8th International conference on turbochargers and turbocharging

8th International Conference on Turbochargers and Turbocharging

Organising Committee Kian Banisoleiman (Chair) Steve Birnie Henry Tennant Ennio Codan Les Smith Andrew Stapleton Ricardo Martinez-Botas ]oerg Seume Chris Brace ] ohn MardeII

Lloyd's Register EMEA Borg-Warner Automotive Holset Turbochargers ABB Turbo Systems MIRA QinetiQ Imperial College London University of Hannover University of Bath Consultant

8th International Conference on Turbochargers and Turbocharging Institution of Mechanical Engineers Combustion Engines & Fuels Group

• IDGTE

CRC Press Boca Raton Boston New York Washington, DC

WOODHEAD PUBLISHING LIMITED Cambridge England

Published by Woodhead Publishing Limited, Abington Hall, Abington, Cambridge CB 1 6AH, England www.woodheadpublishing.com Published in North America by CRC Press LLC, 6000 Broken Sound Parkway NW, Suite 300, Boca Raton, FL 33487, USA First published 2006, Woodhead Publishing Limited and CRC Press LLC © 2006, Institution of Mechanical Engineers unless otherwise stated The authors have asserted their moral rights. This book and CD-ROM contain information obtained from authentic and highly regarded sources. Reprinted material is quoted with permission, and sources are indicated. Reasonable efforts have been made to publish reliable data and information, but the authors and the publishers cannot assume responsibility for the validity of all materials. Neither the authors nor the publishers, nor anyone else associated with this publication, shall be liable for any loss, damage or liability directly or indirectly caused or alleged to be caused by this book or CDROM. Neither this book, CD-ROM, nor any part thereof may be reproduced or transmitted in any form or by any means, electronic or mechanical, including photocopying, microfilming and recording, or by any information storage or retrieval system, without permission in writing from Woodhead Publishing Limited. The consent of Woodhead Publishing Limited does not extend to copying for general distribution, for promotion, for creating new works, or for resale. Specific permission must be obtained in writing from Woodhead Publishing Limited for such copying. Further terms and conditions concerning the CD-ROM are included on the CD-ROM. Trademark notice: Product or corporate names may be trademarks or registered trademarks, and are used only for identification and explanation, without intent to infringe. British Library Cataloguing in Publication Data A catalogue record for this book is available from the British Library. Library of Congress Cataloging in Publication Data A catalog record for this book is available from the Library of Congress. Woodhead Publishing ISBN-13: 978-1-84569-174-5 Woodhead Publishing ISBN-lO: 1-84569-174-1 CRC Press ISBN-lO: 0-8493-0777-5 CRC Press order number: WP0777 Printed by Antony Rowe Limited, Chippenham, Wilts, England

CONTENTS PART I: COMPRESSORS AND NOVEL INTAKE SYSTEMS 1 Prediction and measurement of turbocharger compressor wheel temperature A Yamagata, S Nagai, K Nakano, and T Kawakubo, Ishikawajima-Harima Heavy Industries Company Limited, Yokohama, Japan

3

2 Turbocharger compressor development for diesel passenger car applications H Chen and J F Yin, Honeywell Turbo Technologies Limited, Skelmersdale, UK

15

3 The reduction of turbocharger whoosh noise for diesel powertrains D Evans, Ford Motor Company Limited and A Ward,Ricardo UK Limited

29

4 The influence of installation parameters on turbocharged automotive engine performance G Capon, A Leong, and T Morris, Ford Motor Company Limited, UK

43

5 Using the centrifugal compressor as a cold air turbine M Miiller, S Sumser, P Fledersbacher, K Rofiler, and K Fieweger, DaimlerChrysler AG, Germany, and H-J Bauer, University of' Karlsruhe, Germany

55

6 Extending the knock limit of a turbocharged gasoline engine via turboexpansion J W G Turner, R J Pearson, and N Milovanovic, Lotus Engineering, UK, and D W Taitt, Loughborough University, UK

69

7 Turbo-cooling applied to light duty vehicle engines CD Whelan and R A Richards, WDL Limited, UK

81

PART II: IMPROVED MODELS FOR CYCLE SIMULATION 8 A one-dimensional model for variable and fixed geometry radial turbines for turbochargers J M Lujan, J R Serrano, C Cervell6, and F J Arnau, Universidad Politecnica de Valencia, Spain, and S Soltani, Renault, France

97

9 Analysis of turbocharger non-adiabatic performance S Shaaban and J R Seume, University of Hannover, Germany

119

10 Part-load performance prediction of turbocharged engines S Shaaban and J Seume, University of Hannover, Germany, R Berndt, Technical University Berlin, presently Ingenieurgesellschaft Auto und Verkehr IA V GmbH, Germany, H Pucher, Technical University Berlin, Germany and H J Linho.fj; Linhoff Engineering, Germany

131

v

PART III: ELECTRO BOOST SYSTEMS 11 Development of electrically assisted turbocharger for diesel engine Y Yamashita, S Ibaraki, and H Ogita, Mitsuhishi Heavy Industries Limited, Tokyo, Japan

147

12 The design and testing of an electrically assisted turbocharger for heavy duty diesel engines o Ryder, Holset Engineering Company Limited, Huddersjield, UK, H Sutter, ATE GmbH, Germany, and L Iaeger, Iveco Motorenforschung AG, Switzerland

157

PART IV: TURBINES 13 A numerical study of the performance characteristics of a radial turbine with varying inlet blade angle L Barr, S W T Spence, and A McNally, Queen's University Belfast, UK

169

14 Experimental study on the performance of a variable geometry mixed flow turbine for automotive turbocharger S Rajoo and R Martinez-Botas, Imperial College London, UK

183

15 Turbocharger turbine performance under steady and unsteady flow: test bed analysis and correlation criteria M Capobianco and S Marelli, University of Genoa, Italy

193

16 Flexible turbocharger turbine test rig MONA VI D Filsinger, G Fitzky, and B Phillipsen, ABB Turbo Systems Limited, Baden, Switzerland

207

17 Active control turbocharger for automotive application: an experimental evaluation A Pesiridis and R Martinez-Botas, Imperial College London, UK

223

PART V: MECHANICAL ASPECTS 18 Thermomechanical analysis of a turbocharger turbine wheel based on CRTcalculations and measurements T Heuer, B Engels, H Heger, and A Klein, BorgWarner Turbo Systems Engineering GmbH, Germany 235 19 Dynamics of mistuned radial turbine wheels X Sheng, D C Clay, and I Allport, Holset Engineering Company Limited, Huddersfield, UK

251

20 Improving analysis capability in order to reduce turbine RCF S T Kitson, D C Clay, D H Brown, R 0 Evans, D M Eastwood and P K Tootill, Holset Engineering Company Limited, Huddersfield, UK

261

VI

21 Axial load capacity of V-section band clamp joints K Shoghi, BorgWarner Turbo Systems, Bradford, UK, S Barrans and P Ramasamy, University of Huddersfield, UK

273

PART VI: ADDITIONAL PAPERS 22 Reliability trends, operating issues and acceptance criteria related to exhaust gas turbochargers used in the marine industry - a classification society view K Banisoleiman and N Rattenbury, Lloyd's Register, London, UK

289

23 A novel method of high efficiency pressure charging A 0 Dye, Epicam Limited, Linton, Cambridgeshire, UK

305

24 Turbine wheel design for Garrett advanced variable geometry turbines for commercial vehicle applications H Chen, Honeywell Turbo Technologies Limited, Skelmersdale, UK

3 17

25 Compact long-route exhaust gas recirculation mixer design and optimization J Yin, N Deschatrettes, 0 Han, and P Renaud, Honeywell Turbo Technologies Limited, Skelmersdale, UK

329

26 Transient performance prediction of the turbocharging system with the variable geometry turbochargers H Uchida, A Kashimoto, and Y Iwakiri, Toyota Central R&D Laboratories Incorporated, Aichi, Japan

341

27 Plain and full floating bearing simulations with rigid shaft dynamics I McLuckie, S Barrett, and B K Teo, Advanced Integrated Solutions Limited, Market Harborough, Leicestershire, UK

351

VII

Prediction and measurement of Turbocharger compressor wheel temperature A. Yamagata, S. Nagai, K. Nakano and T. Kawakubo Ishikawajima-Harima Heavy Industries Co., Ltd., Yokohama, JAPAN

ABSTRACT Conjugate heat transfer (CHT) analysis for a high-pressure ratio turbocharger compressor has been conducted to estimate the temperature distribution of a rotating impeller. CHT analysis have been performed with several thermal boundary conditions, which are temperature around a compressor housing, of a back plate of the impeller and of a turbine shaft. We revealed the effects of these conditions on the compressor impeller temperature. Using a radiation thermometer, temperature measurement of a rotating impeller also has been performed. It was utilized to verify the accuracy of CHT analysis for the temperature prediction of a rotating impeller. Subscripts blp Back plate of impeller c/h Compressor housing exit Compressor exit oil Lubricant oil ref Reference condition shaft Turbine shaft tip Impeller tip

Nomenclature Diameter Mu Rotation Mach number N Rotation speed Q Volumetric flow rate T Temperature Tt Total temperature Zb Number of blades It c Compressor total pressure ratio

1)

1. INTRODUCTION In recent years, the need of Diesel engines for passenger vehicles is increasing in the worldwide and the turbo charging for Diesel engines becomes very important due to the economical and environmental reasons. In order to improve the output power and the fuel consumption and also to suppress the emission of Diesel engines, a high efficiency and high-pressure ratio turbocharger is required. Since a single stage centrifugal compressor is used usually in a turbocharger for a passenger vehicle, high rotational speed is necessary to achieve high-pressure ratio at a turbocharger compressor. Therefore, due to the increase of rotational speed, the compressor wheel, impeller, is exposed to high centrifugal stress. High-pressure ratio also leads the increase of discharged air temperature, and this causes higher material temperature and consequently the impeller material strength decreases. In order to guarantee the lifetime of a turbocharger, it is important to know the impeller metal temperature correctly. However, there is much difficulty in a direct temperature measurement of rotating parts under high rotational speed, because thermal sensors set on

3

the impeller surface cannot bear the high centrifugal force. Therefore, analytical method that can predict the impeller metal temperature at high-pressure ratio accurately is desired. As a practical method to predict the impeller metal temperature, heat transfer analysis assuming a heat transfer coefficient on the impeller surface and fluid temperature near the wall as thermal boundary conditions has been applied. Mukherjee and Baker [1] have calculated the metal temperature distribution of a high-pressure ratio turbocharger compressor wheel using a heat transfer analysis. They applied a heat transfer coefficient derived from empirical method, and also have investigated thermal stress occurred at the impeller together with centrifugal load. In the aspect of taking account of heat transfer between solid walls and fluid simultaneously, conjugate heat transfer (CHT) analysis can be useful to obtain the metal temperature distribution with high accuracy. One of the main issues of CHT analysis is computational cost, i.e. limitation of memory size to run a calculation and computational time to obtain a converged solution. However, recent progress of computer hardware and parallel computing technology has realized a large CHT calculation. Bohn [2] has conducted a CHT analysis for the whole of a turbocharger, including a compressor, a turbine and a bearing housing. He described about the effect of heat transfer from a turbine to a compressor through a bearing on a compressor performance. Heuer [3] also has done CHT calculations of some twin-entry turbine housings with an integrated manifold, and avoided a thermal shock occurred at a turbine scroll by the modification of a turbine housing geometry. In the present study, in order to construct the temperature prediction method of a rotating impeller analytically, CHT analyses for a high-pressure ratio turbocharger compressor with high rotational speed have been conducted. This analysis includes a compressor housing and a back plate of the impeller. Further, the metal temperature of the rotating compressor impeller has been measured using a radiation thermometer and the accuracy of CHT analysis was verified.

2. METHODOLOGY 2.1 Compressor configuration RHG8V turbocharger with variable geometry turbine system, which has been developed in IHI for truck size Diesel engines, is chosen for this study. Figure 1 shows a cross sectional view of the RHG8V turbocharger, including the information of a selected computational domain and temperature measurement locations. This turbocharger compressor has a cavity inside a shroud casing, which is called "Casing Treatment". This is made to enhance the compressor operating range by the flow re-circulation through this cavity, and included in a computational model used in this study. The compressor design parameters are summarized in Table 1. Total pressure ratio of the compressor :rtc becomes 3.5 at 100% test rotation speed, and the discharged air temperature reaches over 200 deg. C. The computational domain is limited only to compressor side, which includes a compressor impeller, a back plate behind the impeller, a compressor housing and a shaft connected to a turbine. A bearing housing and a turbine are not included. At the interface between a compressor and a bearing or between a compressor housing and atmosphere, temperature or heat flux is given as a thermal boundary condition.

4

Temperature measurement of the rotating impeller is performed at the compressor inlet and the back surface of the impeller using an infrared thermo camera and a radiation thermometer, respectively, which are shown in Figure 2. Further, thermo couples are set in stationary parts, which are a compressor housing and a back plate, and stationary parts temperatures are measured and used for the verification and the boundary conditions of CRT analysis. Table 1. Specification of RRG8V turbocharger compressor. 7 +7

Number of blades

Zb

Impeller diameter

Dtip

Impeller tip Mach number (*)

MUtip

Total pressure ratio (*)

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(mm)

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5

Radiation

a) Radiation thermometer.

b) Infrared thermo camera.

Figure 2. Thermometers for the compressor wheel temperature measurement. 2.2 Computational model The computational model used in this study is shown in Figure 3. Since a compressor impeller is composed of 7 full-length blades and the same number of short blades, the computational domain is reduced to a 1/7 sector of the annulus to save the computational cost and the periodic boundary condition is applied circurnferentially. Here, though the compressor housing has a non-axisymmetric geometry, a volute region is simulated by the representative geometry of a volute section for the simplicity. At the outer surfaces of solid regions, which are a compressor housing, a back plate of the impeller and a shaft end to a turbine, thermal boundary conditions are applied, i.e. surface temperature or heat transfer coefficient and neighbor temperature are given, and heat fluxes through these boundaries are determined by the heat balance between the interior and the adjacent boundary. Computation is the steady-state calculation performed by FLUENT version 6.2. Computational grid is generated by GAMBIT, a pre-processor of FLUENT, and mainly consists of tetra mesh. Solid and fluid region have about 1.0 million and 1.5 million cells respectively and total number of grids becomes 2.5 million cells. As a flow and heat transfer solver, the segregated solver within FLUENT 6.2 is selected and the Spalart-Allmaras one-equation model determines the turbulent viscosity of the fluid. In order to control stability and convergence of the CHT calculation, the relaxation factor for each equation is adjusted appropriately. Calculation has been done by Intel Xeon multiprocessor machine and continued till mass flow rate, total pressure and total temperature at inlet and outlet boundary and the metal temperature became converged. About 2 days of computational time is required to obtain a converged solution. 2.3 Boundary conditions As flow boundary conditions, uniform total pressure, total temperature and flow angles are fixed at the inlet boundary and static pressure is fixed at the exit boundary. Thermal boundary conditions are summarized in Table 2. CRT analysis has been performed for following four typical conditions and the effect of each condition on the impeller metal temperature are revealed. 1) Outer surfaces are all adiabatic, i.e. there is no heat transfer through the outer surface of the computational domain.

6

2) In order to evaluate the effect of heat transfer through the rotor shaft on the impeller metal temperature, the thermal boundary condition at the shaft end is changed parametrically. At first, it is assumed that the shaft temperature at the compressor side is the same as the lubricant oil temperature. After that, the normalized shaft end temperature is increased to +0.17 and +0.34. Other surfaces are all adiabatic. 3) In order to evaluate the effect of heat transfer through the outer surface of the compressor housing, the thermal condition on this surface is changed. Assuming the heat transfer coefficient at the outer surface of a compressor housing at a fixed value, the neighbor atmosphere temperature is increased from the reference temperature to +0.17 and +0.34. Other surfaces are all adiabatic. 4) In order to evaluate the effect of heat transfer through the back plate surface, the back plate surface temperature is changed parametrically. First, the temperature distribution of the back plate obtained by the stationary part temperature measurement is applied to a solid region of the back plate. After that, the temperature distribution level is changed to both -0.17 and +0.17. Other surfaces are all adiabatic. - No. of grids : Solid 1.0 million : Fluid 1.5 million - Grid type : Tetrahedron

a) Side view of the computational domain

b) Impeller and shroud cavity.

Figure 3. Computational model of CRT analysis. Table 2. Summery of thermal boundary conditions. Case

(1)

(2)

(3)

(4)

LlT shaft

adiabatic

0, +0.17, +0.34

adiabatic



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Figure 1: Spark ignition passenger car engine, simulated acceleration processes from a driving speed of 20 km/h in second gear

THERMODYNAMIC SYSTEM ASSESSMENT Using the enthalpy hidden in the throttling process of a SI engine is supported by a fundamental thermodynamic assessment of the overall system. The following calculation uses the results measured from a turbocharged SI passenger car engine with a displacement of 1.8 I to determine the enthalpy difference on the intake side between compressor entry PI and the pressure downstream throttle valve P2,s (Figure 2, left), and the exhaust gas side, between exhaust gas turbocharger turbine intake P3 and the turbine P4 (Figure 2, right). At high engine speeds and loads, a large isentropic enthalpy difference is available on the turbine side. This allows for high air compression in the exhaust gas turbocharger compressor. In contrast, the enthalpy difference in the left bottom part of the engine map (low engine speed and low brake mean effective pressure) is very small. This is caused by the ineffective back pressure behavior of the exhaust gas turbocharger turbine at small mass flow rates. Bearing in mind the isentropic energy potential of throttling on the air side (Figure 2, left) for the partial load operating points at the lower engine speed range, it is clear that some of this is operates at a much higher level than in the exhaust gas turbine (Figure 2, right). If this throttling energy could be used, significantly higher exhaust gas turbocharger speeds would be achieved at the partial load points.

