Implementation of an exhaust system for an opposed piston two stroke HCCI engine

JAVIER LAINEZ MARTÍ

Master of Science Thesis Stockholm, Sweden 2010

Implementation of an exhaust system for an opposed piston two stroke HCCI engine

Javier Lainez Martí

Master of Science Thesis MMK 2010:15 MFM132 KTH Industrial Engineering and Management Machine Design SE-100 44 STOCKHOLM

Examensarbete MMK 2010:15 MFM132

Anpassning av ett avgassystem för en motkolvs tvåtaktsmotor med HCCI-förbränning

Javier Lainez Martí

Godkänt Examinator Handledare 2010-06-29 Hans-Erik Ångström Hans-Erik Ångström Uppdragsgivare Kontaktperson HCCI Technologies AB Rafael Villasmil

Sammanfattning

Målet för detta examensarbete har varit att förbättra driften av en motkolvs tvåtaktsmotor med HCCI-förbränning. Huvudfokus för arbetet har varit på gasväxlingsprocessen, då främst avgasprocessen.

Motorn har studerats med utgångspunkt från hur en tvåtakts Otto-cykel fungerar. Gasväxlingsprocessen i tvåtaktsmotorer kännetecknas av behovet av att snabbt få ut den förbrända gasen och införa ny blandning vid varje expansionstakt, samt avsaknaden av ventiler. Behovet att kunna kontrollera gasflödet genom cylindern ger upphov till två av de mest karaktäristiska systemen på tvåtaktsmotorer: spolpump samt avgaspipa. Utformningen och genomförandet av dessa två system har skapat ett flertal uppgifter som det varit nödvändigt att lösa: porttider har uppmätts, ett luftintag med trycksatt luft har byggts för att kunna ändra förhållandet på insugningsluften. Olika insprutningssystem har också testats.

Slutligen har en avgaspipa installerats och trimmats in. Utifrån målsättningen, så har ett logiskt sätt att angripa problemet gjorts:

• Beteende hos kompressibelt gasflöde i HCCI-motorns kanaler har studerats.

• Avgasrörets dimension och dess påverkan på motor prestanda har undersökts genom att prova olika utformningar på avgasröret.

• Experimentella observationer av tryckpulsationer i avgasflöden har gjorts för att kunna gör finjusteringar.

Master of Science Thesis MMK 2010:15 MFM132

Implementation of an exhaust system for an opposed piston two stroke HCCI engine

Javier Lainez Martí

Approved Examiner Supervisor 2010-06-29 Hans-Erik Ångström Hans-Erik Ångström Commissioner Contact person HCCI Technologies AB Rafael Villasmil

Abstract

The aim of this thesis is the improvement of the operation of an opposed piston two stroke engine with HCCI-combustion. The main focus is the charge exchange process, concretely the exhaust process.

The engine has been studied from the perspective of the two stroke Otto cycle. Charge exchange process in two stroke engines is characterized by the need of expelling the burnt gases and introducing the fresh mix each expansion stroke, and the absence of valves. Control of the gas flow through the cylinder motivates two of the most characteristic systems on two-stroke engines: scavenging compressor and exhaust pipe. The design and implementation of these two systems has needed the development of several tasks: port timing has been characterized; a compressed air intake installation has been built in order to be able to modify the inlet air conditions; different injection systems have been tested as well.

Finally, an exhaust pipe has been installed and tuned. With this aim, a logical approach to the problem has been done:

• The behavior of compressible gas flow through ducts has been studied.

• The influence of the dimensions of the exhaust pipe has been understood by doing a design.

• The experimental observation of the pressure pulse propagation through the exhaust flow has allowed the exhaust pipe tuning.

Acknowledgements

First of all, Hans-Erik Ångström allowed me to work in an incredible project in one of the best universities along the world. Thank you for all your advising and help.

HCCI technologies AB and specially Tom Whitlock have delivered a continued interest and support for this project.

Thanks to Patrick Hellgren, for all his solutions for whatever mechanical problem that took place. Without him it could have been impossible to develop this work. Thanks as well to Bengt Aronsson, Jack Ivarsson and Ulf Andorff.

The people of the department have made my stay more pleasant. Thanks to Stefan, Anders, and all the others.

The Shell Ecomarathon experience in Lausitzring was amazing. I am so grateful to all the members of the Agilis-V team for his passionate work and dedication. There is nothing like see the fruits of one´s labour.

This experience would not have the same without all the friends that have accompanied me. Carlos, Adrián y Antonio, thank you for sharing this great year, for sure one of the best in all my life.

I am really grateful to my father, for his push, and to my sister and my grandparents, for giving me happiness from Valencia. To my mother, for allowing me to live the life I want. Ester, it is impossible to explain with words what I feel towards you.

Finally, to Rafael Villasmil: this is probably my first and last Master of Science, but it has been the second one for you. You should feel holder of this work as it was yours. Thank you for behaving like a brother.

Contents

1. Introduction ...... 3

2. Definitions and considerations on angular references ...... 5

3. Optimization of the engine and test bench ...... 7 3.1. Second block and spare parts ...... 7 3.2. Fuel ...... 8 3.3. Rotary encoder ...... 8 3.4. Torque signal filters ...... 9 3.5. Pressure and temperature sensors on intake manifold and exhaust circuit ...... 14 3.6. Intake compressed air supply system ...... 15 3.6.1. Objective ...... 15 3.6.2. Design and configuration ...... 16 3.6.2.1. Pressure regulator valve ...... 16 3.6.2.2. Safety valve ...... 16 3.6.2.3. Flow meter ...... 17 3.6.2.4. Air heater ...... 17 3.6.3. Adjustment of an intake pressure ...... 17 3.7. Oil pump ...... 18 3.8. Injection system modifications ...... 19 3.8.1. Bosch 0280150995 injector ...... 19 3.8.2. Injection calibration ...... 19 3.8.3. Empirical determination of the Injection Time vs Mass of fuel per injection curve ...... 22 3.8.4. Comparison between different injection systems ...... 23 3.8.4.1. Port Injection ...... 24 3.8.4.2. Direct Injection vs Axial Deflected Direct Injection comparison .. 26 3.9. Approaches to two-stroke combustion with high compression ratio ...... 29

4. Implementation of an exhaust system ...... 31 4.1. Exhaust gases flow in a two-stroke engine ...... 31 4.2. Design of the exhaust pipe ...... 32 4.3. Selection and modification of a commercial exhaust pipe ...... 35 4.4. Exhaust pipe tuning ...... 37

5. Results ...... 40 5.1. Tests procedure ...... 40 5.2. Exhaust pipe tuning results ...... 40 5.2.1. Tuned Length ...... 40 5.2.2. Tail pipe diameter ...... 44 5.3. Comparison between the old exhaust pipe designed with 1-D simulation software and the new tuned commercial exhaust pipe...... 45

6. Conclusions ...... 47

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7. Future work ...... 49

8. References ...... 50

Appendix 1: Demostration concerning the Internal DC ...... 51 Appendix 2: Geometrical description MS Excel worksheet ...... 55 Appendix 3: Exhaust Pipe designer MS Excel worksheet ...... 62 Appendix 4: Calibration of the Optrand C33233-Q pressure transducer...... 65 Appendix 5: Calibration of the Gems 2200AGA4001A2UA pressure transducer ... 69 Appendix 6: Calibration of the torque sensor ...... 72 Appendix 7: Calibration of the air flow meter ...... 75 Appendix 8: 12 V power electric map ...... 79

2 1. Introduction

22 is an opposed piston, one cylinder, 2-stroke engine, with two operation modes: Spark Ignition and Homogeneous Charge Compression Ignition. Spark Ignition combustion consists of burning a premixed charge of gasoline and air by mean of a spark, so a flame would travel all along the combustion chamber. Combustion takes place with spatial discontinuity, and close to the flame front, high temperatures promote NOx formation. On the other hand, the HCCI mode burns a premixed homogeneous lean air-fuel mixture like a Diesel engine, Compression Ignition. There are differences between HCCI combustion and Diesel combustion: charge is homogeneous and lean, so there are not rich zones close to the spray like in diesel engines, and therefore soot is practically avoided; and flame does not exist. All the regions of the combustion chamber have a simultaneous temporal evolution. Combustion process, unlike diesel cycle, happens at constant pressure, as in Otto cycle. Higher compressions ratios are used, which together with a not-spatial-discontinuity fast combustion allows efficiencies better than both SI and CI engines. Moreover, absence of a flame front reduces drastically combustion temperatures and hence NOx formation.

