 0.55 0.6 0.65 0.7

ycneiciffE

U/Cis

In-Phase Flow



50% Equiv Speed 50% Equiv Sp eed


Fig. 11. Comparison cycle-averaged vs. quasi-steady efciency for the in-phase ow at 50% and 80% equivalent speed for 60 vane angle.



11/14

was observed in the cycle-averaged efciency variation. In the out-

of-phase ow conditions, the cycle-averaged efciency shows large

discrepancies from the quasi-steady e at 50% equivalent speed

a difference of 32.9%, 29.2% and 17.9% was measured at 40 Hz, 60 Hz

and 80 Hz respectively. A similar drop was also found at 80%

equivalent speed e the cycle-averaged efciency deviates from the

quasi-steady by 17.8%, 31.8% and 26.0% for the 40 Hz, 60 Hz and80 Hz cases respectively.

Another interesting aspect to taken into account is the effect of

the pulsed ow phases (in-phase or out-of-phase) on the efciency.

From Table 3 it can be seen that the out-of-phase condition is

detrimental to the overall efciency above all at high speed. At 80%

equivalent speed, the cycle-averaged efciency drops from 0.835,

0.616, and 0.823 down to 0.631, 0.520, and 0.560 for 40 Hz, 60 Hz

and 80 Hz respectively, from the in-phase to out-of-phase ow

conditione this corresponds to a drop of 32.2%, 18.4%, and 46.9%

respectively. However, the same decit is not observed at lower

speed where the efciency drop is much lesser. At 50% equivalent

speed, the efciency drops from 0.472, 0.456 and 0.631 to 0.434,

0.415 and 0.526 (corresponding to a drop of 8.75%, 9.8% and 19.96%)

for 40 Hz, 60 Hz and 80 Hz respectively, from the in-phase to out-of-phaseow condition.

In order to further understand the correlation between the

cycle-averaged and the quasi-steady assumption, the ratios

between these efciencies has been calculated for both in-phase

and out-phase ow at 50% and 80% equivalent speeds. The

parameter called efciency ratio is plotted in Fig. 13against pulse

frequency (40 Hz, 60 Hz, 80 Hz). As noted from Fig. 13, at 80%

equivalent speed, the efciency ratio for the in-phase ow does not

vary substantially with frequency, which goes in favour of the

quasi-steady assumption. The dip observed at 60 Hz (efciency

ratioz0.8) can be considered as the result of the transition through

a region where the quasi-steady assumption no longer applies. In

the out-of-phase ow conditions instead the efciency ratio

remains below unity for both 50% and 80% equivalent speeds.

In order to understand the efciency trends, the unsteady time-

averagedpower and massowwerecalculated andcompared tothe

quasi-steady values e these are shown inFig.14for 80% speed. The

quasi-steady values are obtained from the steady curves at the

equivalent unsteady isentropic energy averaged velocity ratio.

The gure show shat the quasi-steady average values generally over

predict the corresponding unsteady values since the power and

mass ow ratio remain below unity. The mass ow and power vary

in a consistent manner for both in-phase and out-of-phase ow

conditions e a decreasing value in the ratio observed with

increasing frequency. Despite the similarities existing between the

in-phase and out-of-phase power and mass ow ratios, Fig. 14

shows that in out-of-phase ow condition, the mass ow ratio

(and hence the isentropic power ratio) is approximately equal to

unity for almost any frequency, thus showing that the quasi-steady

assumption is adequate for a full unsteady calculation. The samedoes not occur for the in-phase conditions for which a dip at 60 Hz

can be observed in the mass ow and isentropic power ratio. As

already observedbefore (Fig.10)suchadipcouldbeattributedtothe

increasingly unsteady correlations between the wave action and

lling and emptying effects in the nozzled turbine system. On the

efciency side, the trend observed for the ratios between the cycle-

averaged and the quasi-steady efciencies (Figs. 11 and 12) are re-

ected in the actual and isentropic power ratios e the isentropic

power ratios remains consistently above the actual power ratio.

