An on-engine twin-scroll turbineperformance estimation

Viktor Olsson

Master of Science Thesis MMK2015:70 MFM161KTH Industrial Engineering and Management

Machine DesignSE-100 44 STOCKHOLM

.

Master of science thesis MMK2015:70 MFM161

An on-engine twin-scroll turbine performanceestimation

Viktor Olsson

Approved: Examiner: Supervisor:

2015-06-22 Andreas Cronhjort Ted HolmbergCommissioner: Contact person:

Scania CV AB Ph.D. Oskar Leufven.

Abstract.In this study the main objective was to estimate the instantaneous turbine shaftpower and turbine efficiency of a turbocharger mounted on a 6-cylinder 13 literheavy-duty diesel engine.

The work was carried out at Scania CV AB in Sodertalje, Sweden fromJanuary to June 2015 as master thesis under the division of Internal CombustionEngines at the Royal Institute of Technology KTH, Stockholm, Sweden.

Normally the turbine performance is estimated during off-engine conditionsin a gas flow bench during steady flow. When the turbocharger is mountedon the engine, the flow is far from being steady. The exhaust gas flow is pul-sating and every thermodynamic property with it as well. In this thesis theturbine performance is estimated 1 crank angle degree resolved with 1 crankangle degree resolved pressures and turbocharger speed. Parameters difficult tomeasure crank angle resolved such as temperature and exhaust mass flow rateare estimated utilizing methods found in literature.

The turbine efficiency is defined as the utilized turbine shaft power dividedwith the isentropic turbine power. The utilized shaft power consists of thepower consumed by the compressor, which is assumed to be constant in thisthesis, the power lost due to bearing friction and the acceleration power of therotating parts. The result of the used method revealed a fluctuating turbine ef-ficiency, very seldom at the same levels as the optimal efficiency obtained duringsteady flow during one engine cycle. The results where promising, but lackedthe needed accuracy to estimate the instantaneous turbine performance withproven confidence. Primarily this is believed to be caused by the instantaneousexhaust mass flow rate approximation based on a calculated dynamical pres-sure. Another factor impacting the overall accuracy is the assumed constantcompressor power. Further development of the method is needed and couldyield results with better confidence in the future.

i

.

Examensarbete MMK2015:70 MFM161

An on-engine twin-scroll turbine performanceestimation

Viktor Olsson

Godkant: Examinator: Handledare:

2015-06-22 Andreas Cronhjort Ted HolmbergUppdragsgivare: Kontaktperson:

Scania CV AB Ph.D. Oskar Leufven.

Sammanfattning.Denna rapport handlar om skattning av momentan turbinaxeleffekt och turbin-verkningsgrad for ett turboaggregat monterat pa en 6-cylindrig 13 liter stordieselmotor.

Arbetet gjordes pa och for Scania CV AB i Sodertalje, fran Januari till Juni2015, som ett examensarbete under institutionen Forbranningsmotorteknik paKungliga Tekniska Hogskolan, Stockholm.

Nar turbinprestandan normalt bestamms, sker detta i en s.k. flodesbank darturbinen matas med ett konstant flode. Nar turbon ar monterad pa en motor serflodessituationen annorlunda ut, avgasflodet och dess termodynamiska tillstandpulserar. I den har rapporten skattas turbinprestandan vevvinkelupplost medhjalp av vevvinkelupplosta tryck och turbovarvtal. Parametrarnastintill omojliga att mata vevinkeluppost sasom temperatur ochavgasmassflode har skattats vevvinkelupplost med metoder funna i litteratur.

Turbinverkningsgraden bestar av den utnyttjade axeleffekten dividerad medden isentropa turbineffekten, dar den utnyttjade axeleffekten bestar av kom-pressoreffekten vilken betraktats som konstant, en lagerfriktionsforlust samt enaccelerationseffekt av de roterande delarna i turbon. Resultatet visade att turbi-neeffekten och turbinverkningsgraden fluktuerade, som vantat med fluktuerandeingangsparmetrar och att verkningsgraden sallan ar den optimala framtagen un-der konstant flode. Den anvanda metodiken visade sig ge ett lovande resultat,men saknade tillracklig noggranhet for att bestamma turbinverkningsgradenmed stort fortroende over en hel motorcykel. Detta tros primart vara en foljdav det uppskattade momentana avgasmassflodet baserat pa ett dynamiskt tryckoch sekundart en antaganen konstant kompressoreffekt. En metodutvecklingskulle kunna ge battre resultat i eventuella fortsatta studier.

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Nomenclature

mair Air mass flow rate [kgs ]

mexh. Exhaust mass flow rate [kgs ]

ηmech. Mechanical efficiency losses due to bearing friction [-]

ηturb. Turbine efficiency [-]

UCs, BSR Blade speed ratio [-]

γexh. Ratio of specific heat of exhaust [-]

ρexh. Density of exhaust gas [ kgm3 ]

At Cross section area of one scroll at turbine inlet [m2]

cp,air Specific heat capacity of air [ JkgK ]

cp,exh. Specific heat capacity of exhaust [ JkgK ]

Jrotor Rotor assembly polar moment of inertia [kgm2]

Nrotor Turbocharger rotational speed [rpm]

Pacc. Rotor assembly acceleration power [W]

Pcomp. Power consumed by compressor [W]

pcyl6 Absolute pressure cylinder 6 [Pa]

pdynamic Dynamic pressure [Pa]

pexh.s Static pressure exhaust manifold [Pa]

pin,t Total pressure at turbine inlet [Pa]

Pisentropic Isentropic total-to-static turbine power [W]

pout,s Static pressure at turbine outlet [Pa]

ps,inst. Instantaneous static pressure at turbine inlet [Pa]

ps,mean Mean static pressure at turbine inlet [Pa]

Putilized Utilized turbine power [W]

Tin,comp. Temperature at compressor inlet [K]

Tin,exh.mean Mean temperature at turbine inlet [K]

Tin,exh. Temperature at turbine inlet [K]

Tinst.exh. Instantaneous temperature at turbine inlet [K]

Tout,comp. Temperature at compressor outlet [K]

u Flow velocity of exhaust [ms ]

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Contents

1 Introduction 11.1 Turbocharging . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.2 Turbocharger layout . . . . . . . . . . . . . . . . . . . . . . . . . 1

1.2.1 Turbine . . . . . . . . . . . . . . . . . . . . . . . . . . . . 21.2.2 Compressor . . . . . . . . . . . . . . . . . . . . . . . . . . 21.2.3 Boost-pressure control . . . . . . . . . . . . . . . . . . . . 31.2.4 Pulsating exhaust flow . . . . . . . . . . . . . . . . . . . . 3

2 Objective 52.1 Problem statement . . . . . . . . . . . . . . . . . . . . . . . . . . 52.2 Delimitations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5

3 Theory 63.1 Unsteady turbine flow . . . . . . . . . . . . . . . . . . . . . . . . 63.2 Turbine power . . . . . . . . . . . . . . . . . . . . . . . . . . . . 73.3 Utilized power . . . . . . . . . . . . . . . . . . . . . . . . . . . . 83.4 Turbine efficiency . . . . . . . . . . . . . . . . . . . . . . . . . . . 93.5 Blade speed ratio . . . . . . . . . . . . . . . . . . . . . . . . . . . 9

4 Methodology 104.1 Experimental setup . . . . . . . . . . . . . . . . . . . . . . . . . . 10

4.1.1 Indication system . . . . . . . . . . . . . . . . . . . . . . . 104.1.2 Instrumentation, measured and estimated quantities . . . 10

4.2 Measuring procedure . . . . . . . . . . . . . . . . . . . . . . . . . 13

5 CAD resolved signal evaluation 155.1 Reference points . . . . . . . . . . . . . . . . . . . . . . . . . . . 155.2 Turbocharger rotational speed . . . . . . . . . . . . . . . . . . . . 16

5.2.1 Digital Micro-Epsilon signal . . . . . . . . . . . . . . . . . 165.2.2 Analog Micro-Epsilon signal . . . . . . . . . . . . . . . . . 17

5.3 Turbocharger rotational acceleration . . . . . . . . . . . . . . . . 185.4 Pressure signals . . . . . . . . . . . . . . . . . . . . . . . . . . . . 20

5.4.1 Total pressure signals . . . . . . . . . . . . . . . . . . . . 205.4.2 Static pressure signals . . . . . . . . . . . . . . . . . . . . 22

5.5 Exhaust mass flow rate estimated from dynamic pressure . . . . 245.5.1 Discussion of accuracy and validity . . . . . . . . . . . . . 29

5.6 Exhaust pulse identification . . . . . . . . . . . . . . . . . . . . . 30

6 Results 326.1 Isentropic turbine power . . . . . . . . . . . . . . . . . . . . . . . 326.2 Utilized power . . . . . . . . . . . . . . . . . . . . . . . . . . . . 346.3 Turbine efficiency . . . . . . . . . . . . . . . . . . . . . . . . . . . 356.4 Flow characterization . . . . . . . . . . . . . . . . . . . . . . . . 406.5 Comments on results & accuracy . . . . . . . . . . . . . . . . . . 42

7 Summary, conclusions and future work 437.1 Conclusions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 437.2 Future work . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 43

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1. Introduction

In this chapter the turbocharger concept is explained, its purpose and way ofoperation are covered.