The control of the ignition in HCCI operation can be mainly done by affecting inlet conditions: lambda, intake temperature and pressure, octane number of fuel, internal and external EGR…But the most influent parameter is compression ratio. The compression ratio control is what motivates engine’s unique configuration, given by two opposite pistons moving in one cylinder. The minimum volume of the combustion chamber (and so the compression ratio) is controlled by phasing both crankshafts while the engine is under operation. This is done with a patented system consisting of 4 gearwheels transmitting the rotation between crankshafts, which can get a phase change because of the positioning control system of the intermediate not-fixed-shaft wheels. The phase is adjusted according to the signals of two Hall Effect sensors and a cylinder pressure transducer. The cylinder pressure measurement allows monitoring the combustion process, focusing in the cycle to cycle evolution of the start of the ignition. The gearwheels attached to the crankshafts have a missing tooth. The phase between both crankshafts is known by identifying when each missing tooth passes in front of each Hall Effect sensor. The linear actuator that controls the phasing mechanism is activated according to these signals. Figure 1 shows a view of the engine from the phasing mechanism side.

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Figure 1. Engine

Figure 2 shows the transversal section of the engine. Particularly relevant is the scavenging process, uniflow type.

Figure 2. Transversal section of the engine

The engine’s charge exchange process study and improvement is the main topic of this thesis. The need of expelling the burnt gases and introducing the fresh mix each expansion stroke and the absence of valves makes control of gas flow through the cylinder difficult, and motivates two of the most characteristic systems on two-stroke engines: scavenging compressor and exhaust pipe. The exhaust pipe tuning is the most important objective of this work. Different injection systems have been tested as well, and an intake compressed air installation for allowing changes on air’s pressure and temperature has been proved. All the development of the engine can be consulted within this thesis.

4 2. Definitions and considerations on angular references

The control of the compression ratio is based on the cylinder pressure and the relative angular position of both crankshafts, measured with a pressure transducer and Hall Effect sensors. Many other parameters such as Ignition Angle, Ignition Current, Injection Angle, are introduced in the ECU with reference to a certain angular position (TDC, Top Dead Center) of one of the crankshafts. Data taken from crankshaft angle logs must be referenced also to a crankshaft angle scale, usually with 0 in the TDC. Due to the fact that the engine has two crankshafts for the same cylinder, it is extremely important to know which parameters are referenced to which crankshaft. For a better understanding, the different TDCs will be placed on a common scale.

Three different scales can be defined: • Crankshaft angle scale with 0 on the Intake side crankshaft TDC (airTDC). • Crankshaft angle scale with 0 on the Exhaust side crankshaft TDC (exTDC). • Crankshaft angle scale with 0 on the Inner DC. The Inner DC (IDC) is defined as the angle where the minimum volume between both pistons occurs.

For building a common scale the most important consideration is that the Exhaust side crankshaft always reaches its TDC before the Intake side crankshaft does. It is easy to realize that the exTDC will happen exactly an angle consisting in the phase difference before the airTDC.

The IDC matches exactly half the phase difference after the exTDC and before the air TDC. A demonstration of this can be read on Appendix number 1.

Figure 3 clarifies better the concepts explained before, where positive increments of angle mean crankshaft displacement on the real rotating direction.

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Figure 3. Volume between pistons vs different angle scales, with 12º of phase difference

Crank Angle PDAQ logs will place its 0 in a TDC that depends on the crankshaft where the rotary encoder is installed (for more information about the rotary encoder, consult section 3.).

All the TDCs described in the ECU parameters menu (see table 3 on section 4.1 in Ref [1]) are referred to a crankshaft scale with 0 on the airTDC. From now on, if no reference is given, TDC would be referred to airTDC.

Other important nomenclature used in this book is A (after) or B (before) before the referenced TDC. For example: AairTDC means After air Top Dead Center.

6 3. Optimization of the engine and test bench

3.1. Second block and spare parts

The achievement of a continuous and systematic research process needs a forecast of failures of any kind in the engine and the test bench. Particularly, mechanic failures of parts can stop the tests during weeks, since components are mostly not commercial but custom manufactured. Even commercial products have a delivery time that can take months. Because of this, a complete set of spare parts was prepared, including:

• Second operative block: cylinder block, pistons, piston rings, connecting rods, crankshafts, gaskets, hubs, gearwheels, linear actuator, oil tank, connections with cooling and oil circuits, hall sensors, and cylinder pressure sensor (Optrand C33233-Q).

Figure 4. Front and back view of the second block

• Extra sets: cylinder block, pistons, piston rings, connecting rods, crankshafts, gaskets, hubs, gearwheels, linear actuator, oil tank, connections with cooling and oil circuits, injectors, rod sensor, and coupling for the rod sensor.

Figure 5. Set of Bosch 0 280 150 995 for Renault models Twingo and Clio

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Figure 6. Set of spare parts

3.2. Fuel

Fuel’s octane number has been lowered from 98 to 95 octanes. This reduction makes the fuel more prone to self ignition.

3.3. Rotary encoder

The original rotary encoder, a Heidenhain sensor model ROD 426, was described on section 3.2.1 in Ref [2]. The big inertia of the rotary parts made the intermediate connection between the crankshaft and the coupling and even the threaded end of the crankshaft to break. It led to a purchase of a smaller model ROD 1000, with 1000 measurements per revolution (every 0.36 degrees).

Figure 7. Comparison between the old and the new rotary encoders

8 3.4. Torque signal filters

The description of torque measurement system can be found on section 3.2.4 in Ref [2]. It consists basically of a force sensor attached to an arm, which is screwed to the housing of a hydrostatic pump moved by the engine through a toothed pulley mechanism, so the pump casing can rotate freely around the impeller, delivering all the reaction torque to the arm. Figure 8 clarifies the concept. A different gear width on one of the toothed belt wheels induced breaks of several belts. After replacement with an adequate gearwheel, torque signal presented a bad behavior since measurements showed cyclic oscillations instead a constant value under a constant break torque (break torque is applied through a valve that regulates the pressure drop on the oil circuit driven by the described hydrostatic pump). Figure 9 shows the cylinder pressure and the measured torque vs time with combustion. Conclusions point to a change of the old belt deformation properties, which allow absorption of vibrations to the torque measurement mechanism.

Figure 8. Torque measuring system

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Figure 9. Cylinder pressure and Torque vs Time

Some different assemblies were tried, but even with a low elastic constant rubber element that dumped the high frequency vibrations on the arm-force sensor coupling , and adding moment of inertia by placing weight a long vertical distance down the rotary housing of the oil pump (figure 10), the structure still showed resonance in operation (figure 11). Once revealed vibrations weren’t interfering with the resonance frequency of the forces sensor (750 Hz according to manufacturer), the natural frequency of the structure was obtained (figure 12), and according to that, it was decided to install a low pass filter with a cutoff frequency of 1 Hz (figure 13), and reestablish the original assembly without rubber element and increased moment of inertia (figure 14). Professor Hans-Erik Ångström selected the cutoff frequency and added the electronic components to the original electronic circuit. Measurements are now more accurate (figure 15).

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Figure 10. Torque measuring structure with low elastic constant and increased moment of inertia.

Figure 11. Cylinder Pressure and Torque vs time with low elastic constant and increased moment of inertia structure. Low frequency oscillation is about 5.4 Hz and high frequency oscillation is about 185 Hz.

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Figure 12. Hit to the torque measuring structure with low elastic constant and increased moment of inertia structure. Natural frequency is about 6 Hz.

300kΩ

300kΩ

Vin 47µF Vout 47µF

Figure 13. 2 stage amplification and 2 stage low pass filter equivalent circuit. The cutoff frequency is 1.11 Hz.

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Figure 14. Original torque measuring structure.

Figure 15. Cylinder pressure and Torque vs Time with electronic filter.

The calibration of the force sensor can be consulted on Appendix 6.

13 3.5. Pressure and temperature sensors on intake manifold and exhaust circuit

Measurements on temperature and pressure of intake and exhaust gases are extraordinary important for understanding the engine operation:

Air conditions on the intake provide information for knowing parameters as described: • Temperature, which will affect the mixing process with the fuel. • Density, which will affect the mass flow entering to the cylinder. • Pressure, which will influence the charge exchange process.