57

The assessment in Figure 2 shows the maximum available energy potential; any losses or component efficiencies were intentionally ignored so it would not influence assessment of the theoretical potential. The efficiencies of the overall DOT system must be identified and optimized for the aspects relevant to engine operation. • Turbocharged SI Engine

a. w :; co

• Displacement 1.8 I

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0

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o 1000

1500

2000

2500

3000

3500

1000

4000

Engine Speed [rpm]

1500

2000

2500

3000

3500

4000

Engine Speed [rpm]

Figure 2: Engine map of a 1.8 I turbocharged spark ignition passenger car engine

The effects of using the enthalpy potential of throttling on the turbocharger speed are significant. While equation (I) yields the power balance at the turbocharger shaft in traditional operation, equation (2) applies to throttle-controlled operation. It is assumed that the compressor impeller will act as a cold air turbine.

+ PFriction

PTurbine

=

PTurbine

+ PColdAirTurbine

Pcompressor

(1)

= PFriction

(2)

Since the right side of the equation describes the "consumers," bearing friction is the only power consumer in cold air turbine operation. In throttle engine operation there are two power sources: the hot gas turbine and the compressor acting as a cold air turbine. The thermodynamic assessment in Figure 2 shows that the cold air side can provide 500 Watts (linear interpolation) of isentropic energy at engine speeds of 1,500 rpm and a break mean effective pressure of 2 bars. Assuming a cold air turbine efficiency of 20 percent, the cold air turbine should provide about 100 Watts of energy. This energy, and that of the hot gas turbine, will then be in balance with the bearing friction loss at the resulting turbocharger speed. When using traditional journal bearings, the power (shown in Figure 3) needs to be overcome. This should allow for maintaining the turbocharger speed at 50,000 rpm using cold air turbine power alone.

58

Hence, the real benefit of the cold air turbine concept is obvious. It should be remembered, however, that the compressor wheel operating at 50,000 rpm and the minimum possible mass flow (limit comes from the surge line) require a power input of 400 Watts. Therefore, even a power output of just 100 Watts, which corresponds to a very low efficiency of only 20 percent, would be of clear benefit. Through additional optimizations, such as increasing cold air turbine efficiency, use of low-friction bearing systems, and variable turbine geometries (to obtain more power from the hot gas turbine during low load conditions), it should be possible to reach significantly higher turbocharger speeds. 350

/

300 250 [

200 ~

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150

u.

c..

100

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50

o

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-/

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25,000

50,000

75,000

100,000

Turbocharger Speed [rpml

Figure 3: Typical curve for exhaust gas turbocharger bearing friction

TECHNICAL IMPLEMENTATION The above shows that it is quite possible to raise the turbocharger speed by using throttling losses. A simple design was used for technical implementation, [12] based on a classic gas exhaust turbocharger. Here use of the throttling energy occurs in the radial compressor impeller. To do this, the cold air side of the turbocharger had to be redesigned (Figure 4), so a movable plunger CD was integrated that allows for adjusting the cross section of the compressor intake opening. Additional guide vanes @) allow radial entry of the air to the compressor impeller. The flow cross section of these guide vanes can be varied with a matrix d:J, which is connected to the plunger. Adaptations to the compressor housing were necessary to house these components. During the charged operating phases of the engine, the air flows axially to the compressor impeller (sketch at top right of Figure 4). The retracted plunger creates a very large cross section of flow at the compressor intake. In this operating state, the guide vanes, which are arranged concentrically around the compressor impeller, are also open. This open position was initially chosen for ease of technical implementation. In addition, the opening also acts as a stabilizing map measure in the compressor surge limit range; which expands the usable compressor operating range [3]. The transition to turbine operation is accomplished by moving the plunger. First it moves into the conical stop C?) shown in Figure 4. When the plunger has reached the stop (sketch at bottom right of Figure 4), air can no longer enter the compressor impeller axially. Now air can only flow radially via guide vanes to the compressor impeller. At the maximum opening of the guide vanes, the impeller can still function as an extreme co-swirl compressor.

59

If the plunger is moved further toward the compressor impeller, the matrix starts to slide over the guide vanes. The cold air turbine thus practically becomes a variable axial slider turbine. This sliding motion reduces the opening cross section of the guide vanes and restricts the air flow, resulting in a high flow speed (which may be close to the speed of sound) at the exit of the guide vanes. The very high co-swirl of the flow drives the compressor impeller in the direction of rotation. The compressor acts as a cold air action turbine. This design makes it possible to control the entire system using a single actuator. The quantity control of the spark ignition engine is now achieved through the expansion process in the DOT guide vanes and the subsequent energy transfer in its rotor, and not in a classical throttling process in the throttle valve. Initial simulations were conducted to test this control strategy. They show that in principle it is possible to regulate the engine using such a component, and to replace the throttle valve. It should be mentioned, however, that moving the throttling point ahead of the compressor requires special sealing measures for the bearing housing. This is necessary to prevent oil leaking from the bearing to the intake system. This seal was established by adding a by-pass component behind the compressor impeller, as well as a labyrinth seal.

Q) Moveable Plunger

Conical Stop

O>,@ Exchangeable guide vanes I Variable guide vanes with matrix

Figure 4: DOT concept to use the centrifugal compressor wheel as a cold-air turbine

EXPERIMENTAL VERIFICATION OF THE BASIC DOT CONCEPT Stationary flow test stand

For the basic functional testing of the cold-air turbine operation, a prototype was set up, using interchangeable guide vanes that represent different fixed plunger and matrix positions. The measurements were carried out on a stationary flow test stand [14]. First the conventional compressor map with fully opened plunger was measured. In a second step, a vacuum was created at the cold air side, at the compressor exit (on the P2 side), while the turbine did not receive any additional power. This set-up allowed

60

generating the cold-air turbine map. In this case, the hot gas turbine dissipates energy, in addition to bearing friction. While this operating mode allows demonstrating the basic function of the DOT concept in principle, the efficiency of the cold-air turbine cannot be accurately determined (see section "Cold-air turbine optimization through 3D flow simulation"). The compressor map can be expanded by adding the turbine operating points (Figure 5). Cold air turbine operation takes place below the 1.0 pressure ratio line. For these conditions turbine operation is illustrated as a VTG 1 turbine map according to Hiereth. [5] The lines connecting the speed points depict the flow rate behavior of the cold air turbine at a constant opening cross section of the guide vanes (similar to a constant VTG opening). Here a typical nozzle discharge characteristic becomes apparent. The measured map thereby underscores the basic functioning of the innovative DOT technique and the possibility of using a compressor wheel as a turbine.

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Figure 4: Diagram and view of an AFT. Variation of R versus the displacement of the turbine shaft and versus the turbine mass flow rate Since the angle of the blades in the VGT stator is directly related to the movement of the rack from the variable-geometry mechanism, the blades angle can be clearly determined once the rack position is established. An angle of the stator blades of 86° corresponds to the closed VGT and an angle of the stator blades of 42° corresponds to the open VGT. Measurements were carried out for seven constant positions of the VGT, and for each position several operating speeds and expansion ratios were tested. The results obtained are shown in Figure 3, where it can be observed that R increases when the VGT opens. In addition, R tends to 0.5 when the blades angle tends to be tangential to the rotor circumference (a\::;;68°). Additionally, an AFT, whose blades present a constant a\ angle (Figure 4) was tested coupled to an engine on a specific bench. This turbine consists of a moving rod to which the stator blades are attached and a vacuum pump which controls the position of the rod. The characteristics of this variable geometry turbine are shown in Table 1. In this case, for a given opening and turbine speed only one point was measured, and then R was calculated by using the average values of pressure and temperature measured on 104

engine tests. The results obtained are shown at the bottom of Figure 4. They clearly show that R is dependent on the motion of the turbine axis. It can be seen that, in contrast to what occurs in the VGT, R increases when the turbine closes. The cause of this behaviour can be the particular way in which the AFT reduces the stator effective section, so that it can be presumed that a lamination process is produced in the stator when the AFT is closed (AFT Disp = 0) and therefore R is increased. There is also a dependency with other turbine variables; for example, R generally decreases when the exhaust gas flow rate increases. The results plotted in Figure 3 show this trend for each position and each turbine speed, except in the case of VGT closed, where R = 0.5 has been imposed as hypothesis. It is interesting to notice that when the AFT is closed, R remains also virtually constant, since the flow rate in the AFT is almost constant for the different expansion ratios considered, (Figure 4). In Figure 4, it can be observed that R varies between 0.1 and 0.25 when the AFT is open (AFT Disp = 10). Therefore, the increament of relative speed should be reduced in the rotor and this would account for the very low efficiency that is usually observed for a completely open AFT. Calculation of pressure at the stator outlet

Once R has been obtained as a function of the corrected turbine variables and of some easily measurable geometric parameters, it is necessary to establish its relation with the intermediate pressure between turbine stator and rotor in order to apply the chosen model of two nozzles plus an intermediate volume. If R is defined according to equation A.l, and considering the gas to be a perfect gas, one may write:

J

Eq.17 1; = (1- R) . (T2 + R 1'a 1'a If the type of evolution occurring in the turbine was known, it would be possible, with the help of Eq 18, to relate the intermediate pressure to the temperature, and therefore, to R. Taking into account that in any thermodynamic evolution, a flow which traverses the turbine may start from certain initial conditions (Po, To) (the terminology established in figure 1 is still maintained) and may achieve certain final conditions (P2, T2), in such a way that PO>P2 and To> Tz. the polytropic exponent of the evolution can take values between 1, for the extreme case in which To=T2, and ~1.33, in the case of exhaust gases and isentropic evolution. Assuming that the process undergone by the gas in the turbine is adiabatic but irreversible, it is physically impossible for the polytropic exponent to take values below 1, since that would imply a temperature increase. Likewise, any value above 1.33 would imply a decrease of entropy in the final state when compared with the initial state. If the evolution experienced by the gas traversing the turbine was a polytropic evolution, with a constant polytropic index (n), it could be represented by:

P2 (T2 Po = 1'a

)%-1

n In(P2/ Po) => n-l = In(I;/1'a)

Eq.18

However, when the turbine is quite closed, the evolution of the gas inside the stator is much more irreversible (since when the passage area diminishes, the losses caused by friction increase) than the evolution inside the rotor. Therefore, it seems convenient to assume the hypothesis that the evolution across the stator and the evolution across the

105

rotor have different polytropic exponents. Thus, Equation 19 must be fulfilled for each evolution, assuming that i is constant throughout the process in the stator or in the rotor: Eq. 19 and for the total evolution from the turbine inlet to the turbine outlet, it can be established that: P2 Po

= P2

. PI PI Po

=

(7; J%-I .(7; JKI = (7; J%-I 7;

1'0

Eq.20

1'0

Since it is not possible to measure pressures and temperatures at point 1, which is located between the turbine stator and rotor, the polytropic exponents k and g cannot be directly obtained. However, it is possible to relate them to n, as shown in equation 21. k

g

( ~_~).lnT2 n-l g-l To

k-1

g-1

In[(1-R)~ +R]

-=-+

Eq.21

Now the problem is that there is only one equation, Eq. 21, with two unknown values, k and g. Therefore, it is necessary to consider an additional hypothesis. To do this, the behaviour of the VGT and the AFT under extreme open and closed positions must be analysed. Closed variable geometry turbine In the case of a VGT, the hypothesis that R is 0.5 when the turbine is closed has been used. In this case, the thermodynamic evolution should be similar to the k-g evolution shown in the left part of Figure 5, due to the fact that the high velocity of the flow at the stator will increase considerably the increment of entropy. If the objective is to calculate the pressure between stator and rotor (P1k-g in Figure 5), it can be confirmed that a satisfactory hypothesis would be to assume the evolution in the rotor to be isentropic (y) and to calculate the polytropic exponent of the stator k' as a function of n and y.

----- y ------ k'-r - - n --k-Q

h

Po

h

~n

8h,

8h.

P.k-g

8h

&1

!=In

Qy !=Ik·.., 8h

I ----------------I. ------2s1

s

s

Figure 5: h-s diagram for a closed VGT (left) and an AFT (right) 106

Thus, p1k'-y should be taken as the sought pressure value, which is much more appropriate thanplyor pIn, obtained in the case where the entire evolution was assumed to be isentropic or with a constant polytropic exponent, respectively, In addition, it has been verified that the more irreversible the evolution in the stator, the more accurate this assumption will be. That is, when the VGT is closed (Figure 5), and the passage area of the stator is very small, the irreversibilities caused by friction are very significant and the entropy increment must be higher in the stator than in the rotor. In the case of a variable geometry turbine of the AFT type, when it is closed and the actual passage area of the stator is very small, R increases in accordance with what was experimentally obtained in Figure 4. If, in tum, the hypothesis of flow lamination is assumed, it should cause a very high increase of entropy across the stator due to the increased flow friction. Therefore, the evolution found will be similar to that labelled in Figure 5 as k-g evolution. Once again, when considering that the final objective is to calculate the intermediate pressure p1k-g, in the right side of Figure 5, it can be observed that the differences between p h-g and p lk'_y are the least significant ones. Open variable geometry turbine In the case of a VGT, when it is open, R is higher than 0.5 (Figure 3), which implies a low level of expansion at the stator. In this case that the turbine directs the flow towards

the rotor, the thermodynamic evolution should be as observed in the left part of Figure 6 with little increase of entropy at the stator and, therefore, the differences in PI calculated using any of the hypotheses are very small, and the real solution is always between Plk'-y andply. In turn, in the right part of Figure 6 a diagram is presented to illustrate what occurs under those hypotheses which are stated at the beginning of the section, for the case of a variable geometry turbine of the AFT type when open. Here, as previously seen, (Figure 4), R is low (~0.20). It can be observed in Figure 6 that once R is fixed and a small increase in entropy occurs, due to the low velocity of the flow at the stator, the differences between any of the proposed evolutions and p1k-g are small and, in any case, the required pressure lies between Ply and Plk'-y' ------ '"1

--n

h

------ k'-y

--k-g

Po

h

Figure 6: h-s diagram for an open VGT (left) and an AFT (right)

107

Sumarizing, in order to simplify the model and since the calculation ofPI is the main objective, it will be assumed for all the cases that the evolution in the rotor is isentropic, due to the fact that for the case of open turbine the differences are not significant for the calculation ofPI. For the evolution at the stator, a polytropic exponent k given by Eq. 21 where g=rand n is calculated using Eq. 18 will be assumed. Thus, the calculation of the intermediate pressure may be rexpressed as:

lX-I lR

n-I/n

12= [ (l-R)· P2

Po

(PJ

+R

=

_

0.5 if a>al;m;t 2·tanal 9={ rh*

R-l

2 7f

'-;'f(-,m

2

P2'*

N

DIDz

Poo

. ,J4,TJT,)lfa::;al;m;t

Eq.22

Obtaining the effective areas of the nozzles equivalent to the turbine stator and rotor Once R is established and considering how it is related to the pressure drops across the stator and the rotor, the calculation of the effective areas of the nozzles equivalent to the stator and the rotor can be completed by using the nozzle equation (Eq. 1). For the stator, one has:

A

eff _ st

=mr · ~ .~.(Poo) POO r PI

(r-rr

Yr

-1

r. ~[1-(J!L) r1 r 1 [

]

Eq.23

PO~

and for the rotor

Eq.24

The calculation of these effective areas has been carried out for three different turbines, a VGT, an AFT turbine and a fixed geometry turbine (FGT) whose characteristics are shown in Table 2. The FGT performance maps were supplied by the manufacturer. An obvious dependency on the effective area of the nozzle equivalent to the stator with the position of the turbine can be observed for the VGT and the AFT in Figure 7. AFT

VGT

'E600 5500 ~'" 400 Q)

~ 300 ~ Q)

~

00

~

o~~~~~~____~~~~~~~ 6 10 o AFT Displacement (mm)

~

200 100 0

""L,-.-~~~~~~~~~~~~ 40

50

60

70

60

90

Stator blades angle (")

Figure 7: Correlations obtained for the effective areas ofthe nozzles equivalent to the AFT and VGT stators

108

VGT

.-E

600

.s 600 m til 400

~

t&

"'0

400


0.5

Q)

.~

o

0.4

U) U)

~ 0.3 E o

o

0.2 +---,---t---..--t--..-j--..--t--.-'-t----r"-t----;;:-j--..-'--l 0.00 0.02 0.04 0.06 0.08 0.10 0.12 0.14 0.16 Compressor reduced mass flow rate

rileed.

c [kg/sf

Figure 7 Error analysis of the measured isentropic efficiency of the GTl749V 55 Trim compressor efficiency is lower than the compressor adiabatic efficiency due to the amount of heat transfer from the turbine to the compressor. The deterioration of the compressor non-adiabatic efficiency is very significant at low rotational speed (low compressor peripheral Mach number). The effect of heat transfer from the turbine to the compressor decreases with increasing the compressor rotational speed (increasing the compressor peripheral Mach number). Therefore, the deterioration of the compressor non-adiabatic efficiency decreases with increasing the turbocharger rotational speed. The measured compressor non-adiabatic efficiency at 140000 rpm and 160000 rpm is almost independent of the exhaust gas temperature at the turbine entry. This behavior coincides very well with the results of the parameter analysis of the turbocharger non-adiabatic performancc presented above. The deviation of the compressor non-adiabatic efficiency from the compressor adiabatic efficiency is higher near the compressor surge linc. This bchavior is due to thc increase in the specific amount of heat transfer to the compressor at low mass flow rates. This increase causes the higher deterioration of the compressor non-adiabatic efficiency at low flow ratc. Note that the behavior of the comprcssor non-adiabatic efficiency with the exhaust gas temperature at the turbine entry is non-intuitive. It was expected that the amount of heat transfer to the comprcssor increases with increasing the exhaust gas temperature at the turbine entry. However, Figure 8 shows that thc mcasurcd compressor non-adiabatic efficiency is not directly dependent on the exhaust gas temperature at the turbine entry. The highest deterioration of the compressor non-adiabatic efficiency for the 60000 rpm compressor performance line takes place at T6t = 873 K. For thc 100000 rpm compressor performance line, this occurs at T6t = 973 K. This behavior implies that the exhaust gas temperature at the turbine entry is not the only parameter affecting the amount of hcat transfcr to the compressor. The reason is that heat is first transferred by conduction from the turbine casing to the bearing housing. Thc lubrication oil removes part of this conduction heat transfer and the rest flows to the compressor. Thus, the lubrication oil also works as a cooling fluid for the bearing housing. It has therefore a significant effect on the amount of heat transfer to the compressor and hence on the compressor heat number. Figure 7 and Figure 8 show that thc mcasured deterioration of the compressor efficiency due to the amount of heat transfer to the compressor during the hot measurements is higher than the maximum probable error in the measured compressor efficiency. This is also tme for low rotational spccd (60000 rpm) despite the maximum probable error being quite high at this rotational speed, Figure 7.

