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Unsteady performance analysis of a twin-entry variable geometry turbocharger turbine


Energy 38 (2012) 176e189

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Energy
journal homepage: www.elsevier.com/locate/energy

Unsteady performance analysis of a twin-entry variable geometry turbocharger turbine
Srithar Rajoo a, Alessandro Romagnoli b, Ricardo F. Martinez-Botas b, *
a b

Transportation Research Alliance, Universiti Teknologi Malaysia, 81310 UTM Skudai Johor, Malaysia 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 ef?ciency 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 con?gurations 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 ef?ciency of the nozzled turbine either twin or single-entry was found to depart signi?cantly 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, twinentry turbines are generally designed and used for better energy extraction from the pulsating exhaust gases. Twin-entry turbine allows the pulsation in the exhaust gas to 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 ampli?ed with the variable geometry con?gurations. The staggered pulsation in both the 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 work to understand the unsteady-state performance of a twin-entry variable geometry turbine in various operating conditions, in comparison to the single-entry and nozzleless unit. This is necessary

* Corresponding author. Tel.: ?44 20 7594 7241; fax: ?44 20 7823 8845. E-mail addresses: srithar@fkm.utm.my (S. Rajoo), a.romagnoli@imperial.ac.uk (A. Romagnoli), r.botas@imperial.ac.uk (R.F. Martinez-Botas). 0360-5442/$ e see front matter ? 2011 Elsevier Ltd. All rights reserved. doi:10.1016/j.energy.2011.12.017

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], Miyashita et al. [4], Benson [5] and Kosuge et al. [6]. All these investigations 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 Gambarotta [9]. Research work by Capobianco et al. [7,8] concentrated on establishing 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 Winterbone and Pearson [11] presented a good review of the experimental 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

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Nomenclature A C D MFP N P PR r T Tr U _ m _ W Area, mm2 Absolute velocity, m/s Rotor Diameter, m Mass ?ow parameter Turbine speed, rps Pressure, Pa Pressure ratio radius, mm Temperature, K Torque, Nm Rotor tip velocity, m/s Mass ?ow rate, Kg/s Power, kW Ef?ciency Phase Angle

j
Cis Subscript actual in ins isent out SE TE tot ts 0 1 5

Azimuth Angle isentropic velocity, m/s

h q

actual conditions inner limb (entry) instantaneous conditions isentropic outer limb (entry) single-entry twin-entry total (inner ? outer limb entry) total-to-static Total conditions Turbine stage inlet Turbine stage exit

performance characteristics under steady, unsteady, equal admission and partial admission conditions. Dale [13] presented results for pulse frequency of 40 and 60 Hz, turbine speed 300e700 RPS, a mean total inlet temperature of 400 K and a maximum instantaneous pressure ratio of 1.8. The turbine was coupled to an eddycurrent 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 acceleration. 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] conducted a study similar to Dale [13] but with a nozzleless turbine. A hydraulic dynamometer was used to load the turbine with higher maximum 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 Watson [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 temperature 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 ef?ciency 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 quanti?ed the turbine ef?ciency depreciation in comparison to quasi-steady assumptions. This paper addresses the turbine unsteady performance under the combination of twin-entry and variable geometry con?gurations,

which has not been shown previously. The presented experimental data and performance characteristics will serve as an essential validation tools for computation models to address the accurate turbocharger component assessments methods in an engine design process. 2. Experimental facility Fig. 1 illustrates the turbine con?gurations used for the current study, which was designed in progression from nozzleless singleentry to nozzled twin-entry. The turbine was designed based on conventional turbomachinery techniques as described in Japikse and Baines [17]. Table 1 gives 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 reciprocating engine test-bed for turbocharger testing. The facility has the 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 (absolute). The two separated streams of air?ow 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 of an engine as shown in Fig. 3. A variable speed D.C. motor controls the rotating frequency of the chopper 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 relative 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 characterizing the turbine behaviour under a given pulse ?ow. The understanding derived from the characterization can be an essential input towards the formulation of a generic assessment tool in engine design process, for any forms of pulsating exhaust gases.