Exhaust gases characteristics also provide are not less important: • Temperature, which can be used for knowing the combustion process or emission characteristics. • Pressure, which will influence the charge exchange process.

Particularly relevant to this work are the exhaust gases conditions: temperature affects the local acoustic velocity on the flow by changing the bulk modulus. Pressure pulses show how the exhaust pipe is behaving. More about this can be read on section 4.

The installation of pressure and temperature sensors on intake and exhaust ports is, as consequence, extremely important. Installed pressure sensors are Gems 2200AGA4001A2UA, while temperature sensors are standard thermocouples. The placement of the sensors was carried out as close as possible to the ports and on a defined position at the exhaust piping (justification can be read on section 4), as showed on figure 16.

Pressure sensors Temperature sensors

Figure 16. Temperature and pressure sensors

14 The exhaust pressure sensors need cooling for operation. Cooling boxes were prepared, taking coolant from the cell installation.

Signals of pressure sensors are read via fast card, while thermocouples measurements are sent with Datascan-Nudam to the PC controller.

Thermocouples don’t need calibration, while calibration procedure of pressure sensors can be consulted on Appendix 5.

3.6. Intake compressed air supply system

3.6.1. Objective

The inlet conditions must be measured in order to determinate its influence on the engine operation, and be able to improve it by modifying them.

The scavenging system was already studied by previous developing teams (section 3.1.2, in Ref [1] or whole contents of Ref [3]). Existent solution was a root compressor connected to the engine via a belt with changeable gear ratio. Modifying the transmission ratio to the root compressor could be done by changing the diameter of the wheels of the transmission system between the intake crankshaft and the root compressor’s shaft, (keeping the size of the tooth constant). Therefore, different speed ratios can be tested. Because it is a positive displacement or volumetric pump, it can be assumed that the delivered volume of air per revolution is constant (losses, volumetric efficiency, air’s density variations and other factors should be taken in consideration), so the mass flow depends basically on the rotational speed, which can be changed in the way that has been described.

Main disadvantages of this system are: • The pump has a limited range of rotational speed and therefore a limited range of delivered flows. Ranges can be consulted on section 4.5 in Ref [3]. • Transmission ratios between the engine and the compressor are limited. For a certain operational speed of the engine, the pump could not deliver the desired air flow. Moreover, several kits of different wheels have to be prepared, and assembly and disassembly periods give discontinuity to the tests. • The inlet temperature cannot be affected unless a system for modifying it is installed.

The solution applied was to use a compressed air installation for supplying directly the intake port: it is possible to simulate the behavior of a scavenging compressor by supplying the engine with compressed air. With the adequate control systems, the air flow and its conditions can be adjusted as it is desired.

15 3.6.2. Design and configuration

The installation was designed as shown in the figure 17. Thermocouple

Manometer Flow meter

From the supply

Stabilizing Pressure tank regulator valve Stabilizing tank Safety valve

To the engine

Air heater

Thermocouple Port pressure sensor

Figure 17. Intake compressed air installation diagram

3.6.2.1. Pressure regulator valve

It consists of a commercial model Festo LRP valve. It allows regulating the intake installation pressure by keeping a constant pressure drop for each degree of openness.

3.6.2.2. Safety valve

For avoiding damage to the intake installation and even personal injuries, the system is provided with a security valve (figure 18) with adjustable opening pressure. The compressed air installation can deliver 6 gauge bar pressure, but the security valve starts opening at 1.5, and gets fully opened at 2.5 bar, not allowing more pressure on the system.

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Figure 18. Disassembly of the homemade security valve

3.6.2.3. Flow meter

It consists of a differential pressure meter, commercial model Setra 265, measuring the pressure drop over a laminar restriction catalyst with some channels plugged). It was already described on section 3.2.4 in Ref [2]. New channels were created on Cell4 for knowing the air flow and Delivery Ratio according to the intake pressure and air flow measurements, whose expressions can be consulted on Appendix 7, as well as the airflow meter calibration procedure.

3.6.2.4. Air heater

The air heater consists of a commercial fan heater, Jula brand, article number 411-089 2009-04-22, with three different configurations: ¾ Stop ¾ Fixed power of 1000 w ¾ Fixed power of 1500 w For getting an exact temperature on the air flow, an output Pulse Width Modulation channel was introduced on Cell4 setup, based on the intake temperature channel, which measurement is given by the described intake temperature sensor in this section.

3.6.3. Adjustment of an intake pressure

The correct way to establish an intake pressure can be done by following next steps: • Close the intake port of the engine by rotating the crankshaft so all the intake system stays under pressure • Measure the pressure using the intake pressure sensor reading on Cell4. Manometer has not an adequate precision so its measurement should be taken as an approximation, and as a security measurement system. • Correct the valve opening by actuating over the manual regulator.

As an indication, intake pressure shouldn´t be bigger than 0.5 bar (gauge pressure).

17 3.7. Oil pump

The implementation of a compressed air supply to the engine carries the substitution of the root compressor, which also consisted on an oil pump attached in the same shaft that provides impulsion to the lubricant, (read on section 4.5 in Ref [3]). To get independence from the dual pump, an individual oil pump was installed: Biltema article 52-3228, with 12 V voltage supply.

Figure 19. Biltema 52-3228 oil pump

Due to bad operation, the ventilation hose of the oil circuit (description on section 2.3.7 in Ref [4]) was removed.

18 3.8. Injection system modifications

The fuel preparation has been studied by changing intake air conditions, and also changing injection characteristics, mainly the injector placement and injection timing.

3.8.1. Bosch 0280150995 injector

The injector used is a commercial Bosch model, code 0 280 150 995, installed on Renault Twingo and Clio. Injector’s selection process can be consulted on section 2.7 in Ref [4]. Its characteristics are shown on table 1.

Table 1. Characteristics of Bosch 0 280 150 995 injector Resistance (Ohm) 14.5 Spray orientation angle - Spray distribution (area, 4°-20° Test medium Heptan spray amount) 70% Operating pressure (kPa) 300 Allowed fuels - Q-stat (at operating pressure) (g/min) 67.5 Version, type 1.3

Q-stat with 300 kPa (ml/min) 96 Electrical connector Jetronic Q-dyn with 300 kPa, locating lug (anti rotation injection time 2,5 ms 1.8 no device) (g/1000 injections) Hydraulic connection to O-Ring Spray type C manifold standard Spray angle alfa-50 Hydraulic connection to fuel O-Ring - (for 2 beam) rail standard Spray angle alfa-80 Distance between O-rings 20° 60.5 (for cone-characteristic) (O-Ring to O-Ring) Axial deviation angle (spray Total length 77 - bent angle)

3.8.2. Injection calibration

The injector operation depends basically of two parameters: Injection Angle and Injection Time. Both of them are implemented on the ECU, which controls the injector by mean of a voltage signal which incorporates these two parameters in the explained way:

• Injection Angle: the voltage is sent according to hall 1 sensor signal, which allows knowing the intake side crankshaft angle since the missed tooth on the gearwheel passes through the sensor a certain known angle before top dead center (“RefAngle missing 1 Ca bef. TDC” on table 3 section 4.1 in Ref [1]).

• Injection Time: signal is maintained during a certain period of time, keeping injector’s needle retracted and allowing fuel flow.

19 The retraction of the needle has an electromechanical delay when opening and closing the injector’s nozzle that depends on the voltage applied to the solenoid actuating the mechanism. Closing delay is usually neglected due to its short duration compared to opening delay. The dependence between delay and voltage is assumed to be linear and was already studied by previous developing team in order to correct the Injection Time. The parameters set on EcoControl are “Inj. Delay Zero Voltage ms” (current value 2.82 ms) and “Zero inj. Delay Voltage” (current value 22.32897 V), which permit to establish dependence of Injection Delay and Supply Voltage shown on figure 20. The description of both parameters is found in Ref [1], where it is specified that only the Injection Time is corrected.

Figure 20. Injection Delay vs Supply Voltage for Bosch 0280150995 injector

Due to the delay, the Injection Time is corrected by the ECU by adding extra time on the tail of each injection, but the Injection Angle does not have any correction. Figure 21 shows the injection signal in a test performed at 5979.4 rpm, 12º constant phase angle, 1.3 ms of Injection Time and Injection Angle of 80.2º.