128

T ,,:

--0-

305 K _ _ B06 K -----A- B73 K

O.B..------'i"'c'c-_-~--

C,)

~

.g

~~ c ~

g~ ~

iij~~I

0.6

0.5

C,)

a c

~.~ ~::E

0.4

~ -ot-T-+-.,......j 0,00 0,02 0,04 0,06 0,08 0,10 0.12 0,14

IDred.C

Figure 4 Comparison between measured and calculated performance map of the GT1749V compressor .'] :

nred,C [lIlIll ,

• 80000 0 100000 SyniJols: m:asured Lines: calculated

•

120000

2.4 2,2

,

---;---~&-,

2,0

~ ~

0:: --...

.r

,

~

1.8

~

u.

1.6

.!

1.4 1.2 1.0 +-r-+-,-+--r--+-T-+-i-ot-T-+-.,......j 0,00 0,02 0,04 0,06 0,08 0.10 0.12 0,14

0

140000 I.

160000

0,8 0,7 0,6 0.5 0.4 0,3 0.2 0.1 0,0 0,00 0,02 0,04 0,06 0,08 0.10 0.12 0.14

Figure 5 Extrapolated performance map of the GT1749V compressor

139

•

•

o 2942

4251 2.1...-------==------, A

~

6.

~ 1.8 ~ 1.5

3594 5555 0.8...--------------., I I I I I I I 0.7 __ ~ __ ~ __ ~ __ ~--~--~--~-V1Xf. 0.6 _lOO'loopen__ '_ 60% open

,

~

.!

0.5 0.4

~

OJ

...!..,

~ 1.2

gp

o

0.9

1 ,s OJ

~~ 0.2

0.6

0.1 0.0 ~~+T+T+T+'r-.Ij-Io,-+-r'-f'ltII.-II-T-I 0.00.1 0.2 OJ 0.4 0.5 0.6 0.7 0.8 0.9 1.0

0.0 +-...,......-r-"T"""-r-"T"""-r-.,.-+-.,........, 1.0 1.5 2.0 2.5 3.0 3.5

pJps [-]

dc, [-]

Figure 6 Comparison between measured and calculated performance map of the GT1749V turbine A very good agreement between the measured and the calculated performance of the GT1749V turbine is shown in Figure 6. It can also be noted in Figure 6 that the physically based regression equations take also the effect of centrifugal force on the turbine performance map at different turbine guide vanes position.

ENGINE PERFORMANCE SIMULATION Engine Simulation Programs One of the main goals of the development of turbocharged engines for passenger cars is to optimize turbocharger response during acceleration from part load. The investigations typically start at very low turbocharger speeds, corresponding to low engine speed and torque. It is very important to have precise knowledge of the efficiencies of the turbine and the compressor in the lower speed range for a precise definition of these starting conditions in the calculation model of the engine. Furthermore, separating the effect of the heat transfer from the turbine to the compressor from the aerodynamic work gives more realistic results when calculating turbocharger dynamics which is vital for the prediction of car acceleration. The TC2005 Program has been implemented in two different engine simulation program models. The first program model is the zero-dimensional model THEMOS®. This model is based on the quasi-stationary filling and emptying method developed at the Technical University of Berlin, Institute of Internal Combustion Engines. The second engine simulation program model is the one-dimensional model PROMO which is based on unsteady, compressible pipe flow (Linnhoff, 1985). To be able to represent the behaviour of turbo machines in unsteady flow, efficiency and mass flow rate must be calculated for each iteration step of PROMO. Therefore accuracy and calculation speed are important demands on the software. These demands are met by the evaluation of the weighted regression equations. The user now can take advantage of a safer reproduction of extrapolated map data, a much more accurate distinction between similar compressors or turbines, and a much improved handling of the input data. 140

Both engine simulation programs were used to simulate engine operating points located inside as well as outside the measured turbocharger performance map to check the effect of the physically based turbocharger simulation model described above in comparison with the classical mathematically based turbocharger performance simulation. Results of Engine Simulation Identical measured turbocharger maps were used as input data for the TC2005 Program as well as for both engine process simulation programs. Running the TC2005 program under the assumption of adiabatic turbocharger behaviour generates a reference calculation. Both engine simulation programs were used with and without the TC2005 Program. Without the TC2005 Program means that the engine simulation program uses mathematical methods to simulate the turbocharger performance. A comparison between reference calculations with and without TC2005 Program shows good agreement for all tested full load and part load operation points (n = 4000 min- 1 / full load and BMEP = 2 bar, n = 2000 min- 1 / full load). This good agreement is due to the fact that the turbocharger operation points in these cases are either located inside the measured turbocharger maps or only a small extrapolation is required for the turbocharger performance simulation. These results were expected because of the small differences due to reproducibility of turbocharger performance maps for both calculation methods and the good agreements for the TC2005 Program already shown above. The operation point n= 2000 min- 1 / BMEP = 2 bar (Figure 6) corresponds to a very low turbocharger rotational speed and shows differences up to 10% because of the relatively high and therefore unacceptable systematic measuring error. Furthermore the influence of the temperature at the turbine inlet on the compressor efficiency has an increasing effect for this engine operation point. The turbocharger map data are based on a turbine inlet temperature of T6t = 873 K, whereas the corresponding measured temperature is 558 K at the engine under consideration, so that the compressor efficiencies are taken from the map are unrealistically small. Based on the experiences from experimental investigations on the GT1749V turbocharger and on the engine operation at steady-state and dynamic conditions it seems to be sufficient to verify the TC2005 Program with its non-adiabatic model at the engine operation point n = 2000 min- 1 / BMEP = 2 bar. Figure 6 includes a comparison between reference and non-adiabatic calculation model for both engine process simulation programs. This comparison shows that the estimation of the temperatures before and after the turbine as well as the temperature at the compressor outlet is clearly improved with the consideration of the turbocharger non-adiabatic performance. With the assumption that the turbocharger was adiabatic during measurements, the maximum percentage difference of the estimated temperatures can be as high as 9%. This high percentage difference is due to the consideration of the heat transfer inside the turbocharger as useful turbine and compressor work. The maximum difference in the estimated temperatures clearly decreases with the non-adiabatic calculation and reaches a maximum value of only 2% (right ordinate). The improvement in the estimated temperatures is achieved by THEMOS® as well as PROMO. The consideration of the turbocharger non-adiabatic performance also leads to a clear improvement of the estimated pressure at the compressor outlet and the turbine 141

inlet. The maximum difference in the calculated pressure at the turbine inlet with the adiabatic calculations is about 4.5%. This maximum difference is reduced to less than 1% (right ordinate) by the consideration of the turbocharger as non-adiabatic. In summary, the correlations which were described theoretically before starting this investigation are reproduced correctly. The compressor power from the non-adiabatic calculation is lower than that from the reference calculation. The turbine power also exhibits a correct behaviour. As a consequence of the different turbocharger power balance, the turbocharger speed and the turbine inlet guide vane position are different for both calculation methods. The operating point presented in Figure 6 was chosen to examine the ability of the presented physically based modeling of turbocharger performance to extrapolate the turbocharger performance maps down to very low rotational speeds where the effect of heat transfer inside the turbocharger is significant. The improved estimated performance with the application of the TC2005 Program proves the ability of the models described in the present paper to simulate the turbocharger performance correctly and on physical basis. However, other operating points will be simulated in future work to further explore the limits of the models described in the present paper 10

-1

. 1"1 ~

T ~

-6

~

n

~

IDI2ZITHEMOS®r PROMO

~

1m I ~ t i ~ ; ~

~ ~ 1- 12ZI ~~ ~ 1 ~ ~~ 'I 'I ~~ ~u ~ ~ I~ ~~ til ~ Ii u ~ ~ ~ ~ ~~ ~ c ~ ~ ~~ ~ ~ e ~~ J::': ~~

n = 2000 min·1 I BMEP= 2 bar I 30%EGR

J::'"

-

r--

i

Figure 6 Comparison between measured performance and the performance predicted with THEMOS® and PROMO at n = 2000 min-I, BMEP = 2bar (Berndt and Pucher, 2005) left ordinate: deviations of the adiabatic reference calculation (= "basis") Note the larger scale. right ordinate: deviations of the non-adiabatic calculation

142

CONCLUSIONS

A method for the simulation of turbocharger performance maps is presented. This method is based on the measured turbocharger performance maps. It aims at extending the simulation of the turbocharger performance to very low rotational speeds with a minimum amount of data regarding the turbocharger geometry. The presented method uses physically based regression equations to simulate the turbocharger performance. These regression equations weight the operating points close to the computed operating condition more highly to achieve higher accuracy in the simulation of the turbocharger performance. A turbocharger performance simulation program is developed on the basis of the method presented. This program can operate either as a stand-alone program or as subroutines in engine simulation programs. Comparison with measured turbocharger performance maps shows a good agreement between calculated and measured turbocharger performance. This also applies for the extrapolation towards low rotational speeds. The subroutines of the TC200S program were implemented in two different engine simulation programs. The verification of the turbocharger performance simulation program included in engine process simulation programs confirms the expected behavior. REFERENCES

[1] Aungier, R.H. (199S): Mean Streamline Aerodynamic Performance Analysis of Centrifugal Compressors. ASME Journal of Turbo machinery, Vol. 117, pp. 360-366. [2] Aungier, R.H. (2000): Centrifugal Compressors; a Strategy for Aerodynamic Design and Analysis. ASME PRESS, New York, ISBN 0 791800938. [3] Baines, N.C. (1998): A Meanline Prediction Method for Radial Turbine Efficiency. IMechE 6th International Conference on Turbocharging and Air Management Systems, Paper No. CSS4/006/98, London, UK. [4] Berndt, R.; Pucher, H. (200S): Einfluss eines diabaten Turboladermodells auf die Gesamtprozess-Simulation abgasturboaufgeladener PKW-Dieselmotoren. 1st Conference Engine Process Simulation and Supercharging, June 30 - July 1, Berlin [S] Chen, H.; Winterbone, D.E. (1990): A Method to Predict Performance of Vaneless Radial Turbines under Steady and Unsteady Flow Conditions. IMechE 4th International Conference on Turbocharging and Turbochargers, Paper No. C40S/008, May 22-24, London, UK. [6] Kolmanovsky, 1.; Stefanopoulou, A.G. (2000): Evaluation of Turbocharger Power Assist System Using Optimal Control Techniques. Society of Automotive Engineers SAE, PaperNo. 2000-01-0S19. [7] Linnhoff, H.-J. (198S): Die Berechnung des Ladungswechsels und Ansprechverhaltens von Verbrennungsmotoren mit Abgasturboaufladung. Diss. RuhrUniversitiit Bochum, ISBN 3-89194-0S3-X [8] Mueller, M.; Hendricks, E.; Sorenson, S. C. (1998): Mean Value Modeling of Turbocharged Spark Ignition Engines. International Congress Exposition, Paper No. 980784,pp.12S-14S. [9] Pucher, H.; Berndt, R.; Grigoriadis, P.; Nickel, J.; Hagelstein, D.; Abdelhamid, S.; Seume, J. (2003): Erweiterte Darstellung und Extrapolation von TurboladerKennfeldern als Randbedingung der Motorprozesssimulation. Final Report for 143

Forschungsvereinigung Verbrennungskraftmaschinen FVV, FVV Project number 754, Heft 774. [10] Rodgers, C. (1984): Static Pressure Recovery Characteristics of some Radial Vaneless Diffusers. Canadian Aeronautics and Space Journal, Vol. 30, No.1, pp. 42-54. [11]Zinner, K. (1961) Diagramm zur Bestimmung des Betriebspunktes einstufiger Abgasturbolader. M.A.N. Forschungsheft 10. © Institute of Turbomachinery and Fluid Dynamics

144

Development of Electrically Assisted Turbocharger for Diesel Engine y YAMASHITA, S IBARAKI, and H OGITA Mitsubishi Heavy Industries, Ltd, Tokyo, Japan

We are developing a Hybrid-turbocharger. At first, we intended to confirm a elimination of turbo-lag and an increase of the engine torque. All components are our original design. This paper describes the design and experimental results. The ultra high-speed motor and inverter, which are key components, are realized by strong permanent magnet and unique circuit. The results indicated that at engine speed IOOOr/min and 1200r/min the engine torque could be increased approximately 17%. INTRODUCTION Background

An electrically assisted turbocharger (which we call "Hybrid-turbocharger") has been attracting attention in the automotive industry. The new-type boosting system has a possibility to respond demands for automotives from society such as corresponding a regulation of exhaust gas amount and fuel consumption, and improving of drivability in particular a elimination of turbo-lag. System composition As shown in Figure 1, a Hybrid-turbocharger is composed by components of conventional turbocharger and a ultra high-speed motor which drives a compressor electrically. The motor is located in a turbocharger. The inverter which drives the motor is located outside and connected to three windings of the motor. Also, the inverter is connected to a battery by two wires. The motor can control the torque independently of the turbine torque. Additionally the motor torque direction can be controlled not only positive to rotation direction but also negative direction. Therefore, it is possible to assist the turbine torque at low speed and to brake at high speed. It indicates that at high speed the surplus energy of exhaust gas which is ordinary discarded through a waste gate can be retrieved as electric energy. inverter

battery

intercooler

Figure 1 A schematic diagram of Hybrid turbocharger

147

Estimation of effects The improvement targets we have intended by the Hybrid-turbocharger are below. •

Improving a fuel consumption by 10% at low speed.

•

Increasing an engine torque by 50% at low speed.

•

Improving a turbo lag by 70% at an initial acceleration.

To determine how large output of the motor was needed, we conducted an engine simulation by an in-house code. The engine specification in the simulation is a turbocharged 2.Oliter diesel engine with an intercooler. The diameter of the turbine impeller and the compressor impeller are respectively 43mm and 44mm. The turbine efficiency and motor one were set to respectively 70% and 90%. The compressor efficiency was based on experimental results. The simulation results are shown in Figure 2. The steady simulation results indicate that a lkW motor input will allow the increase of the engine torque by 50%, and a 2kW will increase to double. As to the engine fuel consumption, a lkW motor input will improve by 8%; a 2kW will improve by 12%. The transient simulation results indicate that using a 1.3kW motor will shorten the turbo-lag by approximately 50%, and using a 2.0kW motor will shorten by 70%. Based on above results, the proto type motor power has determined 2kW. In following chapter, detail specifications of each component are illustrated .

~

25

is

15

·1

10

.