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

3. Unsteady testing and analysis procedures The conditions for the unsteady test points are set up to correspond to the steady-state velocity ratio at peak ef?ciency for a given speed. This is achieved by comparing the energy averaged velocity ratio at unsteady ?ow condition, with the steady-state velocity ratio at peak ef?ciency. 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 concurrently measure all the parameters. The instantaneously acquired data points are processed in few stages before it can be used to calculate the performance parameters. 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

Table 1 Nozzleless vs. nozzled turbine volute dimensions [18]. Nozzleless Volute tongue position Centroid throat radius at ? 0 (mm) Stator throat area (mm2) A/r 50 73.9 2568.6 34.7 Nozzled 30 100.0 3300.0 33.0

every test point during unsteady experiments. For all but the hotwire 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 speci?ed in BS 1042:1983 and only 50 cycles are ensemble averaged. 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 instantaneous 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] (see Fig. 2).

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

4. Unsteady performance parameters 4.1. Mass ?ow parameter vs. pressure ratio The de?nition of the mass ?ow parameter and pressure ratio follows the conventional turbomachinery method, with instantaneous 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. 3. Rotating chopper plate for the air pulse generator to experimentally simulate exhaust gas pulsation.

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T01/P01)

more appropriate. The ?nal equation for the pseudo-dimensionless mass ?ow parameter in a twin-entry turbine is therefore given in _ _ _ Eq. (2), where mins;tot ? mins;in ? mins;out .

8.00 7.00 6.00 5.00 4.00 3.00 2.00 1.00 0.00

x 1e-5
40Hz, 80% Speed

MFPins;TE

s???????????????????????????????????????????????????????????????????? _ _ mins;in mins;out T ? T _ _ mins;tot 01ins;in mins;tot 01ins;out _   ? mins;tot , 0:5 P01ins;in ? P01ins;out

Mass Flow Par. (kg/s

MFPins;SE ?

p??????????? _ mins T01ins P01ins

(1)

(2)

Single Entry Nozzled Nozzeless Steady Nozzeless 1 1.4 1.8 2.2 2.6

The instantaneous pressure ratio is given in Eq. (3). In a twinentry turbine, the pressure ratio can be de?ned using an area averaged pressure ratio between the two entries as given in Eq. (4).

Pressure Ratio (P01/P5) 8.00 Mass Flow Par. (kg/s T01/P01) 7.00 6.00 5.00 4.00 3.00 2.00 1.00 0.00 1 1.4 1.8 2.2 2.6 Pressure Ratio (P01/P5) Single Entry Nozzled
Nozzleless Steady Nozzleless

PRins;SE ?

P01ins P5ins   1 P01ins;in P01ins;out ? P5ins 2 P5ins

(3)

x 1e-5
60Hz, 80% Speed

PRins;TE ?

(4)

4.2. Ef?ciency vs. velocity ratio The turbine instantaneous total-to-static ef?ciency for a singleentry turbine is given in Eq. (5). In the twin-entry cases instead, since power is an extensive quantity, the instantaneous total-tostatic ef?ciency can be calculated considering the sum of the instantaneous isentropic power of the entries, as given in Eq. (6).

?hts ?ins;SE ?

_ W actualins _ W
isentins

(5)

_ W actualins ?hts ?ins;TE ? ? ? _ _ W isent;in ? W isent;out ins

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]).