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Figure 21. Injection Signal vs Crankangle referred to exTDC

As it can be seen, the total injection angular period is 193º-93º=100º, which means an Injection Time of 2.79 ms for the operation speed. The injection voltage can be taken as 12.3 V, which represents 1.27 ms of delay according to figure 20. Therefore, the Injection Time is 2.79≈1.3+1.27. However, the Injection Angle is 93º in an Exhaust Side Crankshaft Angle, which means approximately 81 on an Intake Side Crankshaft Angle, this is, the value set on the ECU. If the Injection Time correction is also considered valid for the Injection Angle, according to figure 20, there is an injection delay of 1.27 ms, or 45.72º at the operation speed. Hence, real Injection Angle is 81º+45.72º=126.72º, while the Injection Signal should be sent on angle 81º-45.72º=35.28º for achieving the desired Injection Angle.

Henceforth the correction for the injection delay applied to the Injection Time would be also considered a valid approximation for the Injection Angle correction.

21 3.8.3. Empirical determination of the Injection Time vs Mass of fuel per injection curve

The fuel flow depends directly on the Injection Time and the engine speed. On the test bench, a scale is used for measuring the mass of the gasoline tank that feeds the injection system, so taking logs of the scale signal allows knowing the fuel flow by calculating the slope of the gasoline tank vs time curve. An important tool for checking and prediction of fuel consumption would be the opposite: a regression curve that calculates fuel flow from the Injection Time. Previous working team configured that curve from injection tests, and it has been compared to values from engine tests performed along the development of the project. Results are shown on figure 22.

Figure 22. Fuel Flow vs Injection Time

As it can be seen, the engine tests show bigger dispersion, but the match between injection and engine tests can be considered quite approximated. Therefore, the regression curve obtained from injection tests can be used for predicting fuel flow with good enough approximation.

22 3.8.4. Comparison between different injection systems

Different configurations for the injector have been tested: • Port Injector • Direct Injector • Axial Deflected Direct Injector Direct injection is the one that potentially allows a better improvement of the engine operation: fuel injection can be done the closest possible to the angle when ports are about to close. This avoids the unburnt mix to exit through the exhaust ports without burning. However, the fuel preparation could be improved by using Port Injection, since more time is given for the mixing process. Port Injection can solve this problem, but then the control of the ports overlap still remains. In this engine, with uniflow-scavenged configuration and placement of the injector with Direct Injection aligned with intake ports, benefits of Direct or Port Injector are not definitively defined. Even another kind of injection, Axial Deflected Direct Injection, has been tested in order to improve injection quality and achieve better air fuel mixture. It has to be noticed that Bosch 0 280 150 995 injector is designed as a Port Injector, but with Direct Injection configuration as described and showed on figure 18, it is not exposed to the extreme conditions of all the combustion cycle, due to the fact that it is exposed to cylinder pressure and temperature only when the injection port opens and the environment is not extremely aggressive. Therefore, it can be used in a Direct Injection mode.

Notice also that the direct injection port timing is not exactly the same as intake ports timing although they are aligned on the same circumference, since the injector port is bigger than intake ports (figure 23). Direct injector port opens at 106º BairTDC while intake ports open at 119º BairTDC. Another important port timing characteristic is that with less than 8º phase difference, the direct injection port opens before the exhaust ports do. Injector Port

Intake Ports Figure 23. Section of the block

23 3.8.4.1. Port Injection

Port Injector was installed as shown on figure 24. No specifically tests for comparing its performance compared to other injection systems have been done.

Direct Injection

Port Injection

Figure 24. Port and Direct Injection

Figure 25 shows results obtained, while table 2 reflects tests characteristics. Notice that stability reflects the engine behavior in a scale from 1 to 5, meaning 1=inoperability; 2=unstable operation (oscillations of the engine speed over 15% of the test velocity); 3=stability with oscillations of the engine speed between 15% and 3% of the test velocity; 4=stability with oscillations of the engine speed between 3% and 1% of the test velocity; and 5 completely smooth operation (oscillations of the engine speed lower than 1% of the test velocity). Table 2. Tests characteristics Port Injection Injection Time 1.6 ms Ignition Angle 65º BairTDC Ignition Ignition coil charge start at 137º BairTDC Injection angles set on EcoControl from 20º to Injection 280º with 20º steps. Delay correction of 42º. Pressure 0.25 bar Intake air Temperature 70ºC Phase 12º constant phase Speed 5500 rpm Not tuned exhaust pipe on its first configuration (Tuned Length 865 mm, tail pipe diameter 19 mm)

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Figure 25. Power and Stability (see previous paragraph for the points scale) vs Injection Angle (Injection Angle referred to airTDC)

The main characteristic revealed while testing was the increase of stability for a same amount of injected fuel when injecting when intake ports are closing or already closed. This could mean that with this kind of injection, mixing process is improved by allowing fuel and air to interact during a longer period of time. Results related on power, however, are difficult to assess.

The lower fuel efficiency and power compared to Direct Injection determined the change to the last one. Further investigation should be done with this type of injection, once the exhaust pipe has been tuned and other improvements have been applied to the engine.

25 3.8.4.2. Direct Injection vs Axial Deflected Direct Injection comparison

A detailed observation of figure 23 would reveal that with Direct Injection the spray is pointed to the intake ports. The fuel preparation could be improved by making the gasoline evaporate faster against warmer surfaces, these are, cylinder walls or exhaust piston head. Axial Deflected Direct Injection consists of a special injector head that deviates the spray a certain angle from the injector longitudinal axis, with the aim of focusing the spray to the combustion chamber and allowing the faster evaporation. Figure 26 shows a detail of normal Direct Injector (DI) and Axial Deflected Direct Injector (Did).

Figure 26. Axial Deflected Direct Injector and Direct Injector

Direct Injection (DI) and Axial Deflected Direct Injection (Did, with an approximated angular deflection of 45º according to the machining of the hole) have been tested in order to determinate the better option. Tests characteristics are shown on table 3, while figures 27 and 28 show Injection Time (in an Intake side crankshaft angle scale), Power and Specific Fuel Consumption for both types of injection. Power to electric and electronic issues such as injection, ignition, phase regulator, electronic circuits, electric lubrication pump, electric cooling pump, electric petrol pump, electric air heater and intake air pressurization are not considered when calculating specific fuel consumption.

Tests have been done trying to keep the same Injection Time and therefore a constant lambda (measured values from 0.82 to 0.92), but when it was not possible, Injection Time was modified till stable operation was reached. Injection angles were swept till the ports closed: bigger Injection angles showed instability or the need of Injection Times longer than 2 ms, so no interesting result was achieved. Also, small inexactitudes on Specific Fuel Consumption can be expected are due to fuel flow calculation by using the scale measurements.

26 Table 3. Tests characteristics Direct Injection- Axial Deflected Direct Injection Ignition Angle 60.1º BairTDC Ignition Ignition coil charge start at 130.1º BairTDC Injection angles set on EcoControl from 10º to Injection 160º AairTDC with 30º steps. Delay correction of 45.6º. Pressure 0.21 bar Intake air Temperature 70ºC Phase 12º constant phase Speed 6000 rpm Tuned exhaust pipe

Figure 27. Power and Injection Time vs Injection Angle (Injection angle referred to airTDC) for both types of injection

27

Figure 28. Specific Fuel Consumption and Injection Time vs Injection Angle (Injection angle referred to airTDC) for both types of injection

Engine shows a not affected by injection angle behavior with not Axial Deflected Direct Injection. It reaches its best specific fuel consumption when fuel is injected while ports are closed, which reveals that better mixing mechanism is similar to 4-stroke Port Injector engines, where closed valves play an important role. Notice that injecting over the piston lateral surface makes the gasoline to evaporate the diluted oil and this worsens the engine lubrication (the engine uses a mix of 2% oil-98% gasoline as fuel for ensuring the pistons rings lubrication). This means that it is dangerous to inject when ports are closed. On the other hand, Axial Deflected Direct Injection behaves better when injecting while ports start opening or are fully open, which can be understood as a different mixing process, which was explained as a faster vaporization process due to impact of the spray with warm surfaces. Since the angular duration of the injection goes from 46.8º for 1.3 ms to 72º for 2 ms (not corrected duration, since the injection delay is already considered), it is an expected result that power decreases and specific fuel consumption increases when injecting from 200º in advance. The injection is stopped because of the closing of the port. It is also remarkable that the injection port timing does not influence the injection as the intake ports timing does. The specific fuel consumption and the power show bad behaviour when injecting at 115.6º AairTDC (70º AairTDC on the ECU). For that Injection Angle, the injection port is open but the intake ports are still closed.