1.1 Turbocharging

Turbochargers are an effective way to meet todays high legislations regardingparticulate and NOx emissions for heavy duty diesel engines. The turbochargeras a concept is used to increase power and to lower the fuel consumption andemissions, and is found on almost every diesel engine sold today.

The implementation of turbochargers begun in the 1950s and with it camea new set of challenges, increased pressure in the combustion chamber amongothers. The necessity to overcome these problems has been the main focus ofturbocharger development in the earlier years and has forged the turbochargerconcept towards current technology.

Different turbocharger concepts exist, the most common are mechanical su-percharging and exhaust gas turbocharging [1]. Mechanical supercharging isnot covered in this report.

An exhaust gas turbocharger consists of a turbine connected via a shaft to acompressor. The turbine is connected to the exhaust manifold and the turbinewheel is driven by the exhaust gases leaving the cylinder at the end of theexpansion stroke. Thus the power extracted from the exhaust gases drive thecompressor, the compressor in turn increases the pressure of the intake air andin combination with a charge air cooler the density is increased as well beforeentering the combustion chamber. The higher pressure and higher density allowsmore fuel to be burned efficiently per cycle, which increases the power outputof the engine. The focus of this report is on exhaust gas turbocharging, referredto as just turbocharging in the rest of the report.

1.2 Turbocharger layout

The turbocharger consists of three major parts, a turbine, compressor andbearing-housing shown in Figure 1.1. In the bearing housing, the turbine-compressor shaft is supported by a bearing system that is crucial for the overall function of the turbocharger. The bearing system controls both the ax-ial and radial movement of the shaft assembly (turbine and compressor wheelsmounted on the connecting shaft). Typically journal bearings are implementedbut applications with ball bearings exist as well.

Other key components in the bearing-housing are seals, located at both theturbine and compressor side. The seals main purpose is to keep exhaust gasesand intake air out of the bearing housing, since the pressures in the turbine andcompressor are typically higher than the pressure inside the bearing-housingwhich is mostly the same as the pressure inside the engine crankcase. Thuspreventing oil from leaking out into the turbine and compressor is not the sealsmain purpose, this is typically handled by oil deflectors before the oil evencontacts the seals. One problem with the seals in general is that they need tohandle a shaft movement caused by a needed bearing clearance, this often affectthe total friction loss in the turbocharger [2].

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Figure 1.1: Cut through of a twin-scroll radial flow turbocharger, showing thecompressor on the left and the turbine on the right.

1.2.1 Turbine

The purpose of the turbine is to extract energy from the exhaust flow usedto drive the compressor. The turbine stage is composed of two parts, a rotor(turbine wheel) and a stator (turbine housing). Different types of turbines exist,and not all are suited for exhaust gas turbocharging. The three best suited are,axial-flow, radial-flow and mixed-flow turbines [3], since the main focus in thisreport is on the radial-flow turbine the other concepts are left unexplained.

The radial-flow turbine can be configured in different ways, either as a single-entry turbine, a double-entry or twin-entry turbine. The different configurationsinfluence the dynamic behavior of the turbocharger, explained later in this sec-tion. The purpose of the turbine inlet is to guide the flow with a minimal lossin total pressure so that a torque can be conveyed to the rotor blades. Thisis achieved by the scroll shape of the housing called volute. In some casesthe redirection of the flow can be aided by small curved airfoils referred to asvanes in the stator, but frequently the radial-flow turbines used for exhaust gasturbocharger are vaneless.

1.2.2 Compressor

The compressor stage consists of two parts, a compressor housing often calleda cover and a impeller (compressor wheel). The flow through the compressorenters the center of the cover in an axial direction to the compressor wheel, theflow through the compressor wheel changes the flow direction by 90◦. When theflow exits the compressor wheel, it is changed from an axial to a radial direction.The radial flow is directed through a diffuser converting the kinetic energy ofthe air into pressure by diffusing the flow velocity. The diffused air enters a

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plenum often referred to as a volute, either assisted by vanes or vaneless. Theflow is redirected through the volute and is discharged from the compressor.

The size and shape of the compressor wheel highly impact its performance,and the most critical parts are the compressor blades. The blade profile overthe wheel and the wheel tolerance against the housing are two very importantperformance factors along with the size of the blades. A trade-off between per-formance at different operating conditions and manufacturing aspects is alwaysmade.

1.2.3 Boost-pressure control

The amount of pressure above atmospheric pressure produced by the compressoris often referred to as boost-pressure, too high boost-pressure is not desirablesince it can brake the engine and cause the turbocharger to overspeed, eventuallycausing it to fail.

The amount of boost-pressure is in a diesel engine application often con-trolled by reducing the amount of exhaust flow over the turbine wheel. Theflow can be regulated in different ways, one option is to implement a wastegate.A wastegate is a type of valve often controlled pneumatically which allows a cer-tain amount of excess exhaust gas to bypass the turbine wheel, thus controllingthe amount of extracted energy from the exhaust flow. The wastegate actuatoris often mounted on the compressor side and connected to the valve by a rod,due to the high thermal load on the turbine side of the turbocharger.

Another way of controlling the the boost-pressure is to implement a vari-able geometry turbine (VGT). This specific type of turbine, as compared to awastegated turbine, allows the turbocharger to work more efficiently within thetotal operating range [4]. In the VGT application the entire exhaust mass flowin controlled within the turbine and no flow is bypassed. This is achieved byvarying the turbine cross-sections and thus the turbine’s resistance to flow canbe regulated against the required boost-pressure level. Implemented designstypically feature adjustable vanes, either by controlling the vane angles or thearea of the vanes, allowing a wide control range with high efficiency.

In some applications turbines are fitted with both a wastegate and a VGT,though most often the systems are used separately. The turbocharger coveredin this report is fitted with a wastegate and have a fixed geometry.

1.2.4 Pulsating exhaust flow

The turbocharger is coupled to the engine thermodynamically and not mechan-ically. The enthalpy of the exhaust gas is the source from which the mechanicalenergy used to drive the compressor is converted. Thus the exhaust gas/turbineinteraction is very important in order to achieve as high turbine efficiency aspossible.

The exhaust flow from a 4-stroke engine, whether its spark ignited or com-pression ignited, is highly pulsating. The dynamic operation of a turbochargeris highly affected by these pulsations. Previously the pulsations where oftendampened in order to make the pressure at the turbine inlet as constant aspossible. This dampened effect can be achieved by increasing the volume ofthe exhaust manifold and configuring the turbine with a single-entry housing,

3

shown in Figure 1.2. This method, referred to as constant pressure turbocharg-ing ensures a continuous flow over the turbine wheel, but doesn’t maximize useof the available exhaust gas energy due to the dampening of pressure peaks.

Turbine configurations that utilize the fluctuating behavior of the exhaustflow tend to have turbines with multiple inlets. The exhaust manifolds feedingthe multiple inlets are divided, and depending on the engine configuration theycarry the exhaust from a couple of cylinders each. The most common turbineconfiguration is the twin-entry turbine [4], often referred to as a twin-scrollturbine shown in Figure 1.2. The volute is meridionally divided into two scrolls.In a 6-cylinder engine configuration each scroll is fed by three cylinders and in a4-cylinder engine each scroll is fed by two cylinders etc. One of the advantageswith the scroll separation, besides the maximized exhaust energy utilization, isthe improved gas exchange by avoiding overlap of the exhaust pressure pulses [5].

Figure 1.2: Comparison between a twin-entry turbine on the left and a single-entry turbine on the right.

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2. Objective

In this chapter the objective and limitations of the thesis are explained.

2.1 Problem statement

The objective with this thesis carried out for Scania CV AB in Sodertalje, Swe-den is to estimate the on-engine instantaneous turbine power and instantaneousturbine efficiency.

The knowledge of how the turbine performs on-engine is crucial when op-timizing the engine performance and fuel consumption, and understanding ofthe flow process trough the turbine is a key factor in the turbocharger-enginematching.

Compared to standard turbine performance, most often measured in off-engine conditions during steady state flow, the on-engine performance behavesdifferently in terms of the flow diverging from being steady. The assignment willbe to estimate the turbine performance during the pulsating flow that occurswhen the turbocharger operates on the engine. The aim of an initial attemptis that the results will raise conclusions and further development regarding themeasurement setup, i.e. types of sensors used/needed, sensor positions andindication system requirements. Another result could be a better understandingof how the turbocharger behaves as a dynamical system, how the energy is storedand utilized in the turbine at different operating conditions the turbocharger isexposed to during on-engine operation.

2.2 Delimitations

The turbine performance estimation is limited to operating conditions wherethe wastegate is closed, due to the wastegate flow being unknown. The onlyevaluated points are static in the sense of constant engine speed and constantengine load, no turbine performance was estimated during engine transientsbetween different engine loads or engine speeds.