§

____ +3m assiltEd

-*-- +2m

assiltEd --..... +1m assiltEd - -

20

0

500

1000

120

1500

~. 105

------------- - ---

ti

3000

------------ :=:: == - -

-------------~--

OJ

2500

____ +3m assEtEd

J1''' nl ~

2000

110

+12%

+8%

-----------------------

-

100

0

500

1000

1500

2000

2500

3000

Eng:ile speed [r/m:h]

350 300 ~ 250 200 B 150

~

.J

100 50 0 2

3

assisted

4

5

6

7

tine [;sec]

Figure 2 The engine simulation results with the Hybrid turbocharger 148

DESIGN Turbocharger and Compressor

The diameters of the turbine impeller and the compressor impeller are respectively 43mm and 44mm as same as the simulation. The turbine impeller is diagonal flow type, and the compressor impeller has a backward curve. The magnet rotor of the motor is located between the turbine impeller and the compressor impeller. There are two bearings between the turbine impeller and the magnet rotor. As a result, the rotor is supported as an overhang configuration. It means the rotor dynamics must be considered carefully to prevent a fatal vibration. The vibration analysis results have indicated that it is sure that under operating speed, 220 000 r/min, there are two critical speeds, but the effective damping of full floating bearing can prevent from a growth of vibration. Power electronics Ultra high speed motor

As an ultra high-speed motor, PMSM (permanent magnet synchronous motor) was adopted. The magnets used for a source of magnetic fields have very high performance. The material is Nd-Fe-B (Neodymium-Iron-Boron) magnet, and the remanent flux density Br is approximately 1.26 tesla. The strong magnet can make the motor size smaller than the other type of motors, such as a DC motor, an induction motor, a reluctance motor. Another approach for downsizing is a concentration winding method for coils that induce rotation magnetic fields. In addition the motor has a large gap between the magnet surface and the inner surface of the stator teeth. Since an inductance of motor windings is almost inverse proportional to gap length and react as a resistance for a rapid current change, a small gap leads to the quick response of a motor torque. In an ultra high-speed motor, it is very important to retain the rotor against its own centrifugal force. In our motor, the rotor is wound by a carbon fiber. As a negative side effect of downsizing, it is difficult to cool down a motor against its self-heating due to power loss. The excessive temperature rise causes a deterioration of insulation performance and irreversible demagnetization of magnets. In an ultra highspeed motor, the heat generation of magnet can not be negligible because the high frequent fluctuation of the magnet flux induces a eddy current loss in the magnet. To reduce the fluctuation of magnet flux on magnet surface, the stator has six teeth. Loss analysis of the motor was conducted by using a finite element method (FEM) and a boundary element method (BEM). For enough cooling, the stator and rotor of the motor are cooled by forced oil cooling and by forced air cooling respectively. By this cooling, the maximum temperature rises of rotor and stator are held under 140 degC and 155degC respectively.

149

Figure 3 Overview of the Hybrid turbocharger and motor stator

Inverter

In general, a voltage-source PWM inverter is widely used to operate a PMSM, where a vector control scheme is usually employed. However, the sinusoidal motor current regulation is hardly possible in the case of ultra high-speed drives because the fundamental frequency of the motor current is several kilo-hertz even at the rated speed, which does not allow the inverter to perform the pulse width modulation appropriately. Therefore, we adopted a pseudo-current source inverter for ultra high-speed drive, which is easier to control the current value and to synchronize the current phase to rotor phase. This inverter consists of two parts, i.e., a current-controlled buck-boost chopper and a 120 (electric) deg conduction 6-step inverter. The former circuit employs to control the motor current according to the current demand. In other words the circuit regulates the current by means of a pulse amplitude modulation (PAM) technique which adjusts the chopper duty ratio proper. The switching frequency of chopper is 30kHz. When the motor is in a motoring mode, the chopper acts as a buck chopper. On the contrary, when the motor is in a generating mode, the chopper acts as a boost chopper. The latter circuit employs to synchronize the current phase to the rotor phase which is detected by mechanical sensorless method. In permanent magnet motor, the rotor position can be detected by induced voltage of coil and additional analogue detect circuits without position sensor. Strictly speaking, since it is difficult to detect the induced voltage at low speed, the phase control is conducted by a open loop control at starting.

150

C

Vdc:DC power source, C: electrolytic capasitor, L: series reactor, D:Diode, PM:Permanent magnet Motor, Sc\, Sc2, S\, S2, S3, S4, S5, S6: FETs

Figure 4 The inverter circuit diagram and overview

EXPERIMENTAL RESULTS AND DISCUSSION Stand-alone experiment

Prior to the experiment with an engine, the stand-alone experiment has been conducted to verifY the basic performance. The turbine was given a torque and controlled its rotation speed by the source of air in laboratory. The voltage source of inverter which drove motor was not a battery but rectified DC voltage by 3-phase rectifier and capacitors. The reason was that the proto type Hybrid-turbocharger was designed as nv specification, and it was expected to operate for a certain amount of time without worrying about a limit of battery storage capacity. Table 2 shows a difference of rotation speed and pressure ratio between the normal turbocharger (without motor assist) and the Hybrid-turbocharger (with motor assist). Under this stand-alone experiment, the Hybrid-turbocharger's torque was controlled constant at about 0.l6 Nm which was equal to a power of2kW at 120000 r/min. Figure 5 shows comparisons of boost pressure and rotation speed between the normal turbocharger and the Hybrid-turbocharger. Focusing an attention on starting point where the effect of motor assist is significant, the time to the rotation speed 74 000 rlmin was cut to 33% and the time to boost pressure 26 kPa was cut to 41%. A few seconds after the start, the turbine torque was more dominant than motor torque. Therefore large differences of increasing rate between normal turbocharger and Hybrid-turbocharger were not seen.

151

Table 2 Comparisons of speed and pressure ratio between with and without motor assist Before assist 40000 81000 100000 122000 Speed[r/min] After assist 85000 109000 121 000 138000 Pressure Before assist 1.05 1.21 1.50 l.33 ratio After assist 1.23 1.41 1.51 1.66

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152

12

Experiment on engine The effect of the Hybrid-turbocharger for the engine performance has been verified on the engine test bench in our laboratory. The engine displacement is 1.7liter turbocharged with intercooler. The rated power is 75kW at 4600r/min, and rated torque is 240Nm at 2400r/min. In a default configuration, the turbocharger is VG (Variable Geometry) turbocharger whose diameters of the turbine and compressor are 40mm and 44mm respectively. When the motor speed was under 120 OOOr/min of the rated speed, the torque was controlled at O.l6Nm in constant, and above the rated speed the power was controlled at 2kW in constant. Figure 6 shows a difference of steady state engine torque between using the VG turbocharger and using the Hybrid-turbocharger. At engine speeds 1000r/min and 1200r/min, the engine torque was increased approximately 17%. This improvement is assumed to be caused mainly by an increase of air-fuel ratio. Figure 7 shows the difference of transient turbocharger speed and boost pressure between using the VG turbocharger and using the Hybrid-turbocharger. The start condition ofthe engine in this experiment was high idling, and suddenly engine was full loaded. In the Hybridturbocharger, the motor torque was generated at the same time as the engine load. The increasing rate of the Hybrid-turbocharger was larger than the VG turbocharger. Therefore, at least the response of the engine torque by Hybrid-turbocharger would be equal to the one by VG.

Assist motor output: 2kW

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______----.

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Maximum stress 318MPa at rotor bore

-

Fig. 1: Mechanical stresses in motor rotor at 130 000 rpm. Improvements to design

During the course of the project, three different designs of induction motor were tested. The Mkl.O has a 056 x 38 mm stator, Iron Silicon laminations and an industry standard potting compound. In order to improve the thermal situation within the same dimensions, a Mk1.l motor was created with Iron Cobalt laminations and potting compound with a temperature capability up to 200°C. The Mk2 motor also has the Iron Cobalt laminations but the dimensions are larger and the potting compound has a limit of 220°C. TURBOCHARGER DESIGN Basic design

The turbocharger design had to incorporate the following features:1. Fitting a motor rotor onto a turbocharger shaft that gave acceptable shaft motion. 2. Location and assembly of bearings. 3. Accommodate the motor in a bearing housing that provided suitable cooling. The turbocharger design features a motor rotor heat shrunk onto a conventional shaft and wheel, with the turbine end bearing preassembled into a bearing housing end cap. The stator is pressed into the bearing housing which has an internal cooling passage. Conventional floating hydrodynamic bearings were specified. This allows operation at high speed and the rotor benefits from the cooling effect ofthe oil flow. Mkl turbocharger design

The first design had a 056 x 38 mm motor that would fit within the size of a conventional HY 40V turbocharger. This had spatial restrictions that meant several compromises were made, most notably that the cooling passage was within the bearing housing and not in direct contact with the stator. This was smaller than ideally required for a motor of 6kW output. After the first test work was done it became clear that a larger motor and better cooling would be required.

160

Mk2 ELEGT Turbocharger • Motor 090 x 65 mm • Water cooling (wet stator) • 78mm bearing span

Fig. 2: Cross section through the Mia ELEGT turbocharger Mia turbocharger design

A second turbocharger was designed that placed motor requirements ahead of the turbocharger design. The stator was larger at 090 x 65 mm, requiring a longer shaft which may have caused an increase in shaft motion. The larger diameter also meant that the turbine volute was pushed away from the turbine entry and inlet vanes had to be used to maintain turbine efficiency. Shaft motion

The Mkl design of motor rotor was modelled along with a standard HY 40V shaft and wheel in Holset's rotor dynamics software. The results showed that natural frequencies were lowered and problems may be encountered at the maximum speed of 130 000 rpm but that improvements may be seen at lower speeds. A prototype turbocharger was tested with a dummy mass of the correct size and density and the test results were more encouraging, showing that the shrunk on motor rotor stiffened the shaft and reduced shaft flexing at high speeds. The end result was that the shaft motion was reduced and the oil film thickness increased relative to the standard HY 40V rotor system. For the Mk2 turbocharger, the longer shaft was expected to lower the frequency and weaken the shaft further, but the simulations show that the shrunk fit motor rotor is beneficial. The stiffening benefit allows the shaft to be increased in length by 20mm before shaft motion at the turbine wheel would become too large. -,--'

Fig. 3: Shaft motion predictions for the 20mm longer shaft. 161

Thermal considerations

The motor stator has to be kept below 220°C in order to avoid damage to the potting material and the winding insulation. Turbine temperatures are typically around 700°C and so water cooling was considered to be necessary. A full conjugate heat transfer calculation was performed on three design iterations using the CFX-5 code. The designs were: 1) Water jacket within the bearing housing. 2) Water jacket integrated into the stator. 3) Water jacket integrated into the stator with spiral guide vanes. Design 1 with water jacket within the bearing housing resulted in a hotspot where the oil drain cavity was located. Design 2 with the simple water jacket around the stator gave more even cooling than Design 3 with the spiral vanes. Whilst more time could have been spent optimising the vane design, the simpler design was expected to give sufficient cooling and so the simple design was chosen as it also gave more even cooling across the stator. An even temperature across the motor windings was considered important to ensure optimum motor performance and durability. Consideration was given to hot shut downs and it was found that the coolant must be circulated after the turbocharger is stopped in order to avoid the high metal temperatures in the turbine housing soaking through to the stator and causing damage. Assembly considerations

The assembly of the ELEGT turbocharger differs from a conventional design because the turbine end bearing is located in an end cap to allow the motor rotor to be heat shrunk onto the shaft. The end cap must be a press fit into the rest of the bearing housing to ensure accurate concentricity with the compressor end bearing. Oil must be supplied to the end cap for the turbine end bearing and sealed to avoid leakage to the turbine or atmosphere. In order to reduce the amount of oil splashing onto the motor parts, baffles are installed around the bearings to guide oil towards the oil drain and away from the rotor to stator gap. TURBOCHARGER TESTING

The ELEGT turbochargers were tested on a gas stand test rig at Holset in order to establish the performance of the motor, in particular, the length of time the motor could be operated at before reaching its maximum allowed temperature. Motor power only

From stationary, the motor was used to power the turbo up to a target speed of 30 000 rpm. The stator temperature was then monitored until a temperature rise of less than 1°C per minute was observed. Fig. 4 shows the comparison between the different motor designs. The Mkl motor with the first power controller gave a temperature of 150°C within 5 minutes. The Mkl.l motor with the second power controller was better at 135°C but the Mk2 motor gave a significant improvement running at 82°C in 26°C ambient conditions.

162

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Fig. 4: Motor stator temperatures whilst running at a target speed of 30 000 rpm powered by the electric motor only. This is a shaft power of around lkW. Altering turbo speed

Whilst speed increases were shown on the test rig, it was not possible to replicate the feedback loop that would occur on an engine, thus the true benefit of electric assist must be evaluated on the engine. The motor was tested on the gas stand up to the speed of 100 000 rpm to verify that the motor could operate at such conditions. A change in shaft power of2kW was achieved for more than 2 minutes. A reduction in shaft power of 2 kW was achieved at 100 000 rpm also for 2 minutes. A graph of the temperature rise is shown in Fig.5. Although unintentional at the time, a rapid braking event reducing the speed to 86,000rpm was achieved over a 3 second period. This represents a change in shaft power of 12kW and is shown towards the end of the data captured in Fig.5. Despite being significantly more than the rated power for the motor, this event did not damage the motor in any way. 200

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163

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ENGINE TESTING The Mkl.1 ELEGT turbocharger was mounted on a standard production CURSOR 8 engine (7.8Iitre, 228kW), installed in a test cell at Iveco Motorenforschung AG in Arbon. The initial days of operation were to verify that the ELEGT is capable of: 1. Improving the transient response of a fixed geometry turbocharger, and to 2. Limit efficiently turbo speed, by generating electric energy. Fig. 6 is a schematic that shows the arrangement of hardware and information exchange. The electric power supply is from mains and any generated energy is dumped as heat via resistors. ,....------1 Test rig management

Combustion Engine

Power analyser Rapid prototyping tool

Chopper & resistance Power supply

Torque

Fig. 6: Scheme of ELEGT installation on test engine. Test equipment set-up A major part of the effort was dedicated to the electric part of the set up. The challenging control of a high power density, high speed electric motor is embedded right from the start in the electronic control system of the diesel engine. Lubrication and water cooling (not shown in the diagram) of the turbocharger are provided from the test facility, independent from the engine. Oil and coolant are supplied at the correct temperatures and pressures even without the IC engine running. Many functional tests of the bulky prototype electronics could thus be accomplished without exposing the first turbocharger prototype to high load conditions in a hot environment. It is noted that the Mkl.l electric motor can not accelerate the rotor to more than roughly 30 000 rpm, even with the compressor outlet disconnected from the intercooler, breathing free into ambient. Typical power flows through turbines are many times larger than what can be conveyed as current through an electric motor which is small enough to fit into a turbocharger's bearing housing. It is clear that the electric machine's assistance and recovery can account only for a "delta" around most turbocharger operation points. The ELEGT is built into a bearing housing that is derived from the Variable Geometry Turbine (VGT) turbocharger that is in production on CURSOR 8. The swallowing capacity of the ELEGT prototype can be matched for test purposes to the engine by means of a batch of fixed nozzle rings similar to the one that normally slides in ourVGT.

164

Load response, operation as a motor The chosen nozzle ring gives a good fuel economy at higher loads and speeds, at the price of poor response to demanded sudden load increases from part loads. In fact the engine operates in smoke limitation status for 7 seconds, during a typical response at lower speed, with this nozzle ring. This value is typical for wastegated turbo charging. Giving a 1 second long burst of electric power, corresponding to 0.3 Nm at the turbine shaft, the time in smoke limitation is reduced to 2 seconds. This is a value typical for VGT turbocharging, accomplished without changing the turbines geometry. Power for this test was provided by the mains.

Turbocharger speed limitation, operation as a generator The IC engine is operated at four speeds with 100 000 rpm turbo speed. Without changing the IC engine's fuelling, turbo speed is reduced to 96 000 rpm by generating about 1.5 kW electric, before converter. This corresponds to 0.09 Nm torque at the turbine shaft, applied during several minutes. Then the fuelling is increased until turbo speed is again 100 000 rpm. The previous turbo speed is run now with 5 to 9 % more power. The same scheme is shown here also for 90 000 rpm, where the 0.09 Nm torque at the turbine shaft reduce turbo speed to 85 000 rpm. This is shown in Fig.7. 20 18 16 14

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165

CONCLUSIONS The induction motor has been developed such that it can operate in a turbocharger, with the air gap that is necessary for the hydrodynamic bearings. The additional mass added to the rotor system is not detrimental to the shaft motion of the turbocharger and in some cases it improves it. The temperature of the stator is a limiting factor, with power levels of 2kW sustained continuously. Higher powers can be developed for short periods of time, providing the stator can be effectively cooled prior to the next event. The transient response of a turbocharger on the engine can be improved significantly with a built in electric motor, which reduces the engine's turbo lag and transient emissions. Mk2 will be able to generate electric power over a considerable area of the engine's operation range, whenever an air surplus allows this. Integration of an electric machine into the turbo-machinery for a diesel engine system has been achieved with a reliable first prototype, limited however in its temperature, respective power capabilities and electronics. A major effort is required on the integration of electric machinery into the architecture of future automotive powertrain systems.