28 3.9. Approaches to two-stroke combustion with high compression ratio

The configuration of the engine has significantly changed along the course of this thesis. After every change, several parameters have been systematically swept in order to achieve an operation that ensures a fast heat release combustion process every working cycle. Ranges of values for some parameters that ensure a good approach to this type of combustion are: • Intake air pressure: from 0.2 to 0.3 bar. • Intake air temperature: from 55 to 70ºC. • Injection angle: depending on the injection system. See previous section. • Ignition angle: from 55 to 65º BairTDC. • Ignition current: taken according to section 5.3 in Ref [2] as 60º before ignition angle at 5000 rpm (2 ms), and changed consequently to every test speed.

The combustion process was already described on section 5.2 in Ref [2]. As it was said, it is a two stroke fast combustion that needs of the spark plug activation energy for initiate the ignition by compression heat.

Next results are shown as an example of the type of combustion performed when using these values. Table 4 describes the test characteristics. Figure 24 shows 20 consecutive cycles of Cylinder Pressure vs Exhaust side crankshaft angle. Figure 25 shows cycles from 6 to 9 of the previous figure.

Table 4. Tests characteristics Port Injection Injection Time 1.6 ms Ignition Angle 56.3º BTDC Ignition Ignition coil charge start at 137º BTDC Injection angle: 244.5º Injection aTDC Pressure 0.25 bar Intake air Temperature 70ºC Phase 12º constant phase Speed 5600 rpm Not tuned exhaust pipe on its first configuration (Tuned Length 865 mm, tail pipe diameter 19 mm)

29

Figure 29. Cylinder Pressure vs Crankangle referred to exTDC

Figure 30. Cylinder Pressure vs Crankangle referred to exTDC for cycles from 6 to 9 of figure 29

30 4. Implementation of an exhaust system

4.1. Exhaust gases flow in a two-stroke engine

The exhaust gases flow in a two stroke engine is an unsteady flow, i.e., the pressure, temperature, and the gas particle velocity are variable with time. This affirmation is valid even with constant speed and load, since internal combustion engine operate on the basis of a reciprocating principium. The operation of an exhaust pipe is based on the propagation of pressure pulses in this kind of flow. See chapter 2 in Ref [5].

Pressure pulses in a two-stroke engine exhaust pipe are created because of the ports opening and closing operation by the piston. This extremely fast perturbation on the flow that moves through the port becomes a finite pressure wave (pulse) that travels along the pipe. This pulse is characterized by having a pressure ratio greater than a sonic wave and a finite spatial extension. When the pulse crosses a transversal section variation, a part of it is transmitted and a part of it is reflected. If the section increases, the reflected pulse switches its sign, while if the section decreases, the reflected pulse continues with its original sign. The pressure of the flow traveling through the exhaust pipe is taken as reference pressure when talking about compression (positive) or expansion (negative) pressure waves.

Figure 31. Compression pulse traveling through a pipe, with the same direction for the flow and the pulse. Source: Ref [7].

31

Figure 32. Expansion pulse traveling through a pipe, with opposite direction for the flow and the pulse. Source: Ref [7]

The length and the amplitude of the transmitted and reflected pulses will depend, among others, on the transversal surface change and its slope referred to the duct´s axial coordinate.

According to this theory, pressure pulses along the exhaust pipe of a two-stroke engine can be used to increase the efficiency of the charge exchange process. The installation of a diffuser-nozzle tune pipe would create a reflected expansion pressure pulse on the section-diverging, which will help the fresh mix from the intake port to enter the cylinder; and a reflected compression pressure pulse would be created on the converging- belly, arriving later than the previous negative pressure pulse, pushing the fresh mix that comes out of the cylinder to the exhaust pipe back. The lack of port timing control through valves can be done by affecting the flow through an exhaust pipe on 2-stroke engines.

4.2. Design of the exhaust pipe

An exhaust pipe must fit certain dimensions for behaving correctly: pressure pulses must affect the exhaust flow adequately. Although the objective of this project was not to build a new pipe according to a design, but implementing a commercial model to the engine, it was important to know how its shape modifies its behavior, which can be studied by designing an exhaust pipe. The design process was based on a set of theoretical-empirical formulas that can be found on section 6.2.5 in Ref [5].

32 Diagram on figure 33 represents the design process.

Figure 33. Exhaust pipe design process

Figure 34 represents a generic exhaust pipe and its dimensions.

Figure 34. Nomenclature of the dimensions of the designed exhaust pipe

The design aims mainly to make the positive pressure pulse originated at the end of the converging section, to arrive to the exhaust port when it is about to get closed. All the previous negative pressure pulses created on the diffuser and the positive pressure pulses generated on the nozzle affect the flow while the exhaust port is open.

The original pressure pulse will travel all the tuned length till the end of the nozzle, where the reflected positive pressure pulse will be generated, and will again cross all the tuned length till the exhaust port. This will be done during the time interval when the exhaust port is open.

33 All the numeric calculations and data from the engine can be consulted on Appendix number 2 (exhaust port opening angle interval) and 3 (exhaust pipe design).

Since pressure pulses are moving on a flow, their propagation speed would be the sum of the local speed of sound plus the gas particle velocity:

α (m / s) = a(m / s) + c(m / s) Expression 1

The value of the local acoustic velocity:

kJ a (m / s) = γ ⋅ R ( ) ⋅T (K) Expression 2 kg ⋅ K where γ is the adiabatic coefficient of the exhaust gasses mix, R is the universal constant for the ideal gasses applied to it, and T its temperature.

Assuming isentropic conditions, the value of the local acoustic velocity on the particles affected by the pressure pulse (see section 2.1.3 in Ref [5]) is:

γ −1 ⎛ p ⎞ 2·γ ⎜ ⎟ Expression 3 a (m / s) = a0 (m / s)·⎜ ⎟ ⎝ p0 ⎠

The value of the particle velocity is, according to Earnshaw (see section 2.1.3 in Ref [5]):

⎡ γ −1 ⎤ 2·a (m / s) ⎛ p ⎞ 2·γ c (m / s) = ± 0 ⎢⎜ ⎟ −1⎥ Expression 4 ⎢⎜ ⎟ ⎥ γ −1 ⎝ p0 ⎠ ⎣⎢ ⎦⎥ where p is the pressure, and subindex 0 indicates the reference state (as it has been said, the conditions of the flow). The sign depends on the pressure ratio: plus when compression waves, and minus when expansion waves.

For the design process, the value of the local acoustic velocity will be taken, according to Blair (section 2.1.3 in Ref [5]), as the propagation speed in the flow (Expression 2), simplifying the calculations, since it is taken a unique propagation speed for all pressure pulses, without considering their pressure ratio. Values of R and γ, which are functions of T and compound properties of the exhaust gas, have been also simplified and can be consulted on Appendix 3

A speed for peak power must be selected, since it determinates the exhaust port opening angle interval, and the exhaust pipe will be tuned for it.

34 Once the speed for peak power and the exhaust port opening angle interval have been selected and calculated, it is easy to know the exhaust port opening time interval:

θ ()º t (s) = Expression 5 ⎛ rev ⎞ 360 ()º / rev N ⎜ ⎟ ⋅ ⎝ min ⎠ ⎛ s ⎞ 60 ⎜ ⎟ ⎝ min ⎠

So, the tuned length must be, according to the following expression:

c (m / s) L (m) = Expression 6 T 2·t ()s

The other dimensions are obtained according to Blair´s empirical formulas. Because of the fact that, as described before, it is not the objective of the project to build an exhaust pipe according to a design but acquiring and changing a commercial model, the part of the process of design left is explained on the Appendix 3, as well as the final dimensions for the obtained design. Only two dimensions will be considered critical: • Tuned length (LT): 867.4 mm. • Tail pipe diameter (D7): 13.2 mm.