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3. Theory

In this chapter the underlying theory and governing equations for calculating theon-engine instantaneous turbine performance in terms of power and efficiencyis explained. The validity of the theory during on-engine operation will bediscussed based on methods and conclusions found in literature.

3.1 Unsteady turbine flow

Turbine performance calculations are usually based on turbine maps producedby the turbo manufacturer. These standard maps are produced under steady-state conditions in flow benches and present turbine efficiency and speed as afunction of mass flow rate and pressure ratio. As mentioned before, when theturbocharger is mounted on the engine the exhaust flow is pulsating and themass flow rate and pressure ratio is far from being steady, so the validity of thesesteady-flow performance maps during unsteady pulsating flow is questionable.

In order to utilize the steady-flow performance maps during pulsating flow,the standard method is to assume the flow as quasi-steady (QS). The QS as-sumption entails that flow behaves at any instant as it would during steady-flow [6], thus the mass flow rate and efficiency can be obtained from the steady-flow maps at any instant. This approach is implemented by several researchers,Benson and Scrimshaw [7] were one of the first to try and validate a QS be-havior of the turbine, many has since followed. Kosuge et al. [8] comparedinstantaneously measured data during pulsating flow with steady-state data ina gas flow bench to try and quantify the validity of the QS assumption. Theproblem in this early investigation is that instantaneously measured data aretime-averaged due to the absence of fast response transducers. Despite this theQS approach was found to underestimate the turbine power.

In order to fully clarify the effect of pulsating exhaust flow on turbine op-eration acquired data must have good enough resolution. Martinez-Botas andcolleagues [9–12] measured highly resolved data in order to find the limitationsof the QS approach. Some insights involved the wave dynamics of the pressurepulse traveling trough the turbine, and how this effect the filling and emptyingof the turbine volute, others revealed the frequency of the pulsating flow asa major parameter affecting the dynamical behavior of the turbine. The gen-eral conclusion was that the turbine performance during unsteady-flow divergedfrom QS flow performance.

Szymko et al. [13] later tried to assess the applicability of the QS assumptionand categorize its validity during different levels of unsteady flow, and concludedthat it is more valid at higher pulsation frequencies but still underestimates theunsteady-flow turbine power.

Even though the QS approach can show rough trends of the instantaneousunsteady-flow turbine performance its not suited for the whole operating rangeof the turbine, thus the steady-state performance maps do not fully describe theturbine behavior during unsteady-flow.

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3.2 Turbine power

The turbine power is often defined from the isentropic expansion of the exhaustgas over the turbine. Since no work can be extracted from the kinetic energyleaving the turbine, the available power is defined as the difference in enthalpybetween a stagnation inlet condition and a static outlet condition [6]. So, theisentropic total-to-static turbine power is the difference between power flowinginto the turbine and out of the turbine:

Pisentropic = mexh.in

Tin∫0

cp,exh.dT − mexh.out

Tout∫0

cp,exh.dT (3.1)

Ideally the isentropic power should be the sum av all gas molecules directlyaffecting the turbine rotor blades, but the actual flow and thermodynamic prop-erties around the turbine wheel and its close proximity are unknown. The poweris thus defined as the difference between two boundary conditions at the turbineinlet and outlet where the thermodynamical properties are easier to obtain. Thisentails a control volume between the boundaries. Since the control volume overthe rotor where the actual power is extracted is smaller than the volume betweenthe boundaries, equation 3.1 is simplified by assuming that mexh.in = mexh.out

and that cp,exh. is constant over the volume, and with isentropic conditions it’srewritten as:

Pisentropic = mexh.cp,exh.Tin,exh.

1 −(pout,spin,t

) γexh.−1

γexh.

(3.2)

As mentioned before filling and emptying of the turbine volute occurs duringpulsating exhaust flow, this causes energy storage in the turbine. Westin [14]attempted to simulate this energy storage and concluded that the most appro-priate assumption when conducting on-engine measurements is to assume thatno energy is stored within the volute, why mexh. in equation 3.2 is the exhaustmass flow rate through the turbine. Equation 3.2 is only valid when the waste-gate is closed and no bypass of exhaust flow occurs in this thesis, explained inchapter 5.

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3.3 Utilized power

The available isentropic power utilized or extracted consists of three terms [14],the power absorbed by the compressor, the power used to accelerate and decel-erate the rotor assembly in conjunction with the pulsating exhaust flow and thepower lost in the bearing system due to friction.

Putilized = Pacc. +Pcomp.ηmech.

(3.3)

The power used to accelerate and decelerate the rotor assembly is derivedfrom the torque of the fluctuating speed changes to the inertial power inequation 3.4

Pacc. =

(2π

60

)2

JrotorNrotordNrotordt

(3.4)

The power consumed by the compressor to compress the intake air is oftenbelieved to be constant during unsteady flow and is then defined according toequation 3.5, where all quantities are average values and not instantaneous.

Pcomp. = maircp,air (Tout,comp. − Tin,comp.) (3.5)

The assumption that the compressor power is constant is debatable, Westin[14] showed that the fluctuation amplitude of the consumed compressor power isin the range of 20-30% on a 4 cylinder SI-engine. Winkler et al. [15] performeda similar study but on a 6-cylinder heavy duty diesel engine and concludedthat the compressor power fluctuations where significantly lower. For this the-sis the compressor power is assumed to be constant, the main reason is due tothe fact that a time lag will be introduced between the calculated performanceparameters. The compressor power will most certainly be out of phase in timecompared to both the isentropic turbine power and acceleration power. Anotherfactor contributing to a mean calculated compressor power is the fact that cer-tain quantities are very hard to measure instantaneously, explained in detail inchapter 4.

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3.4 Turbine efficiency

The isentropic turbine power and utilized power can be used to calculate theturbine efficiency according to equation 3.6.

ηturb. =

(2π60

)2JrotorNrotor

dNrotordt +

maircp,air(Tout,comp.−Tin,comp.)ηmech.

mexh.cp,exh.Tin,exh.

[1 −

(pout,spin,t

) γexh.−1

γexh.

] (3.6)

This simplified efficiency in terms of constant compressor power is not farfrom reality according to Lujan et al. [16] among others. This is motivated byLujan et al. based on the fact that pressure pulsations within the engine inletsystem after the compressor are not important. In order to do a instantaneousestimation of the turbine efficiency, all quantities need to be instantaneouslyobtained, limitations explained more in detail in chapter 4.

3.5 Blade speed ratio

To be able to classify the working conditions of the turbine, whether the flow canbe described with a steady state point or not the Blade Speed Ratio (BSR) orU/Cs is often calculated. The BSR is the turbine wheel blade tip speed dividedby the velocity equivalent of an isentropic enthalpy drop over the turbine [1]according to equation 3.7.

BSR =U

Cs=

U√√√√2cp,exh.Tin,exh.

[1 −

(pout,spin,t

) γexh.−1

γexh.

] (3.7)

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4. Methodology

In this chapter the details regarding all data required for calculating the turbineperformance in terms of power and efficiency are explained. First the measur-ing equipment, sensors and the data acquisition systems and their limitationsare covered. Second, the details regarding measurement locations on the tur-bocharger and corresponding limitations are presented and explained. Last, themeasuring procedure and operating points are presented.

4.1 Experimental setup

The measurements were carried out on a standard production heavy-duty 13liter in-line 6-cylinder diesel engine, fitted with a twin-entry wastegated sym-metrical turbine scroll turbocharger.

4.1.1 Indication system

In order to estimate instantaneous turbine performance, fast measurements haveto be done. Not all parameters are possible to measure instantaneously, due toharsh working conditions for sensors and the lack of fast enough sensors. Forthis reason two different indication systems where used, one system to acquireinstantaneously measured signals (both analogue and digital inputs) and onesystem for slow analogue signals.

The fast indication system consist of an AVL IndiModul [17]. The IndiModulanalogue part has an input range of ±10 V and a maximum sampling rate of800 kHz per channel, it has a built in 100 kHz low-pass filter and a 14 bitanalog-digital converter. The IndiModul digital input counts pulses and themeasurement resolution was set to 1 CAD (Crank Angle Degree), considered tobe instantaneous prior to the measurements.

The slow analog indication system used for thermocouples, air mass flowrate, i.e. for most signals impossible to acquire instantaneously was set with asampling frequency of 10 Hz.

4.1.2 Instrumentation, measured and estimated quanti-ties

To evaluate the instantaneous turbine performance 1 CAD resolved data arerequired, for this reason fast response sensors were implemented. Not all quan-tities are possible to measure instantaneously, due to several reasons explainedfurther in this section.

The sensor installation is depicted in Figure 4.1 and summarized in table4.1.

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Figure 4.1: Measurement locations on the turbocharger and engine.

Table 4.1: Description of measurement locations.No. Name Description Sensor unit

1.pin,sTin,exh.

Static pressure at outer scroll inletTemperature at outer scroll inlet

bar◦C

2. pout,s Static pressure at turbine outlet bar3. pin,t Total pressure at outer scroll inlet bar4. pin,t Total pressure at inner scroll inlet bar5. Tout Temperature at compressor outlet ◦C

6.pin,sTin,exh.