Comparing the cycle-averaged efciency with the quasi-steady

assumption based on full admission conditions is not conclusive

when dealing with twin-entry turbines. It is a known fact that the

twin-entry turbine is meant to work in out-of-phase ow condi-

tions in most cases. The incoming pulses from each bank ofmanifolds occur at staggered intervals, thus the turbine works in

partial admissionconditions in most cases compared to full admis-

sion. Therefore, in order to evaluate the quasi-steady assumption in

the out-of-phase ow conditions, it is appropriate to refer to the

partial admissionmaps instead of the full admission.Table 4shows

the quasi-steady efciency calculated using the full and partial

admissionconditions. The cycle-averaged efciencies and the ratio

0.540 50 60 70 80

0.6

0.7

0.8

0.9

1

1.1oitarydaets-isauQ/yda

etsnU

Frequency (Hz)

Ratio between Cycle-averaged and Quasi-steady efficiency

In Phase 50% Equiv Speed

In Phase 80% Equiv spped

Out of phase 50% Equiv Speed

Out of phase 80% Equiv Speed

Fig.13. Ratio between cycle-averaged and quasi-steady efciency for the in-phase and

out-of-phaseow at 50% and 80 equivalent speed.

0

1

2

40 60 80

oitary

daets-isauQ

/ydaetsnU

Frequency (Hz)

In Phase Flow

Wisnt ratio Wact ratio Mass flow ratioWis ratio

0

1

2

40 60 80

oitary

daets-isauQ

/ydaetsnU

Frequency (Hz)

Out of Phase Flow

Fig. 14. Ratio cycle-averaged and quasi-steady power and mass

ow rate for the in-phase and out-of-phase

ow at 80% equivalent speed and 60

vane angle.



12/14

between the cycle-averaged and the quasi-steady efciencies are

shown inFig. 15. FromFig. 15it can be seen that the ratio between

the cycle-averaged and the quasi-steady assumption based on the

partial admissioncondition is much closer to unity compared to the

cases where the full admission map is used (see Fig. 13). At 80%

equivalent speed, the quasi-steady efciencygoes from 0.768, 0.763

and 0.757 to 0.530, 0.533 and 0.534 for the 40 Hz, 60 Hz and 80 Hz

cases respectively, considering the full to partial admission condi-

tion e such a drop leads to values of the efciency ratio (inpartial

admission) to be 1.03, 0.85 and 0.92 for 40 Hz, 60 Hz and 80 Hz

respectively. At 50% equivalent speed, the ratio between the cycle-

averaged and quasi-steady efciency is approximately 10% higherthan that calculated considering the full admission curve, which

results in an improvement in the evaluation of the quasi-steady

assumption. In summary, the quasi-steady efciency shifts from

0.647, 0.586 and 0.641 to 0.607, 0.607 and 6.08 for the 40 Hz, 60 Hz

and 80 Hz cases respectively, considering the full admission to the

partial admissioncase.

Table 5 shows the velocity ratio, cycle-averaged efciencyat 50%

equivalent speed for different vane angle settings (40, 60, 70)

and ow frequencies (40 Hz, 60 Hz and 80 Hz) e these are plotted

in Fig. 16. The values are energy weighted average as shown in

Eq.(5) and Eq. (6). It can be noticed that the trend of the cycle-

averaged efciency for different vane angles does not seem to

follow a well-dened pattern. For the in-phase ow, the cycle-

averaged efciency shows about 12 percentage points drop from40 to 60 vane angle for 40 Hz and 60 Hz. This can be directly

linked to the high uctuation of the torque exhibited by the turbine

for open vane angles compared to the closed vane settings which

leads to higher cycle-averaged efciency[25]. For 70 vane angle

and in-phase ow 40 Hz and 60 Hz, the cycle-averaged efciency is

similar to that calculated for the 60 vane angle even though a large

departure from one another could be observed at 80 Hz. A different

scenario can be observed for the out-of-phase ow. The cycle-

averaged efciency increases with increasing frequency. At 60

vane angle, the efciency is higher than those measured at 40 and

70, while the 40 vane angle setting seems to perform better than

the 70. The efciency decit for the 70 vane angle compared to

the optimum vane angle is approximately 17.3%, 10.1% and 17.7% for

40 Hz, 60 Hz and 80 Hz respectively. Such a penalty in efciency

could be attributed to the blockage effects due to closed nozzle

settings, similar to what was found in the single-entry investiga-

tions[25]. The mass accumulation in the volume (volute pipe)

reduces the possible momentum imparted to the rotor with

consequent lower power output.