4.3. Selection and modification of a commercial exhaust pipe

The design process had been useful to make a first approach of how the dimensions of the exhaust pipe should be, but the objective of the project was to install and modify a commercial exhaust pipe. This ensured no manufacturing malfunctions and had facilitated its adaptation.

The selection of a commercial model had been done by considering different criteria: • Dimensions: it had to be dimensionally similar to the design, at least with a slightly shorter tuned length. • Engines in which it is installed on: it had to work properly installed on engines with analogue characteristics to the Double Two engine: all moped, 2-stroke, 50 cc engines, with the same range of speed as Double Two Engine. • Easy modification: the selected exhaust pipe had to be easy to modify, for changing its dimensions and adapting it to the testing bench. • Economic cost: given the uncertainty of the outcome, not too much expensive pipes had to be considered, while meeting a minimum quality.

35 The selected pipe was Tecnigas Next R.

Figure 35. Tecnigas NextR exhaust pipe

With the aim of tuning the exhaust pipe, the exhaust installation had been modified, introducing: • A first stretch between the block exhaust port and the initial exhaust pipe flange, which allowed changing its total (and so the tuned) length. • An adequate structure for supporting the exhaust pipe from the test bench. • A tail flanged-based-on coupling that allowed changing the tail pipe and holds a flexible pipe that prevented vibrations of the engine to be transmitted to the rest of the exhaust circuit. Silencer had been removed. • New lambda sensor placement in the exhaust circuit, close to the exhaust pipe end. • An intermediate stabilization tank where the emissions probe was attached, and collected exhaust condensates, allowing its removal through a valve on its bottom. • Commercial particulate filters Festo LF-1/8-D-brrr-MINI (5 to 40 µm) and LFMB-1/8-D-brrr-MINI (0.01 to 1 µm), which protected the emissions measurement instruments.

A To the From the B emission exhaust rack pipe Manual

valve

Lambda Duct Filter

sensor pressure

sensor

Stabilyzing To the cell tank exhaust installation Figure 36. Diagram of the exhaust installation

Since the emission rack, Boo Instruments model, had incorporated heated lines and particulate filters, two points of measure, A and B on the diagram on figure 36, were allowed.

36 4.4. Exhaust pipe tuning

The objective of the tuning was to improve the charge exchange process by adjusting the exhaust pipe dimensions. The design process was extremely simplified. The dimensions of the exhaust pipe (specially the designed tuned length) could be empirically modified in order to achieve an improvement of the engine operation.

Charge exchange process can be defined by the use of three parameters: Trapping Efficiency, Delivery Ratio and Charging Efficiency, whose expressions can be found on expressions 7 to 9. mass of delivered fresh air DR = Expression 7 displacement volume* ambient density

mass of fresh gas trapped in the cylinder TrEff = Expression 8 mass of delivered fresh gas

mass of fresh gas trapped in the cylinder ChEff = Expression 9 displacement volume * ambient density

ChEff = DR * TrEff Expression 10

More about these parameters can be read along section 1.5 in Ref [5].

The tuning process had been done for an operational working point of the engine. Delivery Ratio is kept constant by fixing intake pressure and temperature, and rotating at the selected tuning speed. Therefore, Trapping Efficiency will not be affected by changes on intake conditions, and will depend on the exhaust flow. The exhaust flow would depend on combustion process (which determinates its temperature and other conditions) and, as it has been discussed before, on the dimensions of the exhaust pipe. Notice that interaction between charge exchange process, affected by trapping efficiency, and combustion process, makes the tuning process to be iterative. For a certain exhaust flow conditions, a modification on exhaust pipe dimensions would affect the Trapping Efficiency, ensuring more or less fresh mix trapped on the cylinder and so modifying the combustion process, which would make the exhaust flow conditions change again.

Only two dimensions will be considered on the tuning process: • Tuned length • Tail pipe diameter

Changes on these two dimensions would be done according to the observation of pressure pulses travelling along the exhaust pipe.

37 The interpretation of pressure pulses by means of pressure measurements is extremely difficult. In a specific transversal section of the pipe and instant, the usual circumstance is the superposition of different pressure pulses travelling in opposite directions (see section 2.2 in Ref [5]). Because of that, the measurement of a pressure sensor placed somewhere along the pipe will usually represent the amplitude of at least two pressure pulses superposed. However, some attempts to identify correctly which section variation generates each peak on the pressure measurement can be tried, by using a simply cinematic relation:

L(m) a (m / s) = Expression 11 t(s)

In a crankshaft angle log, with a known speed of the engine, angles can be turned to times, and each peak can be translated to a traveled length along the pipe. If the pressure pulse corresponding to the end of the diffuser reaches the sensor when the exhaust port is almost closed (notice that there is a little distance between the sensor and the exhaust port that should not be mispriced), then the tuning is perfect. But the main problem to make these calculations is the propagation speed of pressure pulses: each pressure pulse, according to expression 3, has a different propagation velocity because of the different pressure ratios. Although despising particle speed and considering only the local acoustic velocity without pressure ratio, the temperature introduced on expression 2 can not be taken as exhaust temperature sensor measurement, since it represents an average of the different temperatures experienced in its position during the whole working cycle of the engine. According to Heywood (section 6.5 in Ref [6]), exhaust gas temperature can vary even a 50% during a cycle (although the referenced tests were performed on a 4-stroke spark ignited engine, results can be applied also to 2-stroke engines).

Due to the difficulties to establish a proper exhaust flow temperature, the pressure pulses’ propagation velocity was obtained by using a second sensor placed on a known distance from the first one on the exhaust installation, so the first opening pressure pulse can be identified in both sensors, and angle which it takes place on is obtained.

Some important considerations should be observed for interpretation of crankshaft angle log with pressure measurements: • The sensor is not exactly placed close to the cylinder’s exhaust ports, but on the block exhaust port. Since pressure logs even reflect the different initial pulses due to the different circumferential position of the cylinder ports and so the different distance to sensor, angle for the initial pulse would be taken between second and third of the total of four peaks. • Gems 2200AGA4001A2UA pressure sensor has a delay of 0.1 ms, so the data collected by the fast card from these pressure sensors at a certain angle is actually measured 0.1 ms before (the conversion to an angle is easy by knowing the angular speed of the engine during the log).

38 By keeping the engine configuration and working point unchanged, a fine tuning will reveal: • Increase of Trapping Efficiency • Increase of Torque • Increase of Power • Increase of Fuel Efficiency • Decrease of Specific Fuel Consumption So these five parameters will be measured and calculated for every change on the exhaust pipe dimensions during the tuning process.

Trapping Efficiency must be calculated with an Oxygen balance in rich conditions, by knowing exhaust and inlet concentrations (see section 1.6.3 in Ref [5]):

C8H18 + a·(O2 + N2) → b·CO2 + c·H2O + d·N2 + f·C8H18 + g·(O2 + N2)Non trapped

mass g O2 g exhaust []O2 exhaust ( ) ⋅ m& exhaust ( ) g air g exhaust min m& air int ake ( ) − min mass g O []O ( 2 ) 2 int ake g air TrEff = Expression 12 g air m& air int ake ( ) min

The need of a rich lambda for calculating the Trapping Efficiency determines low fuel efficiencies and high specific fuel consumptions. It has to be considered that the aim of the calculation of these two parameters is comparing the behavior of the different configurations of the exhaust pipe, and not bringing the engine to an efficient working point. Power to electric and electronic issues such as injection, ignition, phase regulator, electronic circuits, electric lubrication pump, electric cooling pump, electric petrol pump, electric air heater and intake air pressurization are not considered when calculating specific fuel consumption and fuel efficiency.

39 5. Results

5.1. Tests procedure

Configuration and working point of the engine are described on table 5.

Table 5. Engine configuration for the tests Injection Angle 80º BTDC Direct Injection Injection Time 1.5 ms Ignition Angle 60.1º BTDC Ignition Ignition coil charge start at 130.1º BTDC Pressure 0.21 bar Intake air Temperature 70ºC Phase 12º constant phase Speed 6000 rpm

As it was said on previous section, Injection Time has been chosen in order to achieve rich combustion. Lambda measured values oscillated between 0.77 and 0.86.

5.2. Exhaust pipe tuning results

5.2.1. Tuned Length

Figure 37 shows the initial dimensions of the exhaust pipe. Tuned length was taken according to the design done on section 4.2.