Static pressure at inner scroll inletTemperature at inner scroll inlet

bar◦C

7. Tin Temperature at compressor inlet ◦C8. Nrotor Turbocharger speed rpm9. pexh.s Static pressure exhaust manifold bar10. pcyl6 Absolute pressure cylinder 6 bar

Both static and total pressures where measured each CAD with water cooledpiezo-resistive absolute pressure sensors from Kistler (type 4049B10DS) [18],with a pressure range of 0-10 bar. Inlet conditions where measured in each ofthe two scrolls, results will be denoted with O for the outer scroll closest to theoutlet and I for the inner scroll closest to the bearing housing.

The static pressure sensors where flush mounted in order to achieve the mostundamped and fastest system possible. The total pressure sensors were mountedon pressure tubes similar to pitot tubes located approximately 110 mm from theturbine inlet flange at the start of the volute, visible in Figure 4.2.

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Figure 4.2: Total pressure tubes in turbine, the left picture shows the tubeslocated approximately 110 mm down in the scrolls from the flange and the rightpicture shows the sensor mounts on the outside of the turbine housing.

The cylinder pressure in cylinder 6 was measured with a piezo-electric sensorfrom AVL (type GU24D) [17], with a pressure range of 0-250 bar.

The turbocharger rotational speed was measured on the compressor sideusing a Micro-Epsilon turbo speed sensor [19]. The transducer is an eddy-current sensor drilled into the compressor housing sensing the compressor bladesas they pass by. The sensor signal is processed in a controller that outputs bothan analog and a digital signal. The analog signal range is between 0-5 V witha linear scale, 0 V equals 0 rpm and 5 V equals 200 000 rpm. The digitalsignal outputs TTL-pulses, either one pulse per passing blade or one pulse perrevolution, the later was chosen in this work.

The exhaust mass flow rate is one of the most difficult quantity to measureinstantaneously, the high exhaust gas temperatures and the pulsating exhaustflow make it impossible to use standard hot film or hot wire mass flow rate sen-sors. The turbine performance is usually estimated in a flow rig with pulsatingexhaust flow, the lower temperatures in the flow rig allows the mass flow rateto be measured. Winkler et al. [15] calculated the instantaneous on-engine tur-bine performance using a simulated mass flow rate, Westin [14] used simulatedcrank-angle resolved data as well. Another approach is taken in this report, themass flow rate was estimated using the dynamic pressure calculated from staticand total pressures measured at the turbine inlet. The approach is explained indetail in chapter 5.

The compressor inlet air mass flow was measured in the engine test cell usinga hot film sensor.

Temperature is another parameter hard to measure instantaneously. Due tothe slow response of thermocouples, only average temperatures were measured.The thermocouple implemented was a standard 4 mm K-type thermocouple,both on the turbine and the compressor side. The inlet temperature and out-let temperature on the compressor where left as average temperatures sincethe compressor power was assumed to be constant. To estimate the instanta-neous turbine performance, the turbine inlet temperatures needs to be measuredinstantaneously. Since this is not possible, the instantaneous turbine tempera-tures where estimated from the measured mean values. This method based onan adiabatic compression or expansion is explained in chapter 5.

12

The rest of the parameters, cp,exh. and cp,air where calculated for the corre-sponding gas composition and temperature in the engine test cell as standardparameters. The same applies for γexh..

4.2 Measuring procedure

In this section the measurement procedure is described, both CAD-resolved(considered instantaneous) and time-averaged measurements have been carriedout.

The parameters measured instantaneously are static and total pressures,the turbocharger speed and the pressure in cylinder 6. As mentioned before thesignals are acquired every engine CAD, thus making the sampling frequencydependent of the engine speed. The acquisition time for each operating pointfor the instantaneously acquired signals was 100 engine cycles, where each cyclecorresponds to 720 CAD.

The rest non-instantaneously acquired signals where sampled at 10 Hz andthe acquisition time was 60 seconds for each operating point. The results arepresented as a mean over the acquisition time.

Two test series where conducted, one full load test and one part load testwith steady operating points to cover the whole operating range of the tur-bocharger. The full load test consisted of 8 operating points summarized intable 4.2. During the full load test the wastegate operation was active and thewastegate was open during the last three measurement points.

Table 4.2: Measurement points, full load test. Wastegate open for point F, Gand H.

PointEngine speed[rpm]

Engine load[Nm]

A 900 1767B 1000 2305C 1100 2349D 1200 2350E 1300 2357F 1500 2008G 1700 1779H 1900 1630

The part load test consisted of 25 operating points summarized in table4.3. To ensure that the wastegate stayed closed during this test series, it wasmechanically shut using a hose clamp.

13

Table 4.3: Measurement points, part load test. Wastegate mechanically shutfor all points.

PointEngine speed[rpm]

Engine Load[Nm]

PointEngine speed[rpm]

Engine Load[Nm]

1 935 450 14 1200 5592 935 872 15 1200 11833 935 1164 16 1200 15184 935 1457 17 1200 18545 935 1816 18 1300 5706 1000 549 19 1300 11527 1000 1170 20 1300 14538 1000 1520 21 1300 18329 1000 1858 22 1700 39010 1100 562 23 1700 71611 1100 1185 24 1700 91312 1100 1522 25 1700 126213 1100 1859

14

5. CAD resolved signal evaluation

In this chapter all measured signals are analyzed and discussed. Deviations andinteresting phenomena in the acquired data are discussed and the instantaneousparameters in need of calculation are calculated.

5.1 Reference points

To be able to present and compare acquired signals in a comprehensive yetperspicuous manner, three of the measurement points thought to represent threedifferent operating conditions of the turbocharger are chosen and discussed asreference points. Part load test point 3 and 24 are chosen, point 3 represents alow mass flow rate through the turbine due to the low engine speed and load.Point 24 represents a higher mass flow rate through the turbine. Full load testpoint C represents a very high mass flow rate through the turbine. To illustratethe different operating conditions, the test points for the part load test and fullload test are marked in Figure 5.1 where the engine load is plotted against theengine speed.

1000 1200 1400 1600 18000

500

1000

1500

2000

2500

Engine speed [rpm]

Eng

ine

load

[Nm

]

A

B C D E

F

GH

1

2

3

4

5

6

7

8

9

10

11

12

13

14

15

16

17

18

19

20

21

22

23

24

25

Full loadPart load

Figure 5.1: Full load and part load test points, chosen reference points markedwith black circles.

15

5.2 Turbocharger rotational speed

In this section the acquired turbocharger speed signals are discussed and eval-uated.

5.2.1 Digital Micro-Epsilon signal

The full load reference point equals an engine speed of 1100 rpm and loadof 2349 Nm. One engine cycle of the acquired digital turbo speed signal forthis point is visible in Figure 5.2. The exhaust flow fluctuations from all 6cylinders during one engine cycle are clearly visible as 6 distinct peaks in theturbocharger speed, with a peak-to-peak amplitude of approximately 300 rpm.The amplitude of the different peaks has a small variation, not strange since itis a thermodynamical system fluctuating. Other discrepancies are visible in thesignal, especially around the peaks. One possible cause for these discrepanciescould be the pulse rate of the sensor. At a mean turbocharger speed of 95600rpm and a engine speed of 1100 rpm one pulse is produced by the sensor roughlyevery 4.1 CAD. When the resolution in the fast indication system is set to 1CAD and the pulse rate is 4.1 CAD per pulse it is believed that the indicationsystem makes an interpolation to the desired resolution.

−400 −300 −200 −100 0 100 200 300 40095350

95400

95450

95500

95550

95600

95650

95700

95750

95800Full load point C

Crank angle [°]

Tur

boch

arge

r sp

eed

[rpm

]

Figure 5.2: Digital turbocharger speed signal for full load point C, one enginecycle. A clear pulsating behavior is visible with a peak-to-peak amplitude ofapproximately 300 rpm.

In Figure 5.3 the turbocharger speed for reference point 24 from the part loadtest is visible, compared to the reference point in Figure 5.2 the peak-to-peak

16

variation is smaller and the signal shifts over the engine cycle. More discrep-ancies are visible, some probably caused by the pulse rate which is roughly 7.8CAD per pulse.

The theoretical maximum resolution of the digital speed signal is thus lim-ited by the pulse rate dependent on the turbocharger speed and engine speed,when utilizing the one pulse per turbocharger revolution setting on the Micro-Epsilon control unit. The overall accuracy of the digital turbocharger speedmeasurements in general is difficult to determine by analyzing the signal alone,this is discussed more in the turbine performance estimation in chapter 6.

−400 −300 −200 −100 0 100 200 300 40078650

78700

78750

78800

78850

78900

78950

79000

79050Part load point 24

Crank angle [°]

Tur

boch

arge

r sp

eed

[rpm

]

Figure 5.3: Digital turbocharger speed signal for part load point 24, one enginecycle. A clear pulsating behavior is visible, less constant and shifting over theshown engine cycle.