(6)
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 nozzleless 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. Wave action effect can be de?ned as a continuous interaction between the travelling pulses and the re?ected 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 leaving the turbine as it enters, but at a different rate, which is de?ned as ?lling and emptying effect. It has been shown in many previous works that 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 of ?lling 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

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

pDNins ? s???????????????????? _ 2W
_ mins
isentins

(7)



U Cis

 ?
ins;TE

pDNins ? ? _ _ 2 W isent;in ? W isent;out ins _ 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 bene?cial to look into the unsteady characteristics of the preceding turbine con?gurations, 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 nozzleless volutes, for 40Hz and 60Hz ?ow and 80% equivalent speed. The nozzled turbine loops shown are for

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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. 5 shows the plots of the turbine instantaneous ef?ciency 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 ef?ciency 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 signi?cant in the latter. It can be noticed that the nozzleless turbine exhibit more negative ef?ciencies at the beginning 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 ef?ciency 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 ef?ciency. This re?ects 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 concentrated in the pulse, both the nozzleless and nozzled turbines shows no signi?cant 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 ef?ciency and the equivalent quasi-steady ef?ciency for the nozzleless and nozzled turbines as shown in Fig. 5. It can be seen that the cycle-averaged ef?ciency is generally lower than the quasisteady assumption in most cases. However, in speci?c, 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

2 1.6 1.2
t-s)

6.1. Swallowing capacity of twin-entry turbine

a

0.8 0.4 0 -0.4 -0.8 -1.2 -1.6 -2 0 90 180 Phase Angle (Deg) 270 360
40 Hz 60 Hz

Nozzleless 80%Speed

2 1.6 1.2
t-s)

b

0.8 0.4 0 -0.4 -0.8 -1.2 -1.6 -2 0 90 180 Phase Angle (Deg) 2 1.6 1.2 270

40deg 50deg 65deg 70deg 60deg

40Hz, 80%Speed 360

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 pro?les 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 8 show the swallowing capacity of the twin-entry turbine during in-phase full admission test at 50% and 80% equivalent speed conditions respectively. In both the ?gures, results for 40 Hz and 60 Hz pulsating ?ow conditions 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 conditions. It can be seen that the outer limb swallowing capacity tends towards steady characteristics during the 50% speed, but falls below the equivalent quasi-steady. However during 80% speed condition,

Efficiency (

Efficiency (

c

t-s)

0.8 0.4 0 -0.4 -0.8 -1.2 -1.6 -2 0 90 180 Phase Angle (Deg) 270

40deg 50deg 65deg 70deg 60deg 60Hz

Table 2 Energy weighed cycle-averaged values of velocity ratio and ef?ciency for singleentry nozzled and nozzleless turbine, 80% equivalent speed conditions (nozzleless results from Szymko, [24]). Frequency & con?guration 40 40 40 40 40 40 Hz Hz Hz Hz Hz Hz Hz Hz Hz Hz Hz Hz 40deg 50deg 60deg 65deg 70deg nozzleless 40deg 50deg 60deg 65deg 70deg nozzleless U/Cis 0.65 0.662 0.652 0.645 0.619 0.624 0.676 0.684 0.682 0.662 0.62 0.653

Efficiency (

hcycle-avg
0.604 0.607 0.613 0.603 0.588 0.626 0.657 0.69 0.686 0.64 0.556 0.691 (?2.8%) (?3.1%) (?12.6%) (?15.4%) (?14.6%) (?4.6%) (?7.3%) (?4.2%) (?7.1%) (?13.2%) (?19.4%) (?0.8%)

hquasi-steady
0.576 0.638 0.739 0.757 0.734 0.672 0.584 0.648 0.757 0.772 0.75 0.683

60Hz, 80%Speed 360

Fig. 5. Comparison of the ef?ciency 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]).

60 60 60 60 60 60

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240 40Hz 80% Speed In-Phase Full Admission Inlet Static Pressure (kPa) Inlet Static Pressure (kPa) 220 200 180 160 140 120 100 0 45 90 135 180 225 270 315 360 Phase Angle (deg) 190
40Hz 50% Speed 180 Out-of-Phase Admission Outer Limb Inner Limb

240 220 200 180 160 140 120 100 0 45 90 135 180 225 270 315 360 Phase Angle (deg) 190 Inlet Static Pressure (kPa) 180
Outer Limb 60Hz 50% Speed 180° Out-of-Phase Admission 60Hz 80% Speed In-Phase Full Admission

Outer Limb Inner Limb

Inlet Static Pressure (kPa)

180 170 160 150 140 130 120 110 100 0 45 90 135 180 225

Outer Limb Inner Limb

170 160 150 140 130 120 110 100

Inner Limb

270

315

360

0

45

90

135

180

225

270

315

360

Phase Angle (deg)

Phase Angle (deg)

Fig. 6. Twin-entry turbine typical inlet static pressure pro?les 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 inphase 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 inphase full admission: given for individual entries, inner and outer limbs as well as mass averaged combination.