Port Duct Sensor Sensor

Ø19mm

L1 L1’ L2 L3 L5 L6 L7

140 300 90 110 60 165 130

1050 Figure 37. Dimensions of the initial configuration of the exhaust pipe

40 Figure 38 shows measurements of pressure transducers over Exhaust port Total Surface Opened vs Crankangle referred to exTDC.

Figure 38. ETSO (Exhaust port Total Surface Opened), Exhaust Port Pressure and Duct Pressure vs Crankangle referred to exTDC, for 865 mm of tuned length

As it can be seen, initial opening pulse reaches the port sensor on angle 154º, while arrives to the duct sensor aproximately on angle 239º. This means that the average propagation speed is 445 m/s. According to this speed and the angle for the reflected pulse of the converging section, 306º, the exhaust pipe should be 370 mm shorter for fitting the last pulse when the exhaust port is about to close. This calculation doesn’t consider changes on exhaust flow temperature, which will change when trapping efficiency increases and combustion and charge exchange processes change, as said before.

A deeper analysis would reveal that reflected compression and expansion pulses arrive to exhaust port when supposed according to the calculated propagation speed.

41 Final configuration after dimensional modifications become as shown on figure 39.

Port Duct

Sensor Sensor

Ø19mm

L1 L1’ L2 L3 L5 L6 L7

85 120 90 110 60 165 130

965

Figure 39. Dimensions of the final configuration of the exhaust pipe

Pressure transducers measurements over Exhaust port Total Surface Opened vs Crankangle referred to exTDC are shown again on figure 40.

Figure 40. ETSO (Exhaust port Total Surface Opened), Exhaust Port Pressure and Duct Pressure vs Crankangle referred to exTDC, for 630 mm of tuned length

As it can be seen, compression pulse reflected from the nozzle reaches now the exhaust port exactly when it closes.

42 Evolution of Fuel Efficiency, Power, Torque, Specific Fuel Consumption and Trapping Efficiency is shown on figure 41.

Figure 41. Evolution of Fuel Efficiency, Power, Torque, Specific Fuel Consumption and Trapping Efficiency vs dimensional change on tuned length

Improvement of the engine operation can be appreciated.

43 5.2.2. Tail pipe diameter

Once the exhaust pipe has been tuned according to the tuned length, the tail pipe diameter influence would be tested for noticing its influence on the engine behavior.

Results on the pressure transducer measurements over Exhaust port Total Surface Opened for the original 19 mm tail pipe diameter were already shown on figure 40. Pressure transducers measurements over Exhaust port Total Surface Opened vs Crankangle referred to exTDC for 14 mm of tail pipe diameter are shown on figure 42.

Figure 42. ETSO (Exhaust port Total Surface Opened), Exhaust Port Pressure and Duct Pressure vs Crankangle referred to exTDC, for 19mm of tail pipe diameter

First fo all, notice that the four characterisitc pressure peaks correspondent to the port opening have turn to two. The reason is that the graph is built from an average of 20 cycles. Individual cycles show the previous behavior.

As it can be seen, changing the tail pipe diameter does not almost affect propagation of the pulses, since they reach the port at about the same angles. Although pressure ratios have increased notably (notice the change of vertical scale from this graph to the previous ones), from expression 3 it can be deduced that these does not modify notoriusly the propagation velocity. On that expression, pressure ratio should be huge for compensating the influence of the exponent.

44 Figure 43 shows the evolution of Fuel Efficiency, Power, Torque, Specific Fuel Consumption and Trapping Efficiency when tail pipe diameter is changed.

Figure 43. Evolution of Fuel Efficiency, Power, Torque, Specific Fuel Consumption and Trapping Efficiency vs dimensional changes on tail pipe diameter

The smaller diameter seems to slightly worsen the engine operation.

5.3. Comparison between the old exhaust pipe designed with 1-D simulation software and the new tuned commercial exhaust pipe

The old exhaust pipe designed with 1-D simulation software looks dimensionally very different to the tuned commercial pipe. Figure 44 shows its really high tuned length (960 mm), as well as its little section variation compared to the commercial pipe (with a tuned length of 630 mm).

Figure 44. Commercial exhaust pipe (left) and designed exhaust pipe (right)

45 Figure 45 shows the Cylinder Pressure and the Exhaust Port Pressure vs Crankangle referred to exTDC for both pipes. The Injection Time was lowered to 1.1 ms for the designed exhaust pipe to reach a stable operation. Lambda did not almost experience change, since for the same intake pressure, the air flow decreased from 150 to 135 l/min.

Figure 45. Cylinder Pressure and Exhaust Port Pressure vs Crankangle referred to exTDC for the commercial and the designed exhaust pipes

It is a remarkable observation that although both tuned lengths are much different, the compression pulse reaches the exhaust port when it should, so both are correctly tuned. However, the 1-D simulation software designed pipe produces more cycle to cycle variation. It does not ensure one combustion per cycle, and the maximum pressure for the average of is lower than the commercial pipe one. It can be also notice that for the designed exhaust pipe, there is a backflow for the first pressure peak after the opening of the exhaust port, since the intake port is already open and the intake pressure oscillates always under 0.21 bar.

46 6. Conclusions

The 2-stroke Otto operation of the engine has been studied and improved.

Charge exchange process has been deeply developed: • The new intake compressed air installation allows changes on inlet conditions: pressure and temperature have been swept and values given on tests ensured an improved combustion process. • Air-fuel mixing has been studied and different injection systems have been tested, giving a first approximation of the issue of where, when, and how to inject. Direct injection has delivered better results than port injection. However, the injection system can be still improved with the concept of axial deflected direct injection. • The exhaust pipe has been tuned and makes the engine operation to be more efficient and smoother. The engine has shown a remarkable improvement when the tuned length of the exhaust pipe has been adjusted. Trapping efficiency has increased a 15% from the not tuned pipe. This justifies the importance of the correct tuned length for the exhaust system. Analysis of pressure pulses has been an important tool for achieving this development. Tail pipe diameter tests do not provide a definitive conclusion of the influence of this dimension on the engine behavior. Apparently, the bigger diameter provides better results.

The main and most notorious improvement for the engine during the development of this thesis has been the increase of reliability for both the engine and the test bench. Many sensors have been recalibrated, new ones have been introduced, and some measuring systems have been optimized. The prevention of failures (preparation of a second block, spare parts, etc) has enabled a more efficient process of investigation. Test continuity has increased exponentially along the course of this thesis.

The learning process about the interpretation and prediction of the engine operation has encouraged the team to reach more efficient solutions to each problem that arises. The planning, close collaboration with the mechanics, the collective discussion of problems, and team work have leaded to a the prosecution of an objective of which all the previous developing teams should fell holders as well: the participation on Shell Ecomarathon 2010, where the Agilis-HCCI-powered car drove for more than 22 km and achieved the Shell Technical Innovation Award.

47

Figure 46. Agilis-V team and car with the Shell Technical Innovation Award in Lausitzring, Germany

48 7. Future work

The tolerances between the gearwheels and shafts of the compression ratio control system cause a big fraction of the engine failures. Although the mechanical problems have been reduced and the engine operation is more secure, a deeper study of the mechanical transmission should be done in order to prevent vibrations and improve the engine performance. . It is required an extensive study with the objective to determinate the exact tolerances for allowing a proper operation of the crankshaft-gearwheels transmission.

The torque measuring system shows a small variability whenever a change on the test bench or engine is introduced. Improvement of this system or skipping the gearwheel/belt transmission is recommended.

The main measuring computer shows slowdown operation when operating different regulators. This affects the engine operation, specially the speed regulator. A change for a faster processor is required.

The fuel preparation and mixing process have to be studied and improved. Values of temperature, pressure, flow, etc, for both air and fuel have to be swept and optimized.

The interactions between spark ignition and compression detonation must be studied in order to achieve a SI to HCCI transfer procedure. Combustion can still be optimized.

The charge exchange process can be still developed for obtaining different configurations through combinations of Delivery Ratios and Trapping Efficiencies.