REFERENCES 1) Millo, F., Mallamo, F., Pautasso, E., Ganio Mego, G.; The potential of Electric Exhaust Gas Turbocharging for HD diesel engines, Society of Automotive Engineers World Congress, Detroit, April 2006 2) Pautasso, E.: "Numerical simulation of the ELEGT (Electric Exhaust Gas Turbocharger) solution for Heavy-Duty diesel engines", Mechanical Engineering Graduation Thesis, Politecnico di Torino, Turin, Italy, 2004. 3) Bumby, J.R., Spooner, E., Carter, J., Tennant, H., Ganio-Mego, G., Dellora, G., Gstrein, W., Sutter, H., and Wagner, J.; "Electrical machines for use in Electrically assisted turbo-chargers", IEE Int. Conf. on Power Electronics, Machines and Drives, Edinburgh, MarchiApril2004, pp344-349. 4) Bumby, J.R., Crossland, S., Spooner, E., Carter ,J.: "The development of turbocharger accelerator motors and drives and their integration into vehicle electrical systems", Proc. of the IEE Conference on Automotive Electronics, London, March 15-16, 2005 © Holset Engineering Co Ltd

166

A Numerical Study of the Performance Characteristics of a Radial Turbine with Varying Inlet Blade Angle Liam Barr, MEng, AMIMechE Stephen W. T. Spence, CEng, PhD, MIMechE, MIEI Tony McNally, PhD, BSc Queen's University ofBelfast, UK SYNOPSIS The inlet blade angle of radial turbines is almost invariably fixed in the radial direction to avoid imposing additional bending stresses at the blade tips. However, a small number of published studies have shown that changing the inlet blade angle can shift the efficiency characteristic to bring improvements at off-design conditions. Such a move in the efficiency characteristic would benefit turbocharger performance where the turbine typically experiences lower than optimum velocity ratios while accelerating during engine transients. This paper details a numerical study of varying inlet blade angle. A ID performance prediction routine has been used to forecast changes in the overall performance characteristics of a turbine with non-radial inlet blading, revealing a shift in the peak efficiency point and an increase in efficiency at lower values of velocity ratio (UlC) when using a back swept blade angle of 30°. The numerical modelling strategy, which was fully validated against existing experimental measurements, was used to verify the findings of the 1D analysis and show how the back swept blade angle significantly reduced regions of flow recirculation at inlet during off-design operation. FEA revealed stress levels up to 51 % of the material yield strength at the inlet region of the pressure side on the 30° back swept blade. A brief review of new nanocomposite material technologies is presented to demonstrate the potential they provide for increased freedom in aerodynamic blade design. NOMENCLATURE 10

3D CFD FEA

One dimensional Three dimensional Computational Fluid Dynamics Finite Element Analysis

CNTs PR UlC

Carbon nanotubes Turbine pressure ratio Blade speed velocity ratio

INTRODUCTION The radial turbine is known to deliver optimum efficiency at a blade speed velocity ratio (UlC) of approximately 0.7. However, in vehicle turbocharging applications the turbine spends a significant part of its operating time accelerating from relatively low speeds in response to engine transients. Consequently, the turbine often operates at a UIC value well below the optimum design value, and rarely above the optimum value. Transient engine performance, and particularly transient exhaust emissions, are closely related to the level of boost pressure achievable. Rodgers [1] stated that a 1% increase in turbine performance would result in around 5% more torque available for turbocharger acceleration and significantly reduce the time taken for the turbocharger to reach full boost pressure during a transient. Therefore, improvements in turbine efficiency at

169

lower values of VIC would yield significant benefit during the engine transients that account for much of engine operating time. Turbine performance at lower than optimum VIC values may be enhanced by employing a back swept inlet blade angle. However, the need to avoid incurring additional bending stresses in highly stressed turbine blades means that there are practically no examples of commercial turbines with non-radial inlet blading and there have been few published studies of non-radial inlet blade angles. Most notably, Mulloy and Weber [2] published their findings from a series of experimental tests conducted on radial turbine rotors with unconventional inlet blade shapes intended to reduce the offdesign losses. Three rotors were produced and tested: a forward curved thick-bladed rotor, a round nosed thick-bladed rotor and a backward curved thick-bladed rotor. The actual value of inlet blade curvature from the radial direction was not specified. The tests, which were conducted by varying rotor speeds at constant turbine pressure ratio, showed that at low rotor speeds the backward curved blades appeared to align better with the flow yielding an efficiency increase. Conversely, the forward curved blades demonstrated better flow alignment and efficiency at higher VIC values. Mass flow rates followed similar trends, with the backward curved blades giving increased flow at lower UIC values. Figure 1 illustrates the change in measured performance. (Throughout the rest of this paper turbine blade angle is described by the terms 'forward swept' and 'back swept', as conventionally used to refer to centrifugal compressors, where forward swept blades are those with an inlet portion angled in the direction of rotation. Likewise, back swept blades are those with an inlet portion angled against the direction of rotation.) Efficiency Vs U/C at 1.S PR

so 75

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Figure 1 Impact of rotor inlet blade angle on turbine efficiency [2] Mulloy and Weber commented on potential problems that might arise with these unconventional blade shapes. The thick blades produced a substantially tighter radius of curvature in the blade-to-blade plane, which it was thought could result in severe velocity gradients producing exit flow distortion, thus increasing the difficulty of

170

recovering the exit velocity head. Another concern was the increased moment of inertia and magnified stress levels that would arise from the thick blading. Meitner and Glassman [3] detailed the modifications made to an existing I D procedure for predicting the off-design performance of radial turbines to incorporate a rotor slip factor correlation that included rotor blade sweep at inlet. The resulting computer programme was used to predict the performance of a radial turbine with back swept inlet blade angles of 0°, 15° and 30° from the radial direction. The authors showed that there was an efficiency advantage at high values of work factor (corresponding to lower than optimum U/C values) as a consequence of employing backward swept blades at rotor inlet. Hakeem [4] presented an equation derived from a ID treatment of inlet and exit velocity triangles to show how the velocity ratio U/C corresponding to the peak efficiency point varies as a function of the inlet blade and flow angle.

(U) C

= optimum

0.707

tan fJinlet

(1)

tana;nlet

The derivation of equation 1 assumes zero incidence between the inlet flow angle and the inlet blade angle {3;nlet. For a purely radial inlet blade angle, tan{3;nlet = 0 and the optimum U/C is 0.7, but for a back swept inlet blade angle the optimum U/C is less than 0.7. Fredmonski et al [5] conducted a research program with the objective of designing, fabricating and testing compact radial inflow turbines with equal or better efficiencies relative to conventional designs while having an axial rotor length 40% less than that of conventionally proportioned radial turbines. Two rotors were designed and tested, one of which incorporated a slightly back swept leading edge with the intention of minimising losses associated with positive incidence. While it was not the focal point of their investigation, Fredmonski et al noted that the back swept rotor attained an efficiency of 88.1 % at its design velocity ratio of U/C = 0.65, which represented an increase of nearly half a percentage point over the purely radial rotor. The authors did not specify the back sweep angle used, but concluded that the improved efficiency may have resulted from the non-radial inlet blading. ainlel

ONE-DIMENSIONAL PERFORMANCE PREDICTION

In order to facilitate 3D numerical investigation of the impact of rotor inlet blade angle, 1D performance prediction modelling was first conducted focused around a duty corresponding to a typical turbocharging application. The ID modelling procedure was based on the method described by Connor and Flaxington [6], and the target design point performance along with the overall turbine dimensions are presented in Table 1. The initial rotor design employed radial inlet blading, which was subsequently varied between values of +30° and -30° while maintaining the other principal turbine dimensions. In this treatment, as in the publication of Meitner and Glassman [3], a back swept inlet blade angle is designated as a positive angle measured from the radial direction, since it aligns with what is conventionally considered to be positive incidence at inlet to a radial turbine.

171

I . Ta ble I DeSI2n . pomt S]!eci'fiIca f Ions an dd'Imenslons Iior 1D analysIs PR (t-s) 2.5 Inlet tip diameter 90mm

Speed

80,000 rev/min

Inlet blade height

20mm

Outlet pressure

101325 Pa

Exducer tip diameter

81 mm

Inlet temperature

873 K

Exducer hub diameter

25mm

Mass flow rate

0.5 kg/s

Exd. RMS blade angle

45°

Rotor blade number

10

Exducer throat area

2740 mm2

Figures 2 and 3 present the predicted efficiencies and flow rates for the five different inlet blade angles plotted against velocity ratio, VIC. The plots are for a constant pressure ratio and, therefore, the UIC value is proportional to rotor speed. Less than optimum UIC values can be considered to represent a transient condition where the turbine is accelerating from a lower speed towards its design speed. In agreement with the results of Weber and Mulloy [2] and Meitner and Glassman [3], it is clear from Figs. 2 and 3 that back swept inlet blades deliver improved efficiency at the lower VIC levels, and as expected, a corresponding increase in flow rate since losses are reduced. From a ID perspective, this is because of better blade alignment with the more tangential relative inlet flow angles that arise at the lower rotor speeds. At around two thirds of the turbine design speed (VIC = 0.46), the 30° back swept rotor delivers an efficiency advantage of2.4% over the radial bladed rotor. While this advantage exists over much of the 'acceleration side' of the D/C characteristic, the 30° back swept rotor suffers a penalty of 1.3% in peak efficiency. The forward swept rotors offer advantages at UIC values above 0.7; however, this is of little relevance to either turbocharger applications or small gas turbine engines. 0.82,-------------------------------, 0.80 0.78

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0.60

0.65

0.70

0.75

0.80

0.85

Velocity ratio (U/C)

Figure 2 ID efficiency predictions of varying inlet blade angle at PR = 2.5

172

0.90

0.56

_y_------------------------------,

0.54 0.52 ~ 0.50

C

E0.48 ,

~~ 0.46 ~

...... 30 deg. back swept

,

- - - 15 deg. back swept

0.44

,

,

--Radial - . - . 15 deg. forward swept

0.42

- .. - 30 deg. forward swept 0.40

+---.,----r------,r------r--~--__._--_r_--_r_--_y_-->---l

0.40

0.45

0.50

0.55

0.60 0.65 0.70 Velocity ratio (U/C)

0.75

0.80

0.85

0.90

Figure 3 ID mass flow rate predictions of varying inlet blade angle at PR = 2.5 Although not presented here, a range of pressure ratios between 1.5 and 3.9 were investigated and it was noted that while the overall trends were the same as presented in Figs. 2 and 3, the efficiency benefits of the back swept rotor at low VIC values were greater at lower pressure ratios. DEVELOPMENT AND VALIDATION OF NUMERICAL MODEL Since the initial I D analysis could not account for the real flow structures existing in the rotor inlet region, a 3D numerical investigation was undertaken using the CFX-IO Computational Fluid Dynamics (CFD) software. In order to carry out calibration and benchmarking of the numerical modelling strategy, validation was undertaken using results from an extensive experimental performance investigation of a 99 mm radial turbine with a vaned stator [7]. The construction of the turbine test rig in question and the measurements taken provided confidence that the turbine flow was axisymmetric in nature, and consequently only a single blade passage was modelled. CFX-TurboGrid 1.6, a turbomachinery specific grid generation package, was used to create a single passage mesh of the stator and rotor. The stator and rotor meshes each comprised 400 000 cells. Although this was more than the minimum requirement indicated by the grid independence study, it was considered desirable in order to effectively capture the complex 3D flow patterns within the blade passage. The cell density of the grid was similar to that used by Dunham and Meauze [8]. Both stator and rotor grids were constructed of H-grid cells, with particular attention being directed towards avoiding extreme skew angle values; the minimum skew angle was 25°, which occurred at the rotor trailing edge. There was no endwall clearance in the vaned stator, but the rotor included a blade tip clearance of 0.4 mm, which was modelled using 10 equispaced cells in the spanwise direction. The 'Stage' interface option was employed between the stationary and rotating domains, which takes the circumferentially averaged fluid properties at exit from the stator domain and applies them to the inlet of the rotor domain. Figure 4 shows the stator and rotor grids used, although the cell density has been reduced for the sake of clarity in the illustration.

173

Stator Domain

Rotor Domain

Figure 4 Stator and rotor grids used for the CFD validation study

The k-& turbulence model was used with a blend factor of 1.0 chosen as the advection scheme setting. Setting the blend factor to 1.0 is the equivalent to using second order differencing to calculate the advection terms within the discrete finite volume equations and is deemed to be the most accurate of the available advection scheme settings. The walls were assigned smooth, adiabatic, no slip conditions. The solution was considered to have reached convergence when the maximum values of all of the residuals and global imbalances fell below 5.0x 10-4• Experimental values for total pressure and total temperature at inlet and static pressure at outlet were used as boundary conditions. The flow upstream of the stator was assumed to approach in a radial direction. Figure 5 compares the numerically predicted and the measured turbine performance at a range of pressure ratios across a constant speed line corresponding to 55 000 rev/min corrected to 288K inlet temperature. The predicted efficiencies lie within a band of 0% to 4% below the measured values, while the predicted mass flow rates are within 1.7% and 3.5% above the measured values. With these results, validation had been achieved and a high degree of confidence had been established in the numerical modelling strategy described.

174

0.80

-r------------------------------.-

0.75

~

.. ····x·······x

0.70

1l

:e:LI

0.65

.~ ••. ~ ••.. ~ •...x •.... ~ ..... ·X··· ···.x

0.60

:::

~ E!

Measured efficiency Predicted efficiency - 0 - Measured mass flow rate x Predicted mass flow rate

-{)r-

~

x

0.24 ...

~

+----r--~--_.r---._--,_--_r--_.----+O.22 ~ 2.0

2.2

2.4

2.6

2.8 3.0 Pressure ratio (t-s)

3.2

3.4

3.6

Figure 5 Comparison of numerically predicted efficiency and mass flow rates with experimental data NUMERICAL STUDY OF INLET BLADE ANGLE A 3D radial element blade geometry typical of a turbocharger rotor was created using the CFX-BladeGen software based on the overall dimensions from the ID analysis given in Table 1. This baseline radial rotor geometry was then modified to incorporate blade curvature at inlet to yield two further rotor geometries for 15° and 30° back swept angles. Blade inlet angle was the only parameter adjusted; all other principal dimensions remained unchanged, including the rotor exit blade angle and the exducer throat area. The forward swept blades were not considered for further investigation since the potential aerodynamic advantages were clearly well outside the operating range relevant to turbochargers. Figure 6 shows the rotor blade geometry for the radial and 30° back swept rotors. The blade curvature was limited to a short distance of approximately 5% of the meridional length.

(a) Radial blade (b) 30° back swept blade Figure 6 Comparison of inlet curvature for the radial and 30° back swept blades Numerical models were constructed for each of the three rotor blade configurations in line with the strategy detailed in the preceding section. Two separate sets of operating conditions were modelled: a rotor speed of 80 000 rev/min and PR =

175

2.5 (Ule = 0.60), and a reduced rotor speed of 50 000 rev/min at the same PR = 2.5 (Ule = 0.38). The influence of the inlet blade angle on the efficiency and mass flow rate, as predicted by the numerical model, is shown in Fig. 7. At Ule = 0.6, the efficiencies of the different inlet angles are practically equal. At the off design condition where U/e = 0.38, the 30° back swept blade shows a clear efficiency advantage of 2% over the radial blade. Less significant are the changes in mass flow rates, which follow the trends of changing efficiency, as expected. 0.80.,---------------------.,t;.

0.75

Radial

o

o 15 deg. back swept

-;,

o 30 deg. back swept

60.70

!6.~ 0.65 WJ

0.60

o o

0.64 ~

t;.

0.55

0.62 o

0.60

o

8

t;.

0.40

0.45 0.50 0.55 Velocity ratio (U/C)

0.60

~

to!

~

0.58.g

+------,-----,-----.-----,------.----+ 0.35

C

0.56

~

0.65

Figure 7 Influence of inlet blade angle on efficiency and mass flow rate

The streamline distribution plots presented in Figs. 8 and 9 clearly show the reason for the variation in loss levels with the different blade angles. The streamline plots represent the mid-span location in the blade-to-blade plane; only the region around the blade leading edge has been shown, and not the full blade passage. At the design speed of 80 000 rev/min (Fig. 8) the relative flow is seen to approach the rotor in an almost radial direction, with the radial blade experiencing some flow reversal next to the pressure surface just after the leading edge. At the off-design condition of 50 000 rev/min (Fig. 9), strong positive incidence is evident at rotor inlet and a substantial zone of recirculation results as the flow separates from the suction side of the blade at the leading edge. However, in spite of the inlet curvature extending over only a short length of the blade, the back swept blade is much more tolerant of the positive incidence and, although still present, the extent of the recirculation on the suction side is much less than for the radial blade. The presence and extent of the recirculating flow was responsible for increased losses and the reduced mass flow rate experienced by the radial rotor at off-design conditions. Although further plots at other spanwise locations have not been shown here, it was apparent that the recirculation zone was concentrated in the half of the passage closest to the hub, and was less evident next to the shroud side. Hence, the angle of inlet blade curvature could be varied from hub to shroud for optimum aerodynamic advantage. It is clear that back swept blading at rotor inlet could offer aerodynamic performance gains for turbocharging applications, where the turbine spends a significant

176

part of its duty cycle at lower than optimum speeds, experiencing strong positive incidence.

(a) Radial rotor (b) 30 0 back swept rotor Figure 8 Streamline distribution at turbine inlet for the radial and 300 back swept blades at 80 000 rev/min (UIC = 0.60)

(a) Radial rotor (b) 30 0 back swept rotor Figure 9 Streamline distribution at turbine inlet for the radial and 300 back swept blades at 50 000 rev/min (UIC = 0.38)

FINITE ELEMENT ANALYSIS OF INLET BLADE ANGLE A stress analysis was performed on both the radial blade and 30° back swept blade using the Finite Element Analysis (FEA) package ABAQUS 6.5-1. Due to the rotor's cyclic symmetric geometry only a single periodic section of each rotor was modelled. Both models had a fillet radius of Imm along the blade root and contained a scallop. Material properties ofInconel 718 at 649°C were used. The rotors were analysed at an angular speed of 80 000 revs/min from which the associated blade root stresses where extracted and compared. A mesh seed independence study was performed in order to identity the number of cells required to adequately predict the levels of stress within the

177

model. It was found that using upward of 330,000 Tri cells for a single periodic model did not have a significant impact on the value of stress at various mesh nodes. Figure 10 shows the periodic section of the 30° back swept turbine (a) and periodic mesh (b) with reduced cell density for the purpose of clarity. The maximum blade root stress throughout the inlet region was found to be 26MPa and 522 MPa for the radial and 30° back swept rotors respectively, located on the pressure side of each rotor. This equates to an increase from 2.5% to 51 % of the material yield strength when changing the inlet blade angle from radial to 30° back swept. Although this was seen as quite a large increase in stress it was confined to a relatively small area within the inlet blade region, which could be further reduced with a larger fillet radius along the blade root. Localised yielding was observed in the centre of the hub on the back-face of both the radial and 30° back swept rotors. Figure 11 shows a contour plot of the stress distribution along the blade root at the inlet region of the radial and 30° back swept rotors.