5.2.2 Analog Micro-Epsilon signal

Presume the digital speed signals in Figure 5.2 and 5.3 depicts the turbochargerrotational speed correctly, a similar peak-to-peak amplitude and signal variationshould be expected in the analog signal. The analog speed signals for full loadpoint C and part load point 24 are visible in Figure 5.4, as seen in the figurethe signals contain a lot of noise. A low-pass filter was applied after checkingthe frequency spectrum of the signal, but the signal-to-noise ratio is too low tobe able to utilize the signal for performance calculations. Since the indicationequipment analog input range is ±10 V and the Micro-Epsilon sensor outputrange is 0-5 V one solution to increase the signal-to-noise ratio could be toamplify the signal and match it with the indication system input. Regardlessif the signal-to-noise ration can be improved, noise will be present in the speed

17

trace of the analog signal. Ensemble averaging the signal over all measuredengine cycles should reduce the noise level. Unfortunately this could not beperformed due to problems with extracting more than one engine cycle of thespeed signal from the indication system.

−400 −300 −200 −100 0 100 200 300 40094000

94500

95000

95500

96000

96500Full load point C

Tur

boch

arge

r sp

eed

[rpm

]

−400 −300 −200 −100 0 100 200 300 400

77000

77500

78000

78500

79000

79500

Part load point 24

Crank angle [°]

Tur

boch

arge

r sp

eed

[rpm

]

Figure 5.4: Analog turbocharger speed signal for full load point C and partload point 24, one engine cycle. The signals show a lot of noise, with a very lowsignal-to-noise ratio.

5.3 Turbocharger rotational acceleration

The turbocharger acceleration is an important parameter in the utilized powerestimation, derived from the time derivative of the measured turbochargerspeed. The turbocharger acceleration for full load point C derived from thedigital speed signal is shown in Figure 5.5. The time derivatives show a lot ofdiscrete levels, making it far from continuous and not very useful as an inputfor the estimation of the acceleration power. The cause of these discrete levelsis the turbocharger speed being a discrete sampled signal, and the derivativebeing approximated numerically.

To make the acceleration more continuous the turbocharger speed needsto be more continuous, this can be achieved by applying a low-pass filter onthe measured turbocharger speed signal. A 5th order Butterworth filter wasapplied to the speed signal to smooth out the data before differentiating it tothe turbocharger acceleration, chosen for its overall good performance. Thesmoothing technique with a filter was used by other researchers [20, 21]. The

18

turbocharger acceleration differentiated from the filtered turbocharger speed forfull load point C is shown in Figure 5.6, and clearly the filtered signal is morecontinuous without loosing too much information.

−400 −300 −200 −100 0 100 200 300 400−100000

−80000

−60000

−40000

−20000

0

20000

40000

60000

80000Full load point C

Crank angle [°]

Tur

boch

arge

r ac

cele

ratio

n [r

pm/s

]

Figure 5.5: Turbocharger acceleration for full load point C, one engine cycle.The acceleration trace show a lot of discrete levels due to the discrete sampledturbocharger speed signal.

19

−400 −300 −200 −100 0 100 200 300 400−80000

−60000

−40000

−20000

0

20000

40000

60000

80000Full load point C

Crank angle [°]

Tur

boch

arge

r ac

cele

ratio

n [r

pm/s

]

Figure 5.6: Turbocharger acceleration differentiated from filtered turbochargerspeed for full load point C, one engine cycle. The filtered speed signal clearlyreduces the discrete levels in the derived acceleration trace.

5.4 Pressure signals

In this section the acquired pressure signals and visible phenomena in the signaltraces are discussed and evaluated.

5.4.1 Total pressure signals

The measured total pressures for full load point C are shown in Figure 5.7. Clearpressure peaks are visible and the overall fluctuating behavior of the exhaustflow is conspicuously present in the pressure trace. Where the pressure peaks inone scroll a peak with less amplitude is visible in the pressure trace for the otherscroll, according to Winkler et al. [15] this lower peak is caused by backflow. Thebuildup of these pressure gradients between the different scrolls is the drivingforce for mass and energy transfer causing the exhaust to partially backflowfrom the high pressure scroll to the low pressure scroll. Another factor possiblyaffecting the lower pressure peaks is located at the turbine entry flange. InFigure 4.2 soot deposits are visible on the scroll divider wall, the gasket betweenthe exhaust manifold and the turbine flange only cover the outer perimeter andnot the divider wall. A small gap is thus open between the scrolls even before theentry of the turbine, which most certainly causes some energy transfer betweenthe scrolls.

20

−400 −300 −200 −100 0 100 200 300 400 2

2.5

3

3.5

4

4.5

5Full load point C

Crank angle [°]

Abs

olut

e pr

essu

re [b

ar]

OI

Figure 5.7: Total pressure signals for full load point C, one engine cycle. Innerscroll denoted I, outer scroll denoted O. Interference between the scrolls is clearlyvisible as small secondary peaks below each of the 6 maximum peaks.

Another phenomenon shown in Figure 5.7 is the fluctuation of the signalafter the pressure peak on each exhaust pulse. The signal from the pressuresensor is analog, thus some background noise will be present, however if theroot cause of these fluctuations was noise it would be present during the wholepressure trace. The fluctuation is instead believed to be induced by the pressuretubes and sensor mounting acting as a resonator. With the pressure tube as aconnecting neck and the threaded mounting socket for the sensor as a volume,the installation was modeled as a Helmholtz resonator. The measurements ofthe resonator dimensions where taken using a caliper and the speed of soundin the exhaust gas was estimated using the measured mean temperature at theturbine inlet flange. The dimensions of the resonator are very rough due tothe measurements being taken by hand. Since it is a small resonator, a pipelength of approximately 10 cm with a inner diameter of 0.6 cm and a volume ofapproximately 1.24 cm3, small variations make a big difference on the calculatedresonance frequency. Another factor contributing to the overall uncertainty isthe speed of sound being approximated from a mean temperature measured acertain distance from the resonator. With that in mind the resonance frequencyfor the resonator at full load point C was found around 1100 Hz, the frequencyspectrum of the total pressure traces for the inner and outer scroll revealed themean frequency of the fluctuations around 800 Hz. Even though the resonancefrequency is 300 Hz from the actual fluctuation frequency, the errors in thecalculations make it plausible that fluctuations are caused by the pressure pipesresonating. Another factor that could contribute to the fluctuations, is the flow

21

direction being some degrees from being perpendicular to the total pressuretube inlet.

The resonating behavior is visible through out the measurement series. Toeliminate these fluctuations, the pressure traces where ensemble averaged overthe 100 engine cycles measured. This did not work satisfactory and a degree offluctuations remained, instead the pressure traces were filtered with the sametechnique used in section 5.3 to smooth out the curves before turbine perfor-mance calculations.

5.4.2 Static pressure signals

The static pressures measured at the turbine inlet for the outer and inner scrollshow the flow pulsations with the same trend as the total pressure signals. Thestatic pressure traces for full load point C is shown in Figure 5.8, where somesmall signal fluctuations are visible. The flush mounted static pressure sensorscould behave in similar way as the total pressure sensors, i.e. the sensor socketand mounting acting as a resonator. Since the fluctuations where so small thesignals where filtered before calculations, and the fluctuations where consideredas noise.

Depending on the load point the amplitude of the fluctuations varies, nonof the measured points contained as severe fluctuations as in the total pressuretraces.

The static pressure trace measured after the turbine for full load point C isshown in Figure 5.9. This pressure trace also shows fluctuations and noise, andthe signal was filtered to make it more suitable for performance calculations.

22

−400 −300 −200 −100 0 100 200 300 400 2

2.5

3

3.5

4

4.5Full load point C

Crank angle [°]

Abs

olut

e pr

essu

re [b

ar]

OI

Figure 5.8: Static pressure signals at turbine inlet for full load point C, oneengine cycle. Inner scroll denoted I, outer scroll denoted O. The fluctuatingbehavior is obvious, with 6 distinct peaks.

23

−400 −300 −200 −100 0 100 200 300 400

1.15

1.2

1.25

1.3

1.35

1.4

1.45Full load point C

Crank angle [°]

Abs

olut

e pr

essu

re [b

ar]

Figure 5.9: Static pressure signal at turbine outlet for full load point C, oneengine cycle. The pressure fluctuations are clearly visible at the turbine outlet,with a obvious amplitude reduction.

5.5 Exhaust mass flow rate estimated from dy-namic pressure

The exhaust mass flow rate is a very important quantity when estimating theavailable energy in the exhaust gas. As mentioned before this quantity is notpossible to measure instantaneously, in this section a method to estimate theinstantaneous exhaust mass flow rate from the dynamic pressure in the turbineis explained. As a reference a mean exhaust mass flow rate was calculated fromthe air and fuel going in to the engine, this is presented as one value for eachmeasurement point from the indication system.

The dynamic pressure can be extracted from the definition of total pressureaccording to the conservation of linear momentum of the fluid, visible in equation5.1. When extracting the dynamic pressure from the static and total pressurenormally a Prandtl tube is used where both the pressures are measured in thesame location. In this application the static and total pressures was measuredapproximately 115 mm apart, thus an assumption was made that the propertyof the moving exhaust gas is unchanged over this distance.

ptotal = pstatic + pdynamic (5.1)

The dynamic pressure is the defined as a function of the flow velocity anddensity of the flowing gas, according to equation 5.2.