49 8. References

[1] David Larsson, “Development of 22” Master of Science Thesis MMK2008:36 KTH Industrial engineering and management

[2] John Larsson, “Development of a counter piston two stroke HCCI engine” Master of Science Thesis MMK2009:74 KTH Industrial engineering and management

[3] Kim Jaktlund, “Development of the HCCI engine” Master of Science Thesis MMK2010:01 KTH Industrial engineering and management

[4] Àlez Poyo Muñoz, “Design and assembly of a throttle for an HCCI engine” Master of Science thesis MMK 2009

[5] Gordon P. Blair, “Design and simulation of two-stroke engines” SAE international, 1996

[6] John B. Heywood, “Internal Combustion Engine Fundamentals” McGraw-Hill

[7] Ingeniería e Investigación vol.29 n.1 Bogotá Jan./Apr. 2009 Internal combustion engine exhaust pipe flow simulation. Part I: theoretical aspects http://www.scielo.org.co/scielo.php?pid=S0120- 56092009000100015&script=sci_arttext

50 Appendix 1: Demonstration concerning the Internal DC

θex crank

EPP

Figure 47. Crankshaft, connecting rod and piston assembly, with its dimensions and angular coordinates. All the units are mm.

According to figure 47, the EPP (Exhaust Piston Position) is:

c·sin(θ ex crank ) EPP = h + c·cos(θ ) + l·cos(arcsin( ) Expression 13 ex crank l

Taking in consideration some of the explanations of section 1, the IPP (Intake Piston Position) is: c·sin(θ ) IPP = h + c·cos(θ ) + l·cos(arcsin( in crank ) = in crank l Expression 14 c·sin(θ − phase) h + c·cos(θ − phase) + l·cos(arcsin( ex crank ) ex crank l

The volume of the cylinder contained between the two piston faces can be calculated as:

⎛ D 2 ⎞ ⎜ ⎟ Expression 15 V = (π·⎜ ⎟)·(L − EPP − IPP) +Vspark plug & pressure sensor ⎝ 4 ⎠

51 For getting the minimum volume, the first and second derivatives are calculated on expressions 16 and 17.

dV dEPP dIPP = − − = dθ ex crank dt dt

h·sin(θ ex crank ) 1 h − c·sin(θ ex crank ) − l·sin(arcsin( )· · cos()θ ex crank l c 2·sin 2 ()θ l 1− ex crank l 2

− c·sin(θ ex crank − phase)

h·sin(θ ex crank − phase) 1 h − l·sin(arcsin( )· · cos()θ ex crank − phase l c 2·sin 2 ()θ − phase l 1− ex crank l 2 Expression 16 d 2V 2 = dθex crank

− c·cos(θex crank ) − c·cos(θex crank − phase) 2 ⎛ ⎞ ⎜ ⎟ 2 h·sin(θex crank ) ⎜ 1 ⎟ ⎛ h ⎞ − l·cos(arcsin( )·⎜ ⎟ ·⎜ cos()θex crank ⎟ l 2 2 ⎝ l ⎠ ⎜ c ·sin ()θex crank ⎟ ⎜ 1− 2 ⎟ ⎝ l ⎠ ⎛ ⎜ 2 h·sin(θex crank ) ⎜ 1 c ·sin ()θex crank h − l·sin(arcsin( )· ·2· cos2 θ ⎜ 2 2 2 ()ex crank l 3 c ·sin θ l l ⎜ 2 ()ex crank ⎜ 1− 2 ⎝ l ⎞ ⎟ 1 h ⎟ − · sin()θ 2 2 l ex crank ⎟ c ·sin ()θex crank ⎟ 1− 2 ⎟ l ⎠ 2 ⎛ ⎞ ⎜ ⎟ 2 h·sin(θex crank − phase) ⎜ 1 ⎟ ⎛ h ⎞ − l·cos(arcsin( )·⎜ ⎟ ·⎜ cos()θex crank − phase ⎟ l 2 2 ⎝ l ⎠ ⎜ c ·sin ()θex crank − phase ⎟ ⎜ 1− 2 ⎟ ⎝ l ⎠

52 ⎛ ⎜ h·sin(θ − phase) ⎜ 1 c2·sin ()θ − phase h − l·sin(arcsin( ex crank )· 2 ex crank cos2 θ − phase ⎜ 2 2 2 ()ex crank l 3 c ·sin θ − phase l l ⎜ 2 ()ex crank ⎜ 1− 2 ⎝ l ⎞ ⎟ 1 h ⎟ − · sin()θ − phase 2 2 l ex crank ⎟ c ·sin ()θex crank − phase ⎟ 1− 2 ⎟ l ⎠ Expression 17

Trying with phase/2 as θex crank in the first derivative the solution becomes 0, as it can be seen on expression 18.

dV = 0 = −c·sin( phase ) − c·sin(− phase ) 2 2 dθ ex crank h·sin( phase ) 2 1 h phase − l·sin(arcsin( )· · cos⎜⎛ ⎟⎞ l l ⎝ 2 ⎠ c 2 ·sin 2 ⎛ phase ⎞ ⎜ 2 ⎟ 1− ⎝ ⎠ l 2 h·sin(− phase ) 2 1 h phase − l·sin(arcsin( )· · cos⎜⎛− ⎟⎞ l l ⎝ 2 ⎠ c 2 ·sin 2 ⎛− phase ⎞ ⎜ 2 ⎟ 1− ⎝ ⎠ l 2 Expression 18 While the second derivative of the volume evaluated in the same value is found on expression 19.

53 d 2V 2 = dθ ex crank − c·cos( phase ) − c·cos(− phase ) 2 2 2 ⎛ ⎞ ⎜ ⎟ ⎜ ⎟ phase 2 h·sin( ) ⎜ 1 ⎟ ⎛ h ⎞ − l·cos(arcsin( 2 )· · cos⎛ phase ⎞ ⎜ ⎟ ⎜ ⎜ 2 ⎟⎟ l 2 2 phase ⎝ l ⎝ ⎠⎠ ⎜ c ·sin ⎜⎛ ⎟⎞ ⎟ ⎜ ⎝ 2 ⎠ ⎟ ⎜ 1− 2 ⎟ ⎝ l ⎠ ⎛ ⎜ ⎜ 2 ⎛ phase ⎞ h·sin( phase ) 2·c ·sin⎜ ⎟ 2 ⎜ 1 ⎝ 2 ⎠ h 2 phase − l·sin(arcsin( )· cos ⎜⎛ ⎟⎞ ⎜ 2 ⎝ 2 ⎠ l ⎜ 2 2 ⎛ phase ⎞ l l 3 c ·sin ⎜ ⎟ ⎜ 2 ⎝ 2 ⎠ ⎜ 1− 2 ⎝ l

⎞ ⎟ ⎟ 1 h phase ⎟ − · sin⎜⎛ ⎟⎞ l ⎝ 2 ⎠⎟ c2·sin 2 ⎛ phase ⎞ ⎜ 2 ⎟ ⎟ ⎝ ⎠ ⎟ 1− 2 ⎟ l ⎠ 2 ⎛ ⎞ ⎜ ⎟ phase ⎜ ⎟ 2 h·sin(− ) ⎜ 1 ⎟ ⎛ h phase ⎞ − l·cos(arcsin( 2 )· ·⎜ cos⎜⎛− ⎟⎞⎟ l ⎜ ⎟ l ⎝ 2 ⎠ c2·sin 2 ⎛− phase ⎞ ⎝ ⎠ ⎜ ⎜ 2 ⎟ ⎟ ⎜ ⎝ ⎠ ⎟ ⎜ 1− 2 ⎟ ⎝ l ⎠ ⎛ ⎜ phase ⎜ 2·c2·sin⎛− phase ⎞ h·sin(− ) 1 ⎜ 2 ⎟ h − l·sin(arcsin( 2 )·⎜ ⎝ ⎠ cos2 ⎛− phase ⎞ ⎜ 2 ⎜ 2 ⎟ l 2 2 ⎛ phase ⎞ l l ⎝ ⎠ ⎜ 3 c ·sin ⎜− ⎟ 2 2 ⎜ ⎝ ⎠ ⎜ 1− 2 ⎝ l ⎞ ⎟ ⎟ 1 h phase ⎟ − · sin⎜⎛− ⎟⎞ > 0 l ⎝ 2 ⎠⎟ c2·sin 2 ⎛− phase ⎞ ⎜ 2 ⎟ ⎟ ⎝ ⎠ ⎟ 1− 2 ⎟ l ⎠ Expression 19

54 Appendix 