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

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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 from Figs. 7 and 8 is 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 partial admission is different between each other only in the unsteady ?ow operation. The swallowing capacity of a turbine is in?uenced 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 operation. Even though the current experimental results do not allow a complete understanding of such a complex behaviour, one can indicate that the ?ow has insuf?cient time to stabilize in the unsteady ?ow, 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 may lead 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. 9 shows 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 covered during a cycle, compared to the individual entry. This can be seen clearly in Fig. 9, as well as the double looping in the overall swallowing capacity representing both entries. 6.2. Power and ef?ciency 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

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 mass averaged combination.

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 admission does 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 ef?ciency representation for majority of the pulse period. However, the instantaneous pressure pro?le, 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 at the higher frequency, as it was observed even at 80 Hz (not discussed here). At higher frequency, the emptying rate of the nozzled 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 ef?ciency however instantaneous 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 de?ning the ef?ciency of a turbine under unsteady ?ow conditions, Rajoo and Martinez-Botas [25].

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90 75 60 Power (kW) 45 30 15 0 0

40Hz 80% Speed In-Phase Full Admission

90
Wisnt Wact

75 60 45 30 15 0

60Hz 80% Speed In-Phase Full Admission

Wisnt Wact

Power (kW)

45

90

135

180

225

270

315

360

0

45

90

135

180

225

270

315

360

Phase Angle (deg) 25 20 Power (kW) 15 10 5 0 0 45 90 135 180 225 270 315 360 Phase Angle (deg) Power (kW)
40Hz 50% Speed 180° Out-of-Phase Admission

Phase Angle (deg) 25 20 15 10 5 0 0 45 90 135 180 225 270 315 360 Phase Angle (deg)

Wisnt Wact

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

Wisnt Wact

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

6.3. Cycle averaging and comparison to quasi-steady Given the uncertainty associated with the meaning of the instantaneous turbine ef?ciency, 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 ef?ciency represents the ratio of energy extracted by the turbine per pulse cycle divided by isentropic energy ?owing into the system, as given in Eq. (9).

Zq ?hts ?energy?avg ?
0

?

? _ hts ? W?t?isent dt Zq
0

Zq ?
0

?

? _ W?t?act dt

_ W?t?isent dt

Zq
0

_ W?t?isent dt (9)

? ?hts ?cycle?avg

The velocity ratio is calculated with energy weighed averaging as shown in Eq. (10),1

Zq  U Cis  ?
energy?avg 0

! U _ ,W?t?isent dt Cis ?t? Zq
0

(10) _ W?t?isent dt

The calculated energy weighed cycle-averaged velocity ratio is then used to read the corresponding ef?ciency from the steady

1 _ _ _ In a twin-entry con?guration W isent ? W isent;in ? W isent;out , while the velocity ratio U/Cis is equal to that provided in Eq. (8).