24

pdynamic =u2ρexh.

2(5.2)

From equation 5.2 it is possible to extract the flow velocity of the gas as afunction of the exhaust gas density and the dynamic pressure. The instanta-neous exhaust gas density can be estimated from a combination of Joules lawand the ideal gas law making the density a function of the instantaneous gastemperature. The instantaneous temperature is not possible to measure, in-stead it was estimated using the assumption that the temperature fluctuationsare caused by an adiabatic compression or expansion of the exhaust gases al-ready considered ideal. This method is utilized by several researchers [9,13,22],and uses the measured mean temperature at the turbine inlet as a referencevisible in equation 5.3.

Tinst.exh. = Tin,exh.mean

(ps,inst.ps,mean

) γexh.−1

γexh.

(5.3)

The instantaneous exhaust mass flow rate is then a function of the estimatedexhaust gas density, flow velocity and cross section area of one scroll at theturbine inlet, according to equation 5.4.

mexh. = uρexh.At (5.4)

The estimated exhaust mass flow rate for the outer scroll with a cross sectionarea at the turbine flange of approximately 17 cm2, the cycle mean of theinstantaneous estimation and the measured mean mass flow rate for full loadpoint C are shown in Figure 5.10. At first sight it seems that the instantaneousestimation underestimates the exhaust mass flow rate when comparing the meanof the instantaneous estimation and the measured mean calculated from theinflow of air and fuel. This mean measured exhaust mass flow rate representthe mass flow rate through both scrolls while the estimated instantaneous massflow rate is only displayed for the outer scroll.

25

−400 −300 −200 −100 0 100 200 300 4000

0.1

0.2

0.3

0.4

0.5

0.6

0.7

Crank angle [°]

Exh

aust

mas

s flo

w r

ate

[kg/

s]Full load point C

1. O2. Mean of 1.3. Mean measured

Figure 5.10: Estimated instantaneous exhaust mass flow rate for the outer scroll.A cycle mean of the instantaneous estimation and the measured mean from theindication system, for full load point C, one engine cycle.

The instantaneous mass flow rate was estimated for the inner scroll as well,and the total mass flow rate through the turbine is the sum of the mass flowrate through both the inner and outer scroll. The total instantaneous exhaustmass flow rate through the turbine for full load point C is presented in Figure5.11. The cycle mean of the total instantaneous exhaust mass flow rate matchthe mean measured exhaust flow with a relative error of 0.0025 in this case, whythe mean values are hard to distinguish from each other in Figure 5.11.

26

−400 −300 −200 −100 0 100 200 300 4000

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

Crank angle [°]

Exh

aust

mas

s flo

w r

ate

[kg/

s]Full load point C

1. Total2. Mean of 1.3. Mean measured

Figure 5.11: Estimated instantaneous exhaust mass flow rate through the tur-bine. A cycle mean of the instantaneous estimation and the measured meanfrom the indication system, for full load point C, one engine cycle.

The instantaneous exhaust mass flow rate for part load point 3 and 25 areshown in Figure 5.12 and Figure 5.13 respectively. The same estimation trendis visible and the relative error between the mean values are 0.034 for part loadpoint 3 and 0.04 for part load point 24. The accuracy and limitations of thisestimation is discussed later in this chapter.

One phenomenon visible for all calculated points, not just the three referencepoints is the exhaust mass flow rate minimum around 140 CAD. This low peakis caused by the flow velocity estimation and specifically the very fluctuatingdynamical pressure turning negative at certain points indicating that the flowchanges direction. This issue is further addressed later in this chapter.

27

−400 −300 −200 −100 0 100 200 300 4000

0.05

0.1

0.15

0.2

0.25

0.3

0.35

0.4

Crank angle [°]

Exh

aust

mas

s flo

w r

ate

[kg/

s]

Part load point 3

1. Total2. Mean of 1.3. Mean measured

Figure 5.12: Estimated instantaneous exhaust mass flow rate through the tur-bine. A cycle mean of the instantaneous estimation and the measured meanfrom the indication system, for part load point 3, one engine cycle.

−400 −300 −200 −100 0 100 200 300 4000

0.1

0.2

0.3

0.4

0.5

0.6

0.7

Crank angle [°]

Exh

aust

mas

s flo

w r

ate

[kg/

s]

Part load point 24

1. Total2. Mean of 1.3. Mean measured

Figure 5.13: Estimated instantaneous exhaust mass flow rate through the tur-bine. A cycle mean of the instantaneous estimation and the measured meanfrom the indication system, for part load point 24, one engine cycle.

28

5.5.1 Discussion of accuracy and validity

The instantaneous exhaust mass flow rate approximation shows fluctuating be-havior as expected, six distinct peaks are visible. After each peak the mass flowis slowed down, if the instantaneous approximation trace reflect real fluctuatingconditions a small increase in mass flow corresponding to mass transfer betweenthe two scrolls should be visible after the distinct peaks. This behavior is clearlyvisible at certain points in Figure 5.11 and Figure 5.13.

The lowest spot found around 140 CAD when the exhaust mass flow rate iszero is believed to be erroneous. If indeed the flow would slow down completelyit is strange that this behavior is only found after one of the pressure peaks.If this low point is caused by cycle to cycle variations it should not be presentin all measurement, and if it is related to the behavior of a specific cylinderthe deviation is very big for a normal engine operation. Instead it is believedto be caused by the exhaust flow velocity based on the dynamic pressure, andforemost the static pressure measurement. The static pressure is measured onthe side of the turbine flange and not in the free stream as the total pressure,this can result in the dynamic pressure estimation being erroneous dependingon how the flow is distributed along the cross section.

Other sources affecting the accuracy of the estimation is that the cross sec-tion area for one scroll at the turbine flange approximated to 17 cm2 differs fromthe cross section area at the total pressure probe location where it increases tosome degree at the start of the volute.

With this in mind when comparing the cycle mean of the instantaneous massflow rate trace with the measured mean, the results are reasonable in terms ofthe mass flow rate levels. The biggest relative error found between the meanvalues trough out the measurement series was 0.084 and if the measured meanis considered correct, the instantaneous approximation seems to overestimatethe exhaust mass flow rate to some degree in most operating points.

The instantaneous estimation is only valid when the wastegate is closed andno bypass around the turbine exists, when used to estimate the energy availablein the exhaust gas. This is based on the fact that the wastegate flow is unknownwhen the wastegate is open.

29

5.6 Exhaust pulse identification

When evaluating the fluctuating behavior of the turbine, identifying the originof each peak in the signal traces obtained could be useful. With the measuredcylinder pressure in cylinder 6 and the static pressure measured in the exhaustmanifold close to cylinder 6 it is possible to detect which exhaust pulse thatcorresponds to which cylinder. In Figure 5.14, the cylinder pressure in cylinder6 is plotted against the total pressures in both turbine scrolls and the staticpressure in the exhaust manifold close to cylinder 6 for full load point C arevisible. After the cylinder pressure peak at approximately 10 CAD, the highestpeak in the exhaust manifold pressure is seen at approximately 180 CAD whichcorresponds to the exhaust pulse that left cylinder 6. Shortly after the increasein the exhaust manifold pressure for this peak, it is seen that there is an in-crease in the total pressure in the outer scroll, this peak in the total pressure isbelieved to correspond to the pressure pulse from cylinder 6. When comparingthe amplitudes of the exhaust manifold pressure and phasing between the totalpressure in the outer scroll for the rest of the pressure peaks, this seem to cor-related with the distance between the sensors. One factor that would indicatethis theory being wrong is the amplitude of the total pressure peaks in the outerscroll related to the distance from the physical position of each cylinder to thetotal pressure sensor. However since these signals are filtered the amplitudesfor each pressure peak is not exactly correct, especially easy to envision whencomparing them to the unfiltered signals for the same operating point in Figure5.7.

With the engine firing order (1-5-3-6-2-4) it is easy to identify the originfor the rest of the pressure pulses, in Figure 5.14 each pressure pulse is markedwith its originating cylinder. The peaks can be correlated in the turbochargerspeed trace as well, in Figure 5.15 the turbocharger speed for full load point Cis shown with each peak marked with its originating cylinder.

30

−400 −300 −200 −100 0 100 200 300 4002

2.5

3

3.5

4

4.5

5

Crank angle [°]

Abs

olut

e pr

essu

re [b

ar]

0

20

40

60

80

100

120

140

160

180

200

Abs

olut

e cy

linde

r pr

essu

re [b

ar]

Full load point C

4 1 5 3 6 2

ptot,O

ptot,I

pexh.s

pcyl6

Figure 5.14: Cylinder pressure in cylinder 6 plotted against the total pressuretraces for each turbine scroll and the static pressure in the exhaust manifoldoutside cylinder 6, for full load point C, one engine cycle. Each pressure peakmarked with the cylinder it originates from.

−400 −300 −200 −100 0 100 200 300 40095350

95400

95450

95500

95550

95600

95650

95700

95750

95800

95850Full load point C

Crank angle [°]

Tur

boch

arge

r sp

eed

[rpm

]

4 1 5 3 6 2

Figure 5.15: Turbocharger speed for full load point C, one engine cycle. Eachpeak marked with the cylinder it originates from.