6.4. Further considerations

The analysis carried out within the current paper comes as

a continuation of a similar study done on the same turbine

congurations under steady-state conditions [18]. The results ob-

tained in steady-state conditions showed that the mass ow

parameters for the twin-entry are almost independent of the

conguration and of the type ofow admission (partialor unequal

admission[18]). A similar behaviour is not observed as the turbines

starts to operate under unsteady conditions. Even though the

hysteresis loops between the entries seem to have a fair degree of

symmetry, there are still some issues, which need to be addressed

in order to be able to make a nal judgement over the turbine

performance. These can be identied mainly in (1) the reliability

of the quasi-steady assumption, for which there are still on-going

debates but it is common to both single and twin-entry congu-

rations, and (2) the physical approach in the evaluation of the

performance parameters for a twin-entry turbine. Within the

current paper, in accordance with previously published works

[18,27], a mass weighed averaged total temperatures and an area

averaged pressure ratios were considered for the calculation of

mass ow parameters. However, the complexity attributed to the

mixing of the travelling pulses and the extra degree of freedoms

due to the admission conditions (either in-phase or out-of-phase)

makes the assessment particularly complex. In this sense the

experimental results presented here provide an initial and rather

unique insight on the response of a variable geometry twin-entry

Table 5

Energy weighted cycle-averaged efciency for different vane angles.

50% equivalent speed

40 vane angle 60 vane angle 70 vane angle

U/Cis hcycle-avg U/Cis hcycle-avg U/Cis hcycle-avg

In-phase

40 Hz 0.593 0.593 0.542 0.472 0.51 0.470

60 Hz 0.548 0.590 0.503 0.456 0.541 0.463

80 Hz 0.519 0.347 0.533 0.631 0.487 0.288

Out-of-phase

40 Hz 0.556 0.425 0.587 0.434 0.472 0.261

60 Hz 0.524 0.344 0.464 0.412 0.478 0.311

80 Hz 0.587 0.433 0.567 0.526 0.485 0.349

Table 4

Comparison between cycle-averaged and quasi-steady efciency considering full andpartial admissionassumption for the quasi-steady value.

Out-of-phaseow (quasi-steady assumption based on partial admission condition)

50% equivalent speed 80% equivalent speed

U/Cis hcycle-avg hqs, full hratio hqs,partial hratio U/Cis hcycle-avg hqs,full hratio hqs, partial hratio

40 Hz 0.587 0.434 0.647 0.67 0.530 0.81 0.595 0.631 0.768 0.82 0.607 1.03

60 Hz 0.464 0.415 0.586 0.70 0.533 0.77 0.587 0.520 0.763 0.68 0.607 0.85

80 Hz 0.567 0.526 0.641 0.82 0.534 0.98 0.577 0.560 0.757 0.73 0.608 0.92

0.5

0.6

0.7

0.8

0.9

1

1.1

40 50 60 70 80

oitarydaets-isauQ

/ydaetsnU

Frequency (Hz)

revs/K^0.5

revs/K^0.5

revs/K^0.5

revs/K^0.5

Full Admission

50% Equiv Speed

Full Admission80% Equiv Speed

Partial Admission50% Equiv Speed

Partial Admission80% Equiv Speed

Fig. 15.Ratio of cycle-averaged and quasi-steady efciency for the out-of-phase owat

50% and 80% equivalent speed, obtained using the full and partial admission quasi-

steady value.



13/14

turbine. The experimental results build up well with previously

published works and they form a benchmark for future works.

Currently on-going work within the Imperial College London

turbocharger research group are focussing on 1D and 3D simula-

tion in the aim to discover more physical insights of the single and

multiple entries turbine behaviour under pulsating ow condi-

tions. Some of the initial ndings of the simulation work are

documented by Costall et al. [19], Copeland et al.[21]and Chiong

et al. [28]. The extensive unsteady experimental results of the

work in this paper enable the accurate validation of computational

tools in turbocharger turbine performance predictions, as shown

by Chiong et al.[28]. This is benecial for turbocharger and engine

designers, since most available turbine model such as shown by

Xiande Fang et al. [29] and Xiande Fang & Yu Xu [30], are stillbased on steady-state ow.