(a) Periodic Section (b) Periodic mesh Figure 10 Periodic section and mesh of the 30° back swept rotor used for the FEA analysis

522MPa

/A=

385 MPa 309 MPa

Y

Y

Y

r

/

/

61 MPa 121 MPa

\"

/ .

26MPa

~.

182MPa

(! ,/ \. ! ;

/

)1

(a) 30° back swept rotor (b) Radial rotor Figure 11 Stress distribution at rotor inlet for the radial and 30° back swept rotors at 80 000 rev/min

178

METAL I CERAMIC NANOCOMPOSITES

Conventionally radial element blades have been used in turbine rotors to avoid additional bending stresses. However, advances in material technologies may provide the aerodynamic designer with freedom to depart from the radial constraints. Ceramic materials have a much lower density than conventional nickel alloys and stainless steels, and would therefore incur lower levels of bending stress in non-radial rotating blades. While much development testing has been conducted with ceramic turbine rotors in the past, they have not been widely employed in commercial products. A primary reason is the low toughness of ceramics, which can give rise to sudden failure and wide variations in rotor life. Recent developments in the use of nano-materials is enhancing the properties of conventional metals and ceramics and yielding new classes of materials called nanocomposites. These composites provide the potential for ceramics with significantly increased toughness levels, or metal alloys with improved high temperature properties. Nanocomposites constitute a novel class of material where a particle having at least one dimension on the nano-scale (10- 9 m) is dispersed in a matrix material, such as a polymer, metal or ceramic. Nanotubes, specifically carbon nanotubes (CNTs) are attracting intense research interest since they were identified by Iijima [9] working at the NEC corporation in Japan in 1991. CNTs have many unique properties, including having a tensile strength and modulus up to 20 times greater than stainless steel, conducting electricity better than copper and having similar thermal conductivity to diamond [10]. These properties combined with their size, shape, large aspect ratio and low density, typically one sixth that of steels, make them ideal candidates for use in novel composite materials. Guo-Dong Zhan and colleagues [11] at the University of California, Davis have reinforced alumina with up to 10 wt.% CNTs and reported a five fold increase in fracture toughness compared to the neat ceramic. Scientists at the NASA Glen Research Centre have developed CNT based coatings for two-phase oxide ceramic eutectics, AI203/Zr02 (Y203) which had enhanced friction and wear properties, both in air and in ultrahigh vacuum. Further research at the centre has also shown that the sintering resistance of certain ceramics can be increased by up to an order of magnitude on addition of CNTs. They reported that the grain size of an alumina ceramic after exposure at 1450 °C was decreased by an order of magnitude on addition of 5wt.% CNTs, indicating the potential for significant increase in the operating temperature of such materials. Bahizsi et al. [12] prepared carbon nanotube reinforced silicon nitride composites with improved bending strength and elastic modulus compared to matrices with added carbon fibre or graphite. Nanotubes made from other elements, such as Boron Nitride have also been synthesized recently. These nanotubes oxidize at higher temperatures than CNTs and their incorporation into ceramics offer higher temperature performance and have better anti-corrosion properties compared to CNT filled ceramics, but they are denser than CNTs. Carbon nanotubes have also been used to reinforce metal matrix composites. Dong et al. [13] reported an increase in fracture toughness, wear resistance and hardness of a CNT reinforced copper composite. Similar improvements were reported by Kuzumaki et al. [14] for a CNT titanium nanocomposite, and Laha et al. [15] having synthesized and characterized a CNT aluminium nanocomposite. Incorporation of CNTs into both ceramics and metals offer exciting developments in materials for turbomachinery applications. It may be possible to produce composites and coatings with a combination of enhanced toughness and hardness, increased thermal

179

conductivity and lower density. The machine-ability and formability of these materials may also be improved. It therefore seems that while past turbine designs have been constrained to use a radial element blade design, nanocomposite materials may provide the aerodynamicist with new freedom when defining blade shapes and angles. CONCLUSIONS

An increase in turbine efficiency, particularly at lower than optimum values of VIC, would result in more torque available for turbocharger acceleration, increasing boost air pressure during engine transients, benefiting engine response and emissions. A review of existing literature demonstrated the potential for enhancing turbine efficiency at low VIC through the use of back swept turbine inlet blades, although published performance data relating to varying inlet blade angle were very limited. A ID analysis for a typical turbocharger turbine indicated an efficiency increase of 2.4% at VIC = 0.46 when the inlet blade angle is back swept 30° from the radial direction. The back swept blade result in a peak efficiency decrement of 1.3% compared with the conventional radial blade. The turbine mass flow rate was found to increase with increasing efficiency. A thorough validation of the numerical modelling strategy was conducted using experimental performance measurements from a similar sized turbine rotor. Two versions of radial turbine geometry were designed with radial and 30° back swept inlet blading. Subsequent numerical modelling showed a 2% increase in efficiency with the back swept blade at VIC = 0.38, but identical efficiency levels at VIC = 0.6. The numerical models revealed a strong recirculation near the suction side of the radial blade at inlet under low VIC, (positive incidence) conditions, which was almost completely eradicated with reduced losses when using the back swept blade angle. The maximum stress throughout the inlet region occurred on the pressure side of the 30° back swept blade and equated to 51 % of the material yield strength. Stress considerations have conventionally constrained turbocharger turbines to adopt a radial element blade design. However, new advances in nanocomposite ceramics and metal materials present the possibility of non-radial blade designs to enhance turbine efficiency at more heavily loaded conditions. ACKNOWLEDGMENTS

The authors would like to acknowledge the technical support of staff at ANSYS Europe in the use of the CFX-IO.O software. Thanks are also due to John Doran for providing turbine measurements for validation purposes. REFERENCES

1. Rodgers, C., & Rochford, K., 2002, Small Turbocharger Turbomachinery, Instn. Mech. Engrs. Turbocharger Conference. C602/003/2002. 2. Mulloy, 1.M. & Weber, H.G., 1982, A Radial Inflow Turbine Impeller for Improved Off-Design Performance, 27th International Gas Turbine Conference and Exhibit. ASME International Gas Turbine Conference, Paper no. 82-GT-101. 3. Meitner, P.L. & Glassman, AJ., 1983, Computer Code for Off-Design Performance Analysis of Radial-Inflow Turbines with Rotor Blade Sweep, NASA TP-2199.

180

4.

5.

6.

7.

8.

9. 10. 11. 12. 13. 14. 15.

Hakeem, 1., 1995, Steady and Unsteady Performance of Mixed Flow Turbines for Automotive Turbochargers, Ph.D. Thesis, Imperial College of Science, Technology and Medicine, London. Fredmonski, A. J., Huber, F. W., Roelke, R.J., & Simonyi, S., 1991, Design and Experimental Evaluation of Compact Radial Inflow Turbines, NASA Lewis Research Centre, ReportAIAA-91-2127. Connor, W. A. and Flaxington, D., 1994, A One-Dimensional Performance Prediction Method for Radial Inflow Turbines, Instn. Mech. Engrs. Turbochargers and Turbocharging Conf., Paper No. C484/041194, pp. 271-282. Doran, W. J., 1999, An Experimental Assessment of the Effects of Shroud Profile on the Performance of a Radial Inflow Turbine, PhD Thesis, School of Mechanical and Manufacturing Engineering, Queen's University of Belfast. Dunham. J., and Meauze, G., 1998, An AGARD Working Group Study of 3D Navier-Stokes Codes Applied to Single Turbomachinery Blade Rows, ASME International Gas Turbine Conference, Paper no. 98-GT-50. S. Iijima, Nature, 354, 7, 56, 1991. J. Li, Y. Lu, Q. Ye, M. Cinke, J. Han, M. Meyyappan Nano Lett., 3, 929, 2003. G-D Zhan, J. D. Kuntz, J. E. Garay and A. K. Mukherjee App!. Phys. Lett. 83, 6, 1228,2003. C. Balazsi, Z. Konya, F. Weber, L. P. Biro, P. Arato Mater. Sci. Eng. C 23, 1133, 2003. S. Dong, J. Tu, X. Zhang Mater Sci. Eng. A 313,83,2001. T. Kuzumaki, K. Miyazawa, H. Ichinose, K. Ito J. Mater. Res. 13,2445, 1999. T. Laha, A. Agarwal, T. McKechnie, S. Seal Mater. Sci. Eng. A 381,249,2004.

181

EXPERIMENTAL STUDY ON THE PERFORMANCE OF A VARIABLE GEOMETRY MIXED FLOW TURBINE FOR AUTOMOTIVE TURBOCHARGER Srithar Rajoo and Ricardo Martinez-Botas

Department of Mechanical Engineering Imperial College London SW7 2AZ Exhibition Road, London SYNOPSIS

This paper investigates a variable geometry (VG) mixed-flow turbine with a novel, purposely designed pivoting nozzle vane ring. The nozzle vane ring was matched to the 3-dimensional aspect of the mixed-flow rotor leading edge. The VG mixed-flow turbine has been evaluated experimentally, in steady and unsteady conditions. The VG turbine shows higher efficiency and swallowing capacity at various vane angle settings compared to an equivalent nozzleless turbine. But the VG turbine unsteady performance was found to deviate substantially from the quasi-steady assumption, compared to a nozzleless turbine. The VG stator with the new unique vane design is expected to further enhance the mixed-flow turbine benefits in term of engine-turbocharger matching and transient performance. NOMENCLATURE

C c Cis

E M R T U W

Absolute flow velocity, mls Vane true chord, m Isentropic velocity, mls Turbine Power, W Mach number Universal gas constant, J/kg.K Temperature, K Rotor velocity, mls Relative flow velocity, mls

b 0

r s m

a

fJ r f//

Vane axial chord, m Nozzle throat width, m Radius, m Nozzle pitch, m Mass flow rate, kg/s Absolute flow angle, deg Relative flow angle, deg Specific heat ratio Tangential lift coefficient

subscript b Rotor blade v Nozzle vane o Total condition 4 Rotor inlet condition 5 Turbine exit condition(atmospheric) INTRODUCTION

Automotive turbochargers almost in their entirety are equipped with radial turbines, due to the efficiency superiority of radial design when compared to other turbine types. But radial turbines are limited in their potential, due to their radial leading edge. A mixedflow turbine differs from a radial turbine in that the leading edge is swept radially downward, as opposed to the zero blade angle of the radial turbine. Research has shown a substantial amount of exhaust gas energy to be available at velocity ratios of less than 183

0.7 - the point of highest energy recovery for a radial turbine. For the past three decades research has been carried out to explore the benefits of the mixed-flow turbine in terms of lower velocity ratio operation and higher swallowing capacity [1,2,3,4], For a nonzero rotor blade angle, the peak turbine efficiency point moves to a higher expansion ratio. Shifting the peak efficiency point to a higher expansion ratio is advantageous in a turbocharger application, which is subjected to pulsating flow from the reciprocating engine, where the greater energy of the flow is contained at high pressures. VGTs have been widely based on radial turbines, particularly in automotive application. Mixed-flow turbines with lower inertial characteristic, coupled with variable geometry operation have the capability of enhancing existing versions of VGTs. Unfortunately, there are not many VGTs currently utilizing mixed-flow turbines, and for the very few which exist [5], the nozzle vane used is adapted from radial turbines, these are invariably straight vanes. Due to the non-radial inlet of the mixedflow turbine, the use of straight vanes is not an optimized option as it creates nonuniform inter-space between the nozzle vane ring and the rotor, hence this work to study an alternative nozzle vane design and its implication on a mixed flow turbine performance. Most available published work on VGTs is still based on steady-state data. The present experimental facility offers the ability for both steady and pulsating flow testing of the VGT, equipped with a mixed flow turbine. It has been shown in the past research that the turbine performance during pulsating flow departs substantially from the quasisteady assumption [3,4]. Thus unsteady-state VGT testing is required in order to better understand the mixed flow turbine characteristics at different nozzle positions, hence the need for an optimized nozzle design for a mixed flow turbine. NOZZLE VANE DESIGN This section describes the design process of the nozzle vane for a mixed flow turbine. The nozzle vane ring was designed based on common axial turbine blade design method and through conformal transformation the designed axial blade was converted to circumferential nozzle ring. As the nozzle ring is preceded by a volute which will provide sufficient swirl to the flow, the nozzle vane is left un-cambered, and effort was concentrated on 3-dimensional variation of the nozzle vane to fit the leading edge of the mixed flow turbine. The nozzle vane profile was designed based on NACA airfoil 0015. The volute used in this study was designed based on a commercial turbocharger (HOLSET H3B), but enlarged for a nozzle vane ring fitting with an inlet area over radius ratio (Air) of 33mm and an exit flow angle of 70°. The nozzle vane ring is aimed for variable geometry operation and the operating range of the vane angles were determined based on the possibility of exploring wider range of turbine performance. The incidence at the rotor inlet is, i=fJ4-f3b (1) The optimum incidence is considered between _20 0 to -30 0 [6]. The rotor velocity, U4 is U4 = C4 sina4 -W4 sinfJ4 (2) Since Cm4 and Wm4 are the same,

W =C4cosa4/ 4 /cosfJ4 Substituting Eq. (3) into (2) will eliminate W4 , U 4 = C4(sina4 -cos a 4 . tan fJ4)

184

(3)

(4)

For the mixed flow rotor used in this study, the blade angle, fit, is 20 0 • By determining the incidence angle, relative blade angle, fJ4 is calculated using Eq. (1). The velocity of the rotor can also be expressed as, U 4 = m·r4, whereby m = 2nN;(;'0

(5)

and according to Euler Equation, the power of the turbine can be approximately deduced as,

E= TilUl (6) By determining the power output or the turbine revolution, U4 is calculated. Consequently, the maximum permitted U4 is determined as the maximum power and turbine rpm of the test-rig are known. With the values of fJ4 and U4 known, using Eq. (4) value of C4 is calculated for different absolute velocity angles, U4. The absolute flow velocity is then normalized and expressed as Mach number, M 4 , C4 C4 _ M 4(7) speed of sound ~ yRTo4 The flow Mach number at different absolute flow angle can be calculated for a range of turbine speeds using Eq (7). By limiting the Mach number to one, the suitable range of inlet flow angle for the power spectrum in the current application was found to be 40 0 to 80 0 • The determined inlet flow angle can be expressed as a function of the nozzle geometry [7], (8) The nozzle throat, 0 and pitch, s are dependant on the number of nozzle vanes. Number of nozzle vanes or indirectly the blade spacing is an important factor in loss contribution and it is determined through Zweifel's criterion, IJI

= 2(s I b)cos 2 a31tana2 -tana31, whereby

b = ccosav

(9)

The optimum pitch/chord (sib) ratio is a compromise between friction losses and good flow guidance. The optimum tangential lift coefficient, IJI is between 0.75-0.85, which will result in best pitch/chord compromise [6]. Since different nozzle exit flow angle will result in different pitch/chord ratios, it was decided to match the optimum pitch/chord ratio at mid-range of the flow angles. Another compromise considered in deciding the number of nozzle vanes and the chord length for a variable geometry system is the suitable nozzle space for capability of pivoting within the range needed. After an iterative analysis as mentioned above, the number of nozzle vanes was chosen to be 15. In order to match the nozzle vane to the mixed flow rotor leading edge, it was decided to apply lean stacking for the whole vane chord. The lean angle was chosen by trial and error and finally determined at 50 0 relative to volute hub surface, which is the same as the mixed flow rotor cone angle. The resulting meridional projection of the assembly (see Figure 1) shows good match between the nozzle vane and the rotor. Nevertheless, the lean stacking and conformal transformation leads to non-uniform alignment between the hub-hub and the shroud-shroud surfaces of the adjacent nozzle vanes. These result in non-uniform suction-pressure surface interaction. To resolve this, the shroud end vane profile was elongated to match the surface interaction at the hub end, and finally producing even interaction. As illustrated in the Figure 1, this results in a small degree of sweep in the nozzle vane trailing edge, but without jeopardizing the mixed flow rotor inlet match.

185

\) l/1j I~ l

I

I I __.JI

.