31

6. Results

In this chapter the instantaneous turbine performance estimations are presented,the accuracy and limitations of the estimation are discussed and characterizationof the exhaust flow is presented.

6.1 Isentropic turbine power

The instantaneous isentropic turbine power was calculated according to equa-tion 3.2, all needed parameters are in previous chapters estimated on a CADbasis. Since the instantaneous exhaust mass flow rate is defined as one param-eter through the turbine, the instantaneous temperature and total pressure arecalculated as an arithmetic mean value between the respective parameters forthe two scrolls, a technique adopted by Uhlmann et al. [23].

The instantaneous isentropic turbine power for full load point C is presentedin Figure 6.1, the fluctuating behavior is clearly visible and a low point atapproximately 140 CAD is visible caused by the instantaneous exhaust massflow rate estimation. The same fluctuating behavior and low power point arepresent in the isentropic turbine power results for part load point 3 and 24,shown in Figure 6.2 and 6.3 respectively.

−400 −300 −200 −100 0 100 200 300 400 0

20

40

60

80

100

120

140

160

180Full load point C

Pow

er [k

W]

Crank angle [°]

Figure 6.1: Estimated isentropic turbine power, for full load point C, one enginecycle. A fluctuating behavior is clearly visible with some discrepancies believedto be caused by the instantaneous exhaust mass flow rate estimation.

32

−400 −300 −200 −100 0 100 200 300 400 0

5

10

15

20

25

30

35

40Part load point 3

Pow

er [k

W]

Crank angle [°]

Figure 6.2: Estimated isentropic turbine power, for part load point 3, one enginecycle. A clear fluctuating behavior is visible, with a low point at 140 CAD.

−400 −300 −200 −100 0 100 200 300 400 0

20

40

60

80

100

120Part load point 24

Pow

er [k

W]

Crank angle [°]

Figure 6.3: Estimated isentropic turbine power, for part load point 24, oneengine cycle. The deviating behavior mentioned before is visible at 140 CAD.

33

6.2 Utilized power

The utilized power consists of the acceleration power, compressor power andthe power lost due to bearing friction, defined according to equation 3.3. Themechanical efficiency of the bearings was set to 97 % independent of the operat-ing point, usually stated by manufacturers somewhere between 95-100 %. Thevalue chosen is used from [14] since no data for the specific turbocharger wasfound. The rotor assembly mass moment of inertia was 3.3751e-4 kgm2 suppliedby the turbo manufacturer. In Figure 6.4 the utilized power for full load pointC is shown. After each of the six big pulses when the utilized power decreasessmaller local peaks are visible, these smaller peaks are believed to be caused bythe energy transfer between the two scrolls. One possible explanation is thatthis transfered energy from one scroll to the other affect the turbine wheel andthus gives a local energy contribution increasing the power output locally.

The same behavior is visible for part load point 3 in Figure 6.5, anotherinteresting phenomena for this load point is that the utilized power drops belowzero. A possible explanation for this phenomenon could be the cycle-to-cyclevariation of the turbocharger speed present at lower engine speeds and loads,since only one cycle could be extracted from the indication system this could notbe fully investigated. Another explanation could be that the compressor powerwhich was assumed constant is in fact fluctuating and could be fluctuating out ofphase when compared to the acceleration power, discussed later in this chapter.

−400 −300 −200 −100 0 100 200 300 40045

50

55

60

65

70

75

80

85

90

95Full load point C

Pow

er [k

W]

Crank angle [°]

Figure 6.4: Utilized power, for full load point C, one engine cycle. Energytransfer between the scrolls is seen as a small increase on the down slope of eachbig peak.

34

−400 −300 −200 −100 0 100 200 300 400−5

0

5

10

15

20Part load point 3

Pow

er [k

W]

Crank angle [°]

Figure 6.5: Utilized power, for part load point 3, one engine cycle. As seen, thepower turns negative at certain points, believed to be caused by the cycle-to-cycle variation of the turbocharger speed at this point.

6.3 Turbine efficiency

With both the isentropic turbine power and utilized power estimated, the in-stantaneous turbine efficiency can be calculated according to equation 3.6. Theinstantaneous turbine efficiency for full load point C is shown in Figure 6.6, thelow point visible at approximately 140 CAD is caused by the instantaneous massflow approximation being zero at the same CAD. The two peaks in Figure 6.6surrounding the low point at approximately 120 and 160 CAD is the result ofthe utilized power being higher than the isentropic turbine power clearly visiblein Figure 6.7 where both the utilized power and the isentropic turbine powerare plotted.

From a pure mathematical point of view the region around the low pointat 140 CAD should be considered as one big peak since the isentropic turbinepower being zero causes this behavior according to the definition of the turbineefficiency. The theory of this consideration is clearly visible for the other bigpeak in Figure 6.6 at approximately -220 CAD where the height of the peakcorresponds directly to the difference in powers at the same CAD in Figure 6.7.

Some peaks greater than unity in addition to the ones previously describedexist in Figure 6.6, these peaks are found by several researchers attemptingan instantaneous turbine efficiency estimation [14, 21]. Winterbone et al. [21]states a couple of reasons for this behavior, one is that the peaks are causedwhen the pressure ratio over the turbine is close to unity. This is not the

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cause in this case, another reason stated by Winterbone et al. is that thecalculated isentropic turbine power is calculated at a measuring plane before theactual energy utilization occur on the turbine wheel. This causes the calculatedisentropic turbine power to be out of phase compared to the utilized power. Themost sensitive part if out of phase is the acceleration power and pressure traceat the turbine inlet and this sensitivity depends on the turbocharger operatingcondition, being less sensitive at higher exhaust flows.

The cycle mean turbine efficiency over one engine cycle for full load pointC is approximately 73 %. Considering the deviations in the instantaneous tur-bine efficiency trace the estimation tends towards being a bit high, but is stillreasonable according to [6].

−400 −300 −200 −100 0 100 200 300 4000

0.5

1

1.5

2

2.5

3

3.5Full load point C

η turb

. [−]

Crank angle [°]

η

turb.

Mean of ηturb.

Exh

aust

mas

s flo

w r

ate

[kg/

s]

0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

Figure 6.6: Turbine efficiency and exhaust mass flow rate, for full load pointC, one engine cycle. The turbine efficiency fluctuates over the engine cycle andvalues over unity are present, believed to be caused by the phase differencebetween the measuring plane and the energy utilization at the turbine wheel.

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−400 −300 −200 −100 0 100 200 300 400 0

20

40

60

80

100

120

140

160

180Full load point C

Crank angle [°]

Pow

er [k

W]

P

utilized

Pisentropic

Figure 6.7: Utilized power and isentropic turbine power, for full load pointC, one engine cycle. In this power comparison the points where the turbineefficiency turns above unity are clearly visible.

The instantaneous turbine efficiency estimation for part load point 3 is shownin Figure 6.8, here the turbine efficiency drops below zero at certain points. Thisis caused by the utilized power being negative. A possible cause for this behavioris believed to be the quite large cycle-to-cycle variations and the rather unsta-ble characteristics of the exhaust flow at low engine speeds and loads causingthe turbocharger speed to fluctuate very much between the engine cycles, andas mentioned before only the first cycle could be extracted for calculations.Another factor believed to aid this behavior is the constant compressor powerassumption. A third possible explanation for this behavior could be the shaft as-sembly freewheeling, i.e. a momentum transfer from the turbine to the exhaustgas [13].

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−400 −300 −200 −100 0 100 200 300 400−1

−0.5

0

0.5

1

1.5

2

2.5

3Part load point 3

η turb

. [−]

Crank angle [°]

η

turb.

Mean of ηturb.

Exh

aust

mas

s flo

w r

ate

[kg/

s]

0

0.05

0.1

0.15

0.2

0.25

0.3

0.35

Figure 6.8: Turbine efficiency and exhaust mass flow rate, for part load point3, one engine cycle. The efficiency drop below zero at certain points, possiblycaused by the phase difference between the obtained powers.

As a comparison the instantaneous turbine efficiency was calculated usinga mean exhaust mass flow rate and a mean turbine inlet temperature as well,thus the only parameter fluctuating in the isentropic turbine power is the pres-sure ratio over the turbine. The difference between the two estimated turbineefficiencies is shown in Figure 6.9 for full load point C and in Figure 6.10 forpart load point 3.

In Figure 6.9 it seems that the turbine efficiency calculated from the constantexhaust mass flow rate and turbine inlet temperature seem to be out of phasewhen compared to the turbine efficiency calculated with instantaneous param-eters, especially visible from -200 to 0 CAD. A possible cause for this could bethe inconsistency of the instantaneously estimated exhaust mass flow rate overone engine cycle, making the boundaries of each exhaust pulse fade. Overall thedifference between the two methods seem quite small for this operating point.