map for that given speed. The cycle-averaged ef?ciency is then compared with the equivalent quasi-steady ef?ciency. By doing this an appropriate comparison between parameters obtained at different conditions can therefore be performed. Table 3 reports a comparison between the cycle-averaged and the quasi-steady ef?ciency obtained for 60 vane angle at 50% and 80% equivalent speeds under in-phase and out-of-phase ?ow, for 40 Hz, 60 Hz and 80 Hz frequencies. It is worth noting that the full admission assumption has been considered here for the quasisteady 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 12 show the ef?ciency 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 ef?ciency remains below the quasi-steady ef?ciency 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 signi?cant and in some cases the cycle-averaged ef?ciency is higher than the quasi-steady ef?ciency. This occurs at 80 Hz and 40 Hz where the cycle-averaged ef?ciency 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 smaller which makes the rotor exposed to more continuous ?ow. At lower frequency instead, the ?ow follows a ?lling and emptying behaviour, thus signi?cant swallowing capacity changes are experienced leading to the differences between the cycle-averaged and quasi-steady ef?ciencies. Similar results were also shown by Karamanis and Martinez-Botas [26], Szymko et al. [23,24] and Rajoo and Martinez-Botas [25], who measured higher cycle-averaged ef?ciency in a single-entry turbine for higher speeds and frequencies. However, it must be noted that no consistent trend

S. Rajoo et al. / Energy 38 (2012) 176e189 Table 3 Comparison of energy weighted cycle-averaged and quasi-steady ef?ciency. 60 vane angle

185

Dh ? ??hcycle?avg =hquasi?steady ? ? 1? ? 100%
50% equivalent speed U/Cis In-phase 40 Hz 60 Hz 80 Hz 0.542 0.503 0.533

Dh ? ??hcycle?avg =hquasi?steady ? ? 1? ? 100%
80% equivalent speed Δh ?25.4% ?25.4% 0.8% ?32.9% ?29.2% ?17.9%

hcycle-avg
0.472 0.456 0.631 0.434 0.415 0.526

hquasi-steady
0.631 0.611 0.626 0.647 0.586 0.641

U/Cis 0.640 0.596 0.629 0.595 0.587 0.577

hcycle-avg
0.835 0.616 0.823 0.631 0.520 0.560

hquasi-steady
0.784 0.768 0.781 0.768 0.763 0.757

Δh 6.5% ?19.8% 5.4% ?17.8% ?31.8% ?26.0%

Out-of-phase 40Hz 0.587 60Hz 0.464 80Hz 0.567

In-Phase Flow 0.9 0.8 0.7 Efficiency 0.6 0.5 0.5 0.4 0.4 0.3 40 50 60 70 80 0.3 0.4 0.45 0.5 0.55 U/Cis Cycle avrg eff - 27.9 revs/K^0.5 50% Equiv Speed 80% Equiv Speed Cycle avrg eff - 43.0 revs/K^0.5 Quasi-steady - 27.9 revs/K^0.5eed 50% Equiv Sp Quasi-steady - 43.0 revs/K^0.5 80% Equiv Speed 0.6 0.65 0.7 Frequency (Hz) Efficiency
Efficiency 40 50 60 70 80

0.9

0.8

0.7

0.6

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

Out-of-Phase Flow 0.8 0.75 0.7 0.65 Efficiency 0.6 0.55 0.5 0.45 0.4 0.35 0.3 Frequency (Hz) 50% Equiv Speed Cycle avrg eff - 27.9 revs/K^0.5 80% Equiv Speed Cycle avrg eff - 43.0 revs/K^0.5 0.8 0.75 0.7 0.65 0.6 0.55 0.5 0.45 0.4 0.35 0.3 0.4 0.45 0.5 0.55 U/Cis 50% Equiv Speed Quasi-steady - 27.9 revs/K^0.5 80% Equiv Speed Quasi-steady - 43.0 revs/K^0.5 0.6 0.65

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.