~

"~

'\: ~.~ 500. __ ,__

Figure 1 Meridional Projection and the Lean Nozzle Vane Assembly

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Nozzle assembly

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(a)

Figure 2 Turbocharger Test-Rig Schematic Diagram EXPERIMENTAL SETUP

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 state testing of single and twin entry turbines as well as unsteady testing with simulated engine pulsations, as previously reported [3,4]. Furthermore, the latest instalment of eddy current dynamometer enables turbine testing within larger velocity ratio range [8]. A schematic diagram of the turbine test rig is shown in Figure 2. The test-rig is supplied by Howden screw-type compressors, capable to deliver the necessary compressed air of up to 1.2 kg/s mass flow rate at a maximum pressure of 5 bar (absolute). Air flow is heated to avoid the occurrence of water vapour condensation during the expansion process within the turbine rotor. The air pulse generator was designed to experimentally simulate the exhaust gas pulsation based on the shape of the cut-outs of the chopper plates, for the unsteady flow turbine testing. A variable speed D.C. motor controls the rotating frequency of the chopper plates, hence the frequency of the pulsation. 186

0.9 80% Speed

0.8 0.7 :I

~

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>; O.S

u c

CD '13 0.5

IE w

• A

0.4

*

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~

+

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(a)

o

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40' Vane Angle 50'

60' 65' 70' Nozzleless

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• 0.7

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ute Is Ci'7

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x10"

is III

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ft

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o

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~

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Po/Ps Figure 3 Turbine Performance Plots at Different Vane Angles: (a) Efficiency versus Velocity Ratio, (b) Swallowing Capacity STEADY-STATE TURBINE PERFORMANCE The newly designed lean vane with the volute and the pivoting mechanism were coupled to a first generation mixed flow turbine developed in Imperial College London [2]. The turbine was tested for different nozzle vane angles and due to the cold flow testing; the flow non-dimensional parameters were matched to the turbine in an actual engine operation. All the necessary flow parameters were measured at the measuring plane (see Figure 2) before the volute inlet, and the turbine speed and torque were measured directly at the dynamometer. Figure 3 shows the turbine efficiency against velocity ratio and the turbine swallowing capacity. The figure shows results for an 80% of the design speed testing which is approximately 48000rpm. The results presented are for different nozzle vane angles and comparison shown for the same mixed-flow turbine tested with a nozzleless volute [4). One significant observation from Figure 3 is the range of velocity ratio in the current results in comparison to the nozzleless result This is due to the new eddy current dynamometer eliminates the surge and choke limits as opposed to conventional turbine testing with compressor coupling. 187

The peak efficiency of the nozzleless turbine is 75% and for the vane angle of 60°, 65° and 70°, the turbine peak efficiency shows improvement, which is 79%, 80% 77% respectively. Meanwhile, for the 40° and 50° vane angles, the peak efficiency drops to 61 % and 68% respectively. This is due to the increase in separation losses in the nozzle as the vanes opens and deviate from optimum incidence condition. Nevertheless, the 50° and 40° vane angles show higher swallowing capacity from the nozzleless turbine with approximately 2% and 8% mass flow parameter improvement respectively between the pressure ratio of 1.5 and 2.0. The swallowing capacity improvement is due to a bigger volute than the nozzleless version and the opening capability of the nozzle vane passage without choking. As for the higher vane angles, the turbine shows capability to achieve higher pressure ratio at lower mass flow parameter, for instance at the vane angle setting of 70°, pressure ratio of 1.7 is achieved with 34% lower mass flow parameter compared to the nozzle less turbine. The change in the nozzle vane setting from 40° to 70°, results in turbine efficiency to increase significantly in the first 25° and then drops slightly, meanwhile the turbine swallowing capacity to reduce slightly in the first 10° and then drops significantly. At 80% design speed, the vane angle setting of 65° shows best performance at peak as well as at higher velocity ratio region. This implies that this setting is suitable for variable geometry operation so as to adapt to the lower incoming mass flow feed. 0.9.,------------------____, •

0.8

~~

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~ 0.6

c .~ 0.5

ffi

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•

•

0.4

0.3 0.2

(a)

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0.5

0.7

0.6

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188

The turbine performance with the new lean nozzle vane was also compared against an equivalent straight vane, which was initially used as baseline design. Figure 4 shows the turbine performance comparison at 80% design speed (48000rpm), with vane angle setting of 65°. As seen in the figure, the turbine efficiency with both the lean and the straight vane is fairly similar at peak condition, even though differences observed at offpeak condition. Due to the leaning, the new vane has higher wetted area than the straight vane. Nevertheless, with similar efficiency the lean nozzle vanes are capable of higher swallowing capacity, due to the larger nozzle area as an effect of leaning the vane. The pressure averaged mass flow parameter of the lean nozzle vane is approximately 10% higher than the straight vane. This indicates the suitability of the lean nozzle vane in turbocharger application, where higher swallowing capacity is often required without sacrificing the turbine efficiency. UNSTEADY TURBINE PERFORMANCE The performance of the turbine with lean nozzle vane was tested under pulsating flow condition for different vane angles. The pulsating frequency of the flow was 40Hz which approximately simulates a 3 cylinder engine running at 1600rpm. Testing was carried out at an average 80% design speed (48000rpm) and flow parameters were measured instantaneously for 1800 cycles. The detailed measuring and analysis procedures were described in Szymko et al. [9]. Figure 5 and Figure 6 illustrate the turbine instantaneous efficiency and swallowing capacity in a complete pulse cycle for 70° and 40° vane angle settings. The instantaneous efficiency is calculated based on isentropic condition measured at the measuring plane (see Figure 2), which is upstream of the volute inlet. Thus, the instantaneous efficiency refers to the stage from the measuring plane to the turbine exit. Also presented in the Figures 5 and 6 are the equivalent quasi-steady curve derived from steady flow testing of the turbine. It is observed that the turbine performance curves exhibit hysteresis loop, but with significant deviation from the quasi-steady condition, in comparison to the nozzleless turbine, as in Figure 7. The deviation is very substantial in the 70° vane angle setting, but reduces as the nozzle vanes open to 40°. During pulsating flow the turbine volute experiences continuous filling and emptying and in the 70° vane angle setting, due to high blockage of the nozzles, the emptying of the volute is delayed before the filling occurs in the consequent pulse,. 1.0

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Figure 5 Turbine Instantaneous Efficiency at 40Hz Pulsating Flow with 70° and 40° Nozzle Vane Angles 189

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1.6

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Figure 6 Turbine Swallowing Capacity at 40Hz Pulsating Flow with 700 and 40 0 Nozzle Vane Angles x10'S

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Figure 7 Nozzleless Turbine Instantaneous Efficiency (a) and Swallowing Capacity (b) at 40Hz Pulsating Flow This creates a back pressure and eventually mass accumulation in the volute. It is suspected that the slow emptying process of the volute may have resulted in higher mass flow rate measurement, which in reality does not reflect the true instantaneous mass flow rate impacting momentum on the rotor. This eventually leads to the lower efficiency calculation for the 70° vane angle setting. The high mass flow rate reading is evident in the Figure 6 and it is also noticed that the turbine experiences choking at high pressure ratio region in 70° vane angle setting, which further indicate the possibility of mass accumulation upstream of the nozzle vane. Similar choking behaviour was also observed for the 65° and 60° vane angle settings. On the other hand, at 40° vane angle setting, the turbine did not experience choking and loops more closely to the quasisteady curve. Table 1 shows the cycle averaged efficiencies and velocity ratios of the turbine at different nozzle vane angle settings as well as the nozzleless condition. These parameters are averaged using isentropic power average method, which was described by Szymko et al. [9]. The equivalent quasi-steady efficiencies were derived from steady curve at the corresponding velocity ratios. It is noticed at all nozzle vane angle settings that the cycle averaged turbine efficiency are significantly lower than the 190

Table 1 Unsteady Cycle-Isentropic Power Averaged and the Equivalent QuasiSteady Efficiencies U/Cis

Nozzle Vane Settings 40° 50° 60° 65° 70°

0.61 0.61 0.62 0.62 0.60 0.62

Nozzleless

Equivalent QuasiSteady EfficiencyffitsJ 0.58 0.67 0.77 0.78 0.75 0.74

Unsteady Cycle Efficiency (TJts) 0.48 0.49 0.53 0.53 0.50 0.62

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80

120

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160

200

240

280

320

360

Crank Angle (Deg)

Figure 8 Turbine Instantaneous Efficiency for One Complete Pulse Cycle

equivalent quasi-steady point and the nozzleless condition. At 70° vane angle, the unsteady cycle efficiency is about 33% lower than the equivalent quasi-steady efficiency and it gradually improves to about 17% at 40° vane angle. Figure 8 shows the turbine efficiency variation plotted for one complete pulse cycle for all the nozzle vane angle settings as well as the nozzleless condition. It can be noticed that the nozzleless turbine exhibit negative efficiency at the beginning and end of the cycle, which is the low pressure ratio region. The negative efficiency was explained by Szymko et a1. [9], as the consequence of turbine rotor impacting momentum on the flow at lower pressure ratio condition. As for the nozzled settings, the turbine did not exhibit negative efficiency at 70°,65° and 60° vane angle settings. This is due to the nozzle constantly providing sufficient flow momentum to the rotor during the pulse cycle. But at 40° and 50° vane angle settings, the turbine gradually exhibit negative efficiency, as the opening of the nozzle reduces the momentum of the flow in low pressure ratio region. It is also noticed that during the first 120° crank angles, where most ofthe isentropic power concentrated in the pulse, nozzleless turbine exhibits better efficiency than the nozzled turbine. For the rest of the cycle, the efficiency of the nozzleless turbine drops in comparison to the nozzled turbine. Another interesting observation is that at 65° vane angle setting, the turbine shows good efficiency towards the end of the cycle, which indicates better energy extraction from the lower pressure region of the pulse. 191

CONCLUSIONS A new nozzle vane ring has been designed to match the 3-dimensional feature of the mixed-flow rotor leading edge. The VG mixed-flow turbine performance was tested at different vane angle settings, under steady and unsteady conditions. The VG mixedflow turbine shows efficiencies at nozzle vane angle of 60°, 65° and 70° higher than the nozzleless turbine for much of the velocity ratio range, with highest peak efficiency of 80% obtained at 65° vane angle. The swallowing capacity of the VG turbine at fully opened condition (40° vane angle) is up to 8% higher than the nozzleless turbine. The VG turbine is capable of achieving high pressure ratio at lower inlet mass flow, which demonstrates its benefit in adapting to off-peak conditions during the turbine operation. Performance comparison with an equivalent straight vane design shows the higher swallowing capacity of the lean vane with similar efficiency at almost all pressure ratio condition. The VG turbine's unsteady performance shows substantial deviation from the equivalent quasi-steady assumption at all the vane angle settings, especially at close nozzle positions (60°, 65°, 70°), as much as 33% at 70° vane angle setting. The observed characteristics of the nozzled turbine under pulsating flow indicate the need to actively control the nozzle position to better adapt to the incoming flow, hence improving energy extraction. REFERENCES

2 3

4

5

6 7

8

9

Baines, N.C., Wallace, F.J. and Whitfield, A. 'Computer Aided Design of MixedFlow Turbines for Turbochargers', Transc. ASME, July 1979 VollOl 440-449. Abidat, M., Chen, H., Baines, N.C., 'Design of a Highly Loaded Mixed Flow Turbine', Proc. Instn. Mech. Engrs., 1992 Vol 206. Arcoumanis, C., Hakeem, 1., Khezzar, L. and Martinez-Botas, R.F. 'Performance of a Mixed Flow Turbocharger Turbine Under Pulsating Flow Conditions', Transc ASME 95-GT-210, 1995. Karamanis, N., Martinez-Botas, R.F. 'Mixed-Flow Turbines for Automotive Turbochargers: Steady and unsteady Performance' IMechE Int. J Engine Research, 2002 Vol 3 No.3. Baets, 1., Bernard, 0., Gamp, T., and Zehnder, M., 'Design and Performance of ABB Turbocharger TPS57 with Variable Turbine Geometry', 6th Int. Conf on Turbochargers and Turbocharging, Proc. of the IMechE, paper C554/017/98, 315325,1998. Japikse, D. and Baines, N.C., Introduction to Turbomachinery, Concept ETI Inc., USA and Oxford University Press, Oxford, 1994. Hiett, G.F. and Johnston, LH., 'Experiments Concerning the Aerodynamic Performance in Inward Radial Flow Turbines', Proc. of the IMechE, 1963 Vol. 178 Part 3I(II) 28-42. Szymko, S., Martinez-Botas, R.F., Pullen, K. R., McGlashan, N.R. and Chen, H., 'A High-Speed, Permanent Magnet Eddy-Current Dynamometer for Turbocharger Research' i h Int. Conf on Turbochargers and Turbocharging, Proc. of the IMechE, paper C602-026, 2002. Szymko, S., Martinez-Botas, R.F. and Pullen, K.R. 'Experimental Evaluation of Turbocharger Turbine Performance under Pulsating Flow Conditions', Proc. of ASME Turbo Expo, GT 2005-68878, 2005.

192

Turbocharger turbine performance under steady and unsteady flow: test bed analysis and correlation criteria Massimo Capobianco and Silvia Marelli Internal Combustion Engines Group (ICEG) Department of Thermal Machines Energy Systems and Transportation (DIMSET) University of Genoa - Italy

ABSTRACT Turbocharging is becoming a key technology for both gasoline and diesel automotive engines. A thorough knowledge of turbine behaviour under steady and unsteady flow conditions is a fundamental requirement for the improvement of engine performance, particularly in transient operation. To this end, a great deal of information can be obtained from investigations developed on dedicated test facilities. This paper presents a new arrangement of the turbocharger test rig operating at the University of Genoa. The results of an extensive experimental programme, focusing on the behaviour of the turbocharger regulating system, on turbine pulsating flow performance and its correlation criteria with steady flow results, are then analysed.

NOMENCLATURE Notations

f n p A K I

M p T

pulse frequency rotational speed pressure waste-gate opening degree ratio between mean pulsating and steady flow performance influence factor mass flow rate power temperature expansion ratio increase

Subscripts 3 4 nd t

M NSF p QSF S T WG

turbine entry turbine exit "non dimensional" (independent of inlet conditions) turbine referred to mass flow rate mean unsteady flow value referred to power average quasi-steady flow value static condition stagnation condition referred to waste-gate valve 193

Definitions and acronyms nnd = nJ-VT3 Mnd = (Mt . -VT3) / P3 et=P3/P4 GDI GP-IB NSF PC QSF VVA

turbine rotational speed factor turbine mass flow factor turbine expansion ratio Gasoline Direct Injection General Purpose Interface Bus Non Steady Flow Personal Computer Quasi Steady Flow Variable Valve Actuation

Note - Turbine performance parameters were defined according to total-to-static conditions in steady flow and static-to-static conditions in pulsating flow operation. INTRODUCTION

Due to the need to reduce CO2 emissions, coupled with increasing fuel prices, a great deal of research effort is currently being expended on the development of high efficiency powertrain concepts. In this regard, turbocharging can both improve automotive diesel engines and substantially enhance gasoline applications. For spark ignition engines, charge boosting is seen as a way of reducing engine displacement at constant rated power (downsizing concept) and its application in conjunction with other technologies, such as direct injection (GDI) (1) and fully flexible valve control systems (VVA) (2), can offer significant reductions in fuel consumption, especially at part load operation. In the meantime, the gasoline engine should be able to maintain its advantage over the diesel engine as regards exhaust pollutants. The application of turbocharging to automotive spark ignition engines has to face various problems related both to the specific operating environment (exhaust gas temperatures) and, principally, to functional aspects (3, 4, 5). There is clearly great interest, therefore, in carrying out dedicated investigations on small turbochargers for gasoline engine applications, focused on different targets such as: i) definition of compressor and turbine characteristics in an extended range; ii) evaluation of unsteady flow turbine performance and development of suitable correlation criteria between steady and pulsating flow operation; iii) analysis of the behaviour of applicable turbine regulating devices; iv) interactions between the turbocharger and the engine intake and exhaust system; v) development of unsteady flow performance prediction procedures. In order to further investigate the above aspects, measurements performed on a components test rig are particularly useful as they allow turbocharger performance to be investigated independently of the engine. In this case, the turbine is usually fed with compressed air which needs to be heated at least to a temperature level that prevents water condensation and outlet freezing problems. To this end, electrical heating is often preferred to a combustion chamber, due to its simplicity and ease of control (6). In any case, compressor and turbine performance characteristics are usually plotted in terms of "dimensionless" or "corrected" parameters regardless of inlet thermodynamic conditions (7). As regards steady flow operation, there are a number of difficulties connected with measuring turbocharger characteristics in a broad range. This is a basic requirement, especially for the turbine that usually operates under unsteady flow conditions and 194

instantaneously experiences expansion ratio levels varying in a substantially wider range than that considered in the maps provided by the turbocharger manufacturer. Extrapolation of turbine curves can generate significant errors, particularly in the field of the lowest expansion ratios. Consequently, if a quasi-steady flow approach is used to predict pulsating performance, a detailed definition of turbine steady flow curves is essential for improving calculation accuracy. Besides, experimental information on the effect of the regulating system setting (waste-gate valve or variable geometry device) on turbine characteristics is a fundamental input for theoretical simulation models and a pre-requisite for developing effective turbocharger control strategies. Test rig investigations are usually performed using the turbocharger compressor as a dynamometer. Since the compressor operating range is limited by chocked flow and surge, specific experimental techniques are required to widen the definition of turbine characteristics, especially as regards tests performed at low inlet air temperatures. To this end, the possibility of controlling the compressor supply pressure in a broad range, thus modulating its power absorption, allows the experimental definition of turbine curves be to considerably extended. If further expansion of measured characteristics for very low levels of the expansion ratio is required, specific devices can be used. For example, an original impulsive system acting on the compressor side of the turbocharger is presented in (8). This allowed the measurement of turbine characteristics near zero mass flow conditions to be significantly extended without modifying the turbocharger rotary assembly. Alternatively, high speed dynamometers can be used to absorb turbine power (9, 10). These devices can considerably extend the measurement of turbine characteristics and in some cases (10) also allow the actual torque produced by the turbocharger bearings to be evaluated. The main drawbacks of turbine dynamometers are complexity of design, restricted rotational speed range (usually below 80,000 rpm) and difficulties in coupling them to different turbines. Additionally, torque values are extremely low, thus requiring very sensitive measuring equipment. The design of a totally flexible turbocharger test rig also has to consider that the turbine usually operates under unsteady flow conditions and sometimes with partial admission. This adds further complexity to the test facility, requiring the introduction of a pulse generator system upstream of the turbine and the use of high speed measuring systems. The possibility of performing tests in unsteady flow conditions is an essential pre-requisite for investigating the effect of the main pulsating flow parameters on turbine performance, developing suitable comparison criteria with steady flow data and analysing the results obtained from theoretical models and quasi-steady flow prediction procedures (11). This paper presents the results of an extensive experimental programme developed on different turbochargers for automotive gasoline engines. The study was performed on the test rig operating at ICEG of the University of Genoa, which allows investigations to be performed both in steady and unsteady flow conditions. The experimental facility was recently upgraded (12) as regards its unsteady measuring capabilities: in the new configuration, tests can be focussed both on analysing the effect of the main pulsating flow parameters on turbine performance and on investigating the engine exhaust subsystem behaviour, including the effect of circuit geometry and of different engine valve actuation strategies. The steady flow characteristics of a small turbocharger turbine for downsized gasoline engines were measured over an extended operating range. The relevant results are presented and discussed in the paper, including the effect of the waste-gate valve 195

setting on turbine mass flow and efficiency. A detailed analysis on the sensitivity of the by-pass regulating device was also performed. Turbine unsteady flow operation was investigated with reference to a different turbocharger. Average pulsating flow performance was analysed, highlighting the influence of the main flow parameters and considering different correlation criteria between mean unsteady results and steady flow data. The turbine feeding circuit pressure diagrams and the results of quasi-steady flow calculations are also presented. EXPERIMENTAL FACILITY