The same comparison for part load point 3 in Figure 6.10 show a bettercorrelation between the fluctuating pressure and exhaust mass flow rate, thoughthe efficiency calculated with a mean exhaust mass flow rate and turbine inlettemperature seem to overestimate the instantaneous turbine efficiency. Whencomparing mean efficiencies over one engine cycle this correlates with theory,that the exhaust mass flow rate is more constant at higher engine speeds andloads.

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−400 −300 −200 −100 0 100 200 300 4000

0.5

1

1.5

2

2.5

3

3.5Full load point C

η turb

. [−]

Crank angle [°]

1. η

turb.

Mean of 1.2. η

turb. (const.)

Mean of 2.

Figure 6.9: Turbine efficiency, for full load point C, one engine cycle. Thered curve is calculated with a constant exhaust mass flow rate and a constantturbine inlet temperature. The black curve is calculated using the instantaneousexhaust mass flow rate and the instantaneous turbine inlet temperature.

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−400 −300 −200 −100 0 100 200 300 400−1

−0.5

0

0.5

1

1.5

2

2.5

3Part load point 3

η turb

. [−]

Crank angle [°]

1. η

turb.

Mean of 1.2. η

turb. (const.)

Mean of 2.

Figure 6.10: Turbine efficiency, for part load point 3, one engine cycle. Thered curve is calculated with a constant exhaust mass flow rate and a constantturbine inlet temperature. The black curve is calculated using the instantaneousexhaust mass flow rate and the instantaneous turbine inlet temperature.

6.4 Flow characterization

When plotting the turbine efficiency generally the optimal efficiency for a radialturbine operating during steady-state flow is approximately 75 % and is found ata BSR of approximately 0.7. A typical performance curve in an ideal steady flowcase gets the shape of a parabola, peaking at the maximum turbine efficiency.The turbine efficiency plotted against BSR for part load point 3 over one exhaustpulse is shown in Figure 6.11, and the shape of the curve deviates from the idealparabola expected under steady flow.

In Figure 6.12, the turbine efficiency is plotted against BSR for one exhaustpulse of full load point C. The ideal flow behavior is not visible here either,indicating that the performance of the turbine cannot be described with steadystate flow maps during pulsating exhaust flow. The fluctuations in BSR aremost certainly caused by fluctuations in pressure ratio over the turbine sincethe turbocharger speed only fluctuates by a couple of percent at most.

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0.5 0.6 0.7 0.8 0.9 1 1.1 1.2−0.1

0

0.1

0.2

0.3

0.4

0.5

0.6

0.7Part load point 3

η turb

. [−]

U/Cs [−]

Figure 6.11: Turbine efficiency vs.BSR, for part load point 3, one exhaust pulse.

0.64 0.66 0.68 0.7 0.72 0.74 0.76 0.780.35

0.4

0.45

0.5

0.55

0.6

0.65

0.7Full load point C

η turb

. [−]

U/Cs [−]

Figure 6.12: Turbine efficiency vs. BSR, for full load point C, one exhaust pulse.

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6.5 Comments on results & accuracy

The estimation of the instantaneous turbine efficiency fluctuates due to thefluctuating behavior of the exhaust flow which is far from being steady, whencomparing the results in terms of mean values over one engine cycle the resultsseem reasonable. However there are a lot of uncertainties related to the turbineefficiency estimation, especially the estimation of the isentropic turbine power.

Compared to the compressor power the isentropic turbine power fluctuatesto a great extent. The fluctuations caused by the pulsating exhaust flow, incombination with quantities such as temperature and exhaust mass flow ratebeing very hard to obtain instantaneously, certainly makes the accuracy fluctu-ate over the investigated operating points. The exhaust mass flow rate basedon the dynamical pressure and the instantaneous temperature based on the as-sumption that the exhaust gas behaves adiabatically over the turbine, affect theoverall accuracy the most, since no instantaneous reference values exist.

The issue of phasing between all calculated powers is another factor affectingthe overall accuracy of the estimation, as explained before the isentropic turbinepower is estimated a certain time before the actual energy extraction occurs onthe rotor. This lag was left in the final results, based on the exhaust mass flowrate approximation being very inconsistent over the engine cycle, especially athigher flow rates. The exhaust mass flow rate is believed to be erroneous basedon exhaust mass flow simulations performed by other researchers [14, 15]. Thisinconsistency made the boundaries of each exhaust pulse fade at higher enginespeeds and loads, and made phase shifting attempts unsuccessful.

Other uncertainties are related to the assumption of a negligible energy lossin the form of heat transfer from the turbine to the compressor and heat lossesfrom the compressor, which are enclosed in the estimated turbine efficiency.

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7. Summary, conclusions and fu-ture work

In this thesis the objective was to estimate the on-engine instantaneous turbinepower and efficiency. As previously described the turbine performance is mostoften measured in off engine conditions during steady-state flow. This is notthe case when the turbocharger is mounted on the engine, when every quantityrelated to the exhaust gas flowing through the turbine is pulsating. To furthercomplicate matters a twin scroll turbine was used. The turbine thus experiencesa very broad range of flow conditions during one engine cycle, most of whichare not the optimal conditions the turbine was designed for.

7.1 Conclusions

The estimated instantaneous isentropic turbine power and the estimated utilizedshaft power revealed promising results. Especially the isentropic turbine powerconsidering that two very sensitive parameters, instantaneous exhaust mass flowrate and turbine inlet temperature where estimated individually since they arenot far from impossible to measure instantaneously in the harsh working en-vironment of the turbine. In terms of turbine efficiency peaks obtained at theoptimal steady state efficiency of 75 % where found at several measured oper-ating points. As a reference the instantaneous turbine efficiency was estimatedwith a constant exhaust mass flow rate and a constant turbine inlet temperature,with the pressure ratio over the turbine as the only fluctuating parameter onthe turbine side. This estimation showed results close to the fully instantaneousestimation at higher engine speeds and loads, pointing towards the exhaust massflow rate being less fluctuating under these conditions.

Still the the fully fluctuating CAD resolved efficiency and turbine powerestimations have a lot of discrepancies and questionable details over an evaluatedengine cycle, unfortunately too many to describe the CAD resolved on engineperformance of the turbine with proven confidence.

Improvements regarding the instrumentation, the measurement techniqueand aspects regarding the isentropic turbine power and utilized shaft power areneeded in order to be able to rely on the CAD resolved turbine performanceestimation.

7.2 Future work

The power consumed by the compressor, assumed constant, would be inter-esting to investigate with CAD resolved measurements, in order to investigateto what extent the constant assumption is valid. A CAD resolved compressorpower could possibly help explain some of the discrepancies found in the tur-bine efficiency. Another part of the utilized power in need of development is theassumed static bearing friction losses, an accurate bearing friction model anda better knowledge of the bearing systems dynamical behavior should improvethe turbine performance estimation.

The isentropic turbine power is thought to be the biggest source of error inthe overall turbine efficiency, due the many uncertain factors being instanta-

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neously estimated and not measured directly. The instantaneous exhaust massflow rate is a parameter in need of further improvements, especially how thedynamical pressure is measured on which the exhaust mass flow rate is based.One idea to improve this estimation could be to implement a Prandtl tube in thelocation of the total pressure measurement. With a Prandtl tube there is a pos-sibility to measure the static and total pressures in the same location in the freestream and thus a more fair dynamic pressure estimation could be achieved.However there is a possibility that the dimensions of the tube needed couldinterfere and disturb the flow through the turbine. The instantaneous temper-ature at the turbine inlet is another parameter with questionable accuracy, dueto the difficulty of measuring this quantity instantaneously. As a reference, asimulation of both the instantaneous temperature and exhaust mass flow ratecould have been useful when evaluating their signal traces.

The instantaneous turbine performance estimations are limited to the closedwastegate region due to the unknown flow rate through the wastegate when itis open. Therefore a model of the wastegate flow is needed in order to expandthe validity range for the method used in this thesis, though determining theon-engine wastegate flow is very complex.

A model of the turbocharger as dynamical system to evaluate the turbineperformance could include the turbines cyclic behavior in order to estimate theissue of phasing between measured quantities. It could describe the filling andemptying of the turbine volume at the arrival of the exhaust pulses and howthis affects the performance. In essence describing the timing between measuredsignals at turbine inlet and the actual energy utilization on the turbine wheel.

Regarding the sampling rate and details of the instantaneously acquiredsignals some changes could be made to improve the signal quality. The tur-bocharger speed sensor could be changed to send a pulse for each passing com-pressor blade. This would increase the detail level in the signal and makingit more appropriate for differentiating to obtain the turbocharger acceleration.The instantaneous total pressure signal trace could be improved by makinga proper frequency analysis of the pressure tube and sensor socket and maybechange the dimensions to avoid the fluctuations caused by the system resonance.

The CAD resolved signals with each of the peaks originating cylinder knowncould serve another purpose. Since the pressure pulses affecting the turbinewheel are so distinct and visible in the turbocharger speed, deviations causedby for example a cylinder not working efficiently could possibly be detected bydeviations in turbocharger speed for that specific exhaust pulse. The instanta-neous turbocharger speed could in a sense serve as on type of reference to seethat the engine works as it is supposed to when evaluating the instantaneousturbine performance.

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