186

S. Rajoo et al. / Energy 38 (2012) 176e189
Ratio between Cycle-averaged and Quasi-steady efficiency

1.1

1 0.9 0.8 0.7 0.6 0.5 40 50 60 70 80

In Phase 50% Equiv Speed

In Phase 80% Equiv spped

Out of phase 50% Equiv Speed

Out of phase 80% Equiv Speed

Frequency (Hz)
Fig. 13. Ratio between cycle-averaged and quasi-steady ef?ciency for the in-phase and out-of-phase ?ow at 50% and 80 equivalent speed.

was observed in the cycle-averaged ef?ciency variation. In the outof-phase ?ow conditions, the cycle-averaged ef?ciency 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 ef?ciency deviates from the quasi-steady by 17.8%, 31.8% and 26.0% for the 40 Hz, 60 Hz and 80 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 ef?ciency. From Table 3 it can be seen that the out-of-phase condition is detrimental to the overall ef?ciency above all at high speed. At 80% equivalent speed, the cycle-averaged ef?ciency 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 condition e this corresponds to a drop of 32.2%, 18.4%, and 46.9% respectively. However, the same de?cit is not observed at lower speed where the ef?ciency drop is much lesser. At 50% equivalent speed, the ef?ciency 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 outof-phase ?ow condition. In order to further understand the correlation between the cycle-averaged and the quasi-steady assumption, the ratios between these ef?ciencies has been calculated for both in-phase and out-phase ?ow at 50% and 80% equivalent speeds. The parameter called ef?ciency ratio is plotted in Fig. 13 against pulse frequency (40 Hz, 60 Hz, 80 Hz). As noted from Fig. 13, at 80%

equivalent speed, the ef?ciency 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 (ef?ciency ratio z0.8) 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 ef?ciency ratio remains below unity for both 50% and 80% equivalent speeds. In order to understand the ef?ciency trends, the unsteady timeaveraged power and mass ?ow were calculated and compared to the quasi-steady values e these are shown in Fig. 14 for 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 same does 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 observed before (Fig.10) such a dip could be attributed to the increasingly unsteady correlations between the wave action and ?lling and emptying effects in the nozzled turbine system. On the ef?ciency side, the trend observed for the ratios between the cycleaveraged and the quasi-steady ef?ciencies (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 ef?ciency 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 conditions in most cases. The incoming pulses from each bank of manifolds occur at staggered intervals, thus the turbine works in partial admission conditions in most cases compared to full admission. Therefore, in order to evaluate the quasi-steady assumption in the out-of-phase ?ow conditions, it is appropriate to refer to the partial admission maps instead of the full admission. Table 4 shows the quasi-steady ef?ciency calculated using the full and partial admission conditions. The cycle-averaged ef?ciencies and the ratio

Unsteady/Quasi-steady ratio

In Phase Flow
2

Out of Phase Flow
2

Unsteady/Quasi-steady ratio

1

Unsteady/Quasi-steady ratio 40 60 Frequency (Hz)
Wisnt ratio Wis ratio Wact ratio

1

0

0

80

40

60 Frequency (Hz)
Mass flow ratio

80

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.

S. Rajoo et al. / Energy 38 (2012) 176e189 Table 4 Comparison between cycle-averaged and quasi-steady ef?ciency considering full and partial admission assumption for the quasi-steady value. Out-of-phase ?ow (quasi-steady assumption based on partial admission condition) 50% equivalent speed U/Cis 40 Hz 60 Hz 80 Hz 0.587 0.464 0.567 80% equivalent speed

187

hcycle-avg
0.434 0.415 0.526

hqs, full
0.647 0.586 0.641

hratio
0.67 0.70 0.82

hqs,partial
0.530 0.533 0.534

hratio
0.81 0.77 0.98

U/Cis 0.595 0.587 0.577

hcycle-avg
0.631 0.520 0.560

hqs,full
0.768 0.763 0.757

hratio
0.82 0.68 0.73

hqs, partial
0.607 0.607 0.608

hratio
1.03 0.85 0.92

between the cycle-averaged and the quasi-steady ef?ciencies are shown in Fig. 15. From Fig. 15 it can be seen that the ratio between the cycle-averaged and the quasi-steady assumption based on the partial admission condition 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 ef?ciency goes 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 condition e such a drop leads to values of the ef?ciency ratio (in partial 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 cycleaveraged and quasi-steady ef?ciency is approximately 10% higher than 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 ef?ciency 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 admission case. Table 5 shows the velocity ratio, cycle-averaged ef?ciency at 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 cycleaveraged ef?ciency for different vane angles does not seem to follow a well-de?ned pattern. For the in-phase ?ow, the cycleaveraged ef?ciency shows about 12 percentage points drop from 40 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 ef?ciency [25]. For 70 vane angle and in-phase ?ow 40 Hz and 60 Hz, the cycle-averaged ef?ciency 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 cycleaveraged ef?ciency increases with increasing frequency. At 60