The test rig operating at ICEG is a continuous flow apparatus which allows tests to be performed on individual components and subassemblies of automotive engine intake and exhaust circuits. The facility is particularly suitable for performing investigations on exhaust turbochargers due to the availability of two independent supply lines. In this case, the turbocharger turbine is fed with compressed air at low temperatures (up to 400 K) while the compressor, which acts as a dynamometer, can operate at a controlled pressure level by means of an appropriate regulating system. Turbine performance can also be investigated under unsteady flow conditions by using different pulse generator systems. The test facility layout has been fully described in previous papers (13, 14). Two different air compression stations are available, one of which consists of three electrically driven screw compressors providing a total mass flow rate of approximately 0.6 kg/s at a maximum pressure of 8 bar. Alternatively, a single stage centrifugal compressor with a delivery of up to 2.2 kg/s and a maximum compression ratio of 2.1 can be used. A proper regulating system makes it possible to feed both the turbine and the compressor supply circuit with air at controlled pressure levels. Mass flow rate in each line is evaluated by means of different measuring devices to allow accurate measurement of this parameter also in unsteady flow conditions. The turbine feeding line is fitted with an electrical heater to moderately raise the air temperature in order to avoid condensation and freezing problems during expansion. Two arrangements of the turbine supply circuit are currently available (Figure 1), allowing detailed investigations to be made both in steady and unsteady flow conditions. Pulsating flow at the turbine inlet can be provided by different pulse generator systems connected to the upstream portion of the feeding circuit through a plenum which acts as damping element and flow distributor. The two line configurations are addressed to the investigation of different aspects and can be easily interchanged by modifYing a few connections; in both cases, the upstream and downstream turbine measurement stations remain the same. The first circuit layout (referred to as arrangement A, Figure la) was extensively used during previous investigations developed at ICEG (8, 13, 14, 15, 16, 17) and was designed to perform parametric studies on the effect of the main unsteady flow parameters on turbine performance. Tests on single and two-entry components can be developed, independently controlling thermodynamic parameters at each entry. In both feed branches, pulsating flow is generated by a diametral slot rotating valve (15). The main pressure pulse parameters (amplitude and mean value) at each turbine entry can be controlled by correctly mixing two flow components (steady and pulsating one) in a Yjunction and adjusting upstream plenum pressure. The flow area diagram of each pulse generator can be changed by replacing a few stator and rotor parts, thus affecting the pulse shape. Dedicated flow control valves allow unequal admission conditions to be reproduced in the event of two-entry devices. A variable rotational speed electrical 196

a)

I"

Plenum

Engine headlqj::fi::W=p Manifold

Compressor air supply

Ir=~

~~__ [;r1~Turbo:arger

Figure 1 Arrangements of the turbine supply circuit on the ICEG test rig

motor allows pulse frequency to be adjusted within the typical range of automotive engine intake and exhaust systems (10 - 200 Hz). For two-entry components, the pulse phase angle can be changed in steps of 1110 of the period. A second configuration of the turbine feeding circuit (arrangement B, Figure Ib) was recently set up in order to more precisely reproduce turbocharger unsteady flow operation when matched to an automotive engine. Furthermore, experimental investigations can be extended to a subsystem level, including the effect of the exhaust circuit geometry and different valve actuation strategies. In this alternative layout of the turbine supply circuit, heated compressed air at the damping plenum outlet enters a flow distributor, designed to reproduce the reference engine cylinder block, on top of which a motor-driven cylinder head is connected. The opening of the engine valves determines unsteady flow in the exhaust circuit, the geometry of which can be easily changed to reproduce various manifold configurations. The effect of different valve actuation strategies on turbine performance can also be investigated since the cylinder head can be fitted with a fully flexible valve actuation system (VVA) which allows any valve opening profile to be reproduced, including cylinder deactivation. In a next stage of the investigation, a further upgrade of the turbine supply circuit will be developed, including butterfly regulating valves in each duct of the flow distributor, with a view to reproducing engine transient flow operation. Turbocharger and cylinder head lubrication is provided by a dedicated circuit, based on two independent lines where oil temperature and pressure can be controlled and kept constant at set levels by means of suitable regulating systems. The oil mass flow rate and lubricant inlet and outlet temperatures are measured in order to estimate turbocharger mechanical losses. The test rig is fitted with a PC-controlled automatic data acquisition system. Average and instantaneous wall static pressures are evaluated by high frequency response straingauge transducers, while temperatures are measured by platinum resistance thermometers and turbocharger rotational speed and pulse frequency by inductive probes. Mean turbine and compressor mass flow rate are estimated by a laminar flow meter and a sharp edged orifice respectively. Transducer signals are managed and processed by various instruments connected to a GP-IB. In the event of unsteady flow tests, pressure transducers are mounted near the duct walls to avoid inaccuracies deriving from the damping effect of the signal connecting lines. High-speed sampling devices are used to simultaneously acquire different 197

pressure signals in transient conditions. Interactive procedures in LabVIEW® environment were set up to control the acquisition process and calculate turbocharger performance parameters both in steady and pulsating flow operation. Additionally, specific analytical tools were developed to allow post-processing of instantaneous pressure signals. DEFINITION OF TURBINE STEADY FLOW CURVES As mentioned above, the ICEG test facility allows an extended definition of components steady flow characteristics to be made. This aspect is particularly important in the case of turbocharger compressors and turbines which are generally simulated within engine global models using steady flow curves. Referring to a small turbocharger matched to a downsized gasoline engine, Figure 2 shows the turbine steady flow characteristics measured on the ICEG test rig. The

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200

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40000

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80000

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Figure 8 Experimental temperature: ttot,in=720°C

Figure 9 Experimental temperature: thermocouple #1

Numerical simulations

Steady state conjugate calculation Some observations concerning the heat transfer in the turbine wheel are discussed for the high load operating point. The occurrence of high temperature gradients shows the effects more clearly than for partial load. In figure 10, the temperature distribution is shown at 50% span. Although a frozen rotor boundary conditions is used, the temperature is equally distributed. Only the tongue of the volute causes some distortions which are mixed out downstream. The blade-to-blade view of the fluid temperature clarifies the high thermal load of the pressure side compared to the suction side. Nevertheless, an equalized temperature distribution is found inside the blade. Along lines of constant radius given in figure 11, static temperature profiles at 6 locations give information about the direction of heat transfer. In figure 12, the tip area along line "A" is heated at both sides of the blade. Downstream, along line "B", the material is heated only at the pressure side and along lines "C" and "D" the blade is cooled on both sides. Further noticeable differences are visible in figure 13. In the rear part of the blade at 90% span the solid is cooled. The strong temperature gradient at the suction side arises from a jet of high kinetic energy in between the suction side and a separation bubble. The much thicker hub area at 10% span (line "F") is bi-directionally heated. However, the temperature gradient in the solid between pressure and suction side is negligible. Even in the thick hub region the differences are less than 3K. In the thin tip region, the high thermal conductivity leads to a gradient of less than 1K. Under consideration of varying wall thicknesses and the scallop shaped back with an undercut at the shaft, the importance of a transient temperature calculation is evident.

242

high thermal load on PS

,

.. /

./

895~

-850

equalized ,./'" "T-distribution in the vanes

~

'.'~~ ~.

,"

~

BCD

A

Figure 10 Temperature distribution at midspan

A

A B C

D E F

%span %stream 50 10 50 40 50 55 50 70 90 55 10 55

Figure 11 Locations of temperature profiles in fig. 12, 13

93°Tr~~===i----------~~----------:

920 910 900 890 ..: 880 870 860 850

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-~70%stream

g

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40

50

70

60

80

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Figure 12 Static temperature at 50% spanwise (profiles A, B, C, D)

g

930 920 910 900 890

E: 90% span - 0 : 50% span

I

F: 10% span

E

PSr;SS I

--1, ..::;
~eiwr

(12)

where,tI> = ['P1'P2 A 'PN]is a matrix formed by the tuned modes and ~is the weight vector. An assumption has been made here that the mistuned vibration is still within the sub-space spanned by these tuned modes. Substituting Eq. (12) into (11) and premultiplyingtl>H , yields - w 2 (M* + tl>H ~MtI»~ + (K* + tl>H ~KtI»~ = tl>H Q (13) where M* and K *are diagonal matrices formed by the modal mass and stiffness of the tuned system. If mass and stiffness matrices mistuning (i.e. tl>H ~MtI>andtl>H ~KtI» is known, then forced vibration can be determined by solving this equation. To identify mistuning, use is made of the tip-to-tip frequency response matrix, H(w) , defined above. Define the force vector of an order excitation: F(t)=(F;,F2 ,A FNfe iwr = Fe iwr (14) in which the forces are applied at the blade tips and normal to the blade surfaces. The global force vector, Q, in Eq. (13) can be generated accordingly. The normal displacement amplitudes of the tips due to this order excitation can be calculated using qe = HF (15)

qe = ({g, q; ,K q~)T is the vector of the normal tip displacement amplitudes. According to Eq. (12), the tip displacement vector, (us, v s' Ws f of the sth blade, is

where,

given by (16) where, !PI(~' !P~~ and !pj~ are the x-, y- and z-components of the tip displacement of the sth blade in the mth tuned mode. If the directional cosines vector of the tip normal is denoted byas ' then from Eq. (16) (17)

In Eq. (17), the term on the left hand side is

q:. Denoting

a sm =a:(!PI :;; 0.8 t···········,·····

q;

0::

0.4

0.45

0.5

0.55

0.6

0.65

0.7

0.75

0.8

0.85

0.9

0.95

Rela1ive corrected mass flow rate

Fig. 8 Comparison of measured turbine performances at const. Pressure ratios

325

1.05

15

Fig. 9 Set up of laser scanning vibrometer

(a) 2nd mode (b) 4th mode Fig. 10 Mode shapes of blade second and fourth vibratory modes (Red color has the smallest displacement and blue color the largest displacement)

326

50000 45000 40000 . .35000 J:

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I

I

I I

-+- 11 -. - 12 13 -+- 14 -e- 15

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327

Compact Long-Route Exhaust Gas Recirculation Mixer Design and Optimization lunfei Yin, Nicolas Deschatrettes, Ocean Han, Philippe Renaud Honeywell Turbo Technologies Ltd., Skelmersdale, UK NOTATIONS d ---- Gas pipe diameter D ---- Main pipe diameter Pt ---- Total pressure Tt ---- Total temperature U ---- Velocity W ---- Mass flow rate 'Y ---- Uniformity Index A ---- Total pressure loss coefficient p ---- Density Subscripts 1 ---- Fresh air inlet 2 ---- Hot gas inlet SYNOPSIS

Reducing engine emissions becomes a very challenging target in automotive industries. Long-route exhaust gas recirculation (EGR) is considered as one potential solution to reduce emissions. The hot-gas is injected back into the turbocharger compressor wheel inlet. The hot gas will cause strong distortion in temperature and pressure fields. From the aerodynamics point of view, that will result in compressor performance loss and potential wheel overheating. In this paper, the design and CFD-optimization of long-route EGR mixers are discussed. Several design configurations were investigated. One selected design was optimized by changing different design parameters to achieve minimum loss and maximum uniformity of compressor inlet temperature. INTRODUCTION

Exhaust gas recirculation (EGR) has been used in recent years to reduce NOx emissions in light duty diesel engines. EGR involves diverting a fraction of the exhaust gas into the intake manifold where the recirculated exhaust gas mix with the incoming air before being inducted into the combustion chamber. EGR reduces NOx because it dilutes the intake charge and lowers the combustion temperature [1]. The US environmental protection agency has reviewed a variety of ways to implement EGR [2]. For passenger vehicles, many systems are under development. A long-route EGR mixer was designed in [3], but the configuration is not suitable for use with a turbocharger compressor. There are two EGR solutions based on such concept [4]:

329

(1) Long route system shown in figure 1: In a long route system, the pressure drops across the air intake and the stagnation pressure in the exhaust gas stream makes the EGR possible. Compressor Aftercooler

Figure 1 Long-route EGR layout (a filter not presented)

Compressor

Cooler

Turbine Figure 2 Short-route EGR layout

(2) Short route system shown in figure 2: These systems differ mainly in the method used to set up a positive pressure difference across the EGR circuit. Another way of controlling the EGR-rate is to use variable nozzle turbine (VNT). Most of the VNT systems have single entrance, which reduce the efficiency of the system by exhaust pulse separation. Cooled EGR should be supplied effectively. Lundquist[5] and others used a variable venturi, in which EGR-injector was allowed to move axially, thus varying the critical area. In the short-route EGR, exhaust gas is taken from after the exhaust valve and passed through a valve to the inlet of the engine. The long-route or low-pressure EGR, can also be used as a retrofit solution. The NOx reduction can be up to about 30-40%. In this case, particulate free exhaust gas is led from after a particulate filter (not shown here), then is cooled and introduced before entering the turbo. The hot-gas is injected back into the compressor wheel inlet in the turbocharger. The hot gas could cause strong temperature and pressure distortion at the compressor wheel inlet. From aerodynamic point of view, 330

that will result in compressor performance loss and potential wheel overheating. The following requirements are essential for an EGR mixer: Compact design due to limited space in a passenger car (the ratio of mixing length to pipe diameter is less than 1.0 ) Minimum flow restriction during the intake process, i.e. minimum (total) pressure loss from air inlet to compressor wheel inlet. Minimum total pressure loss from EGR hot gas inlet to compressor wheel inlet. In the EGR circuit, the loss is related to turbine power An acceptable mixing quality. The objective of this work is to design and optimize long-route EGR mixers in order to achieve the essential requirements. The CFD tool FLUENT was used to evaluate different designs. BASIC DESCRIPTION AND PERFORMANCE PARAMETERS Basic mixer model and conditions

The gas conditions for the fresh air inlet are ambient conditions: ambient pressure, ambient temperature; the mass-flow rate is 85% of the total mass-flow rate. The recirculation gas has a total Temperature of 500 Kelvin, and represents 15% of the total mass-flow rate. Two families of variables are defined to characterise the mixing process: Pressure loss coefficient, and Uniformity index. They are defined as follows. From the hot gas inlet to the outlet: /..(gas)= (Pt(gas inleW P1(outlet average»)/

{0.5 pU2 air inlet}

From the fresh air inlet to the outlet: /..(air)= (Pt(air in1et)~P1(ut1et average»)/ {0.5pU 2 i1ir inlet}

Uniformity Index y: y is the flow distribution index which is 1.0 for a uniform flow[6], and decreases with decreasing flow uniformity.

331

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Where Ai is an elemental area and Vi can be Pt (Total Pressure) and Tt (Total Temperature) to calculate the index 'Y (Pt) or'Y (Tt). The higher the 'Y parameter, the more uniform the flows. 'Y gamma uniformity index is an integral measure of flow uniformity. These metrics can be used to define specific "pass/fail" criteria; the current long-route EGR design will fail if any of the indices falls outside a preset range. While these measures can be determined in FLUENT, it is preferable to have a post-processing utility which automates the process. A code has been written for this purpose. The flow uniformity characteristics of the long-route EGR mixer can therefore be quickly and easily determined. Models

Three cases have been examined, each with d/D=0.5. Case 01 is the simplest model in which there is a main flow pipe, the fresh air flows in and then mixes with the hot gas flow. The gas pipe is normal to the main pipe with a small diameter. In Case 02, there are a number of holes on the main pipe wall and the hole positions and areas are well distributed in order to achieve better mixing and smaller loss. The total area for the small holes are equal to the hot-gas pipe area. Case 03 has a Circumferential Slot on the main pipe wall; the highest and smaller slot widths are 32% and 10% of the hot-gas pipe diameter. CFD RESULTS AND ANALYSIS FOR THE STARTING MODELS

To take the best result from the starting models, the refined designs were made to optimise the number, area, position of the holes and other design parameters. The results on the three stations: 0.5D, ID and 2D were taken downstream from the centre of the hot gas pipe in figure 3, where D is the pipe diameter of the main pipe. The 332

main focus is on the 0.5D station, which represents the shortest EGR mixer and a compact design (the mixing length is.0.5D).

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