7. Summaries

The current paper discusses performance analysis of a twin-

entry variable geometry turbine tested under pulsating ow

conditions. The presented experimental results are obtained from

a high precision pulsating ow turbocharger test-rig at Imperial

College London. The twin-entry turbine was designed in progres-

sion from single-entry nozzleless and nozzled turbines, while

maintaining the main geometrical features in all cases. These were

used as basis of comparison to gauge the operating variation and

the associated twin-entry turbine performance, under pulsating

ow conditions. Some initial results of the nozzleless and nozzledsingle-entry turbine are also shown as to allow readers to follow

through the design progression. Following are the main ndings

from the experimental investigation of this paper.

For the optimum vane angle setting (60) and in-phase ow

condition, the overall ow capacity is larger than the equivalent

quasi-steady. The encapsulation of the quasi-steady curve is only

partially achieved and it occurs mainly in the lowpressure region of

the maps for low speeds and frequencies. A more appreciable level

of encapsulation with the quasi-steady curve is achieved for the

ow capacity owning to each limb; a similar shape for the hyster-

esis loop could be found for the two entries even though the outer

entry (shroud-side) was found to operate at higher pressure ratios

than the inner entry (hub-side). As for the out-of-phase conditions,

the pulsating nature of the

ow at staggered intervals, leads to

a higher rate of lling and empty, this could be observed in the

occurrence of large hysteresis loop amplitude.

The quasi-steady assumption for the turbine efciency under in-

phase ow condition found to be only partially true. At 50%

equivalent speed and low frequency the cycle-averaged efciency is

approximately 25% lower than the quasi-steady whereas at high

frequency the cycle-averaged and the quasi-steady efciency are

almost coincident (0.8% difference). Meanwhile at 80% equivalent

speed, the quasi-steady assumption is satised at low and high

frequencies (40 Hz and 80 Hz), whereas a transition region from

quasi-steady to fully unsteady was observed at 60 Hz.

The quasi-steady assumption for the turbine efciency under

out-of-phase ow condition was found to be unsatised e

a difference from 18% to more than 30% was measured for both 50%and 80% equivalent speed. The full admission quasi-steady ef-

ciency is not fully representative of the out-of-phase ow condition

in the turbine, which at each instant in time is more likely to act as

in partial admission conditions. At 80% equivalent speed, the ratio

between the cycle-averaged and the quasi-steady efciency devi-

ates only by few percentage points (1.03, 0.85 and 0.92 at 40 Hz,

60 Hz and 80 Hz respectively) while at 50% equivalent speed, an

improvement of almost 10 percentage points could be measured.

The ratio between the cycle-averaged and the quasi-steady

assumption passed from 0.67, 0.70 and 0.82 for the full admission

to 0.81, 0.77 and 0.98 for partial admission, at 40 Hz, 60 Hz and

80 Hz respectively.

The effects of vane angle on the cycle-averaged efciency

provided different response depending on the pulse

ow. In-phaseow showed that at lower frequencies the close vane position (60

and 70 vane angle) is detrimental to the efciencycompared to the

fully open position (40 vane angle). The same does not occur in

out-of-phase ow where higher efciency was found for the

optimum vane angle (60 vane angle). This agrees with thendings

found in the steady-state testing which show a higher efciency for

the optimum vane angle.

Acknowledgements

The authors would like to acknowledge Ricardo Plc., Ford Motor

Company Ltd and University of Brighton. This consortium along

with Imperial College is part of funded program (TSB) named

VERTIGO (Virtual Emission Research Tools and Integration).

0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

40 50 60 70 80

ycneiciffede

garevaelcyC

Frequency (Hz)

In Phase Flow

40 deg 60 deg 70 deg

0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

40 50 60 70 80

ycneiciffed

egarevaelcyC

Frequency (Hz)

Out of Phase Flow

Fig. 16. Comparison of cycle-averaged efciency for different vane angles (40, 60 and 70) for the in-phase and out-of-phase ow at 50% equivalent speed.



14/14

References

[1] Wallace FJ, Blair GP. The pulsating ow performance of inward radial-owturbines. ASME 1965;65(gtp-21):1e19.

[2] Benson RS, Scrimshaw KH. An experimental investigation of non-steadyowin a radial gas turbine. Proc Inst Mech Eng 1965e1966;180(3J).

[3] Wallace FJ, Adgey JM, Blair GP. Performance of inward ow turbines underunsteadyow conditions. Inst Mech Eng Proc 1969e1970;184(1).

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