vane angle, the ef?ciency is higher than those measured at 40 and 70 , while the 40 vane angle setting seems to perform better than the 70 . The ef?ciency de?cit 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 ef?ciency could be attributed to the blockage effects due to closed nozzle settings, similar to what was found in the single-entry investigations [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 con?gurations under steady-state conditions [18]. The results obtained in steady-state conditions showed that the mass ?ow parameters for the twin-entry are almost independent of the con?guration and of the type of ?ow admission (partial or 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 identi?ed 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 con?gurations, 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

1.1

Unsteady/Quasi-steady ratio

1 0.9 0.8 0.7 0.6 0.5
40 50 60 70 80

Full Admission 50% Equiv Speed revs/K^0.5 Full Admission 80% Equiv Speed revs/K^0.5 Partial Admission 50% Equiv Speed revs/K^0.5 Partial Admission 80% Equiv Speed revs/K^0.5

Table 5 Energy weighted cycle-averaged ef?ciency for different vane angles. 50% equivalent speed 40 vane angle U/Cis In-phase 40 Hz 60 Hz 80 Hz 0.593 0.548 0.519 60 vane angle U/Cis 0.542 0.503 0.533 0.587 0.464 0.567 70 vane angle U/Cis 0.51 0.541 0.487 0.472 0.478 0.485

hcycle-avg
0.593 0.590 0.347 0.425 0.344 0.433

hcycle-avg
0.472 0.456 0.631 0.434 0.412 0.526

hcycle-avg
0.470 0.463 0.288 0.261 0.311 0.349

Frequency (Hz)
Fig. 15. Ratio of cycle-averaged and quasi-steady ef?ciency for the out-of-phase ?ow at 50% and 80% equivalent speed, obtained using the full and partial admission quasisteady value.

Out-of-phase 40 Hz 0.556 60 Hz 0.524 80 Hz 0.587

188

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

In Phase Flow 0.7 0.6 Cycle averaged efficiency 0.5 0.4 0.3 0.2 0.1 0 40 50 60 Frequency (Hz) 70 80 Cycle averaged efficiency 0.7 0.6 0.5 0.4 0.3 0.2 0.1 0 40

Out of Phase Flow

50

60 Frequency (Hz)

70

80

40 deg

60 deg

70 deg

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

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 simulation in the aim to discover more physical insights of the single and multiple entries turbine behaviour under pulsating ?ow conditions. 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 bene?cial 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 still based on steady-state ?ow. 7. Summaries The current paper discusses performance analysis of a twinentry 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 progression 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 nozzled single-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 low pressure 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 hysteresis 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 ef?ciency under inphase ?ow condition found to be only partially true. At 50% equivalent speed and low frequency the cycle-averaged ef?ciency is approximately 25% lower than the quasi-steady whereas at high frequency the cycle-averaged and the quasi-steady ef?ciency are almost coincident (0.8% difference). Meanwhile at 80% equivalent speed, the quasi-steady assumption is satis?ed 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 ef?ciency under out-of-phase ?ow condition was found to be unsatis?ed 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 ef?ciency deviates 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 ef?ciency provided different response depending on the pulse ?ow. In-phase ?ow showed that at lower frequencies the close vane position (60 and 70 vane angle) is detrimental to the ef?ciency compared to the fully open position (40 vane angle). The same does not occur in out-of-phase ?ow where higher ef?ciency was found for the optimum vane angle (60 vane angle). This agrees with the ?ndings found in the steady-state testing which show a higher ef?ciency 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).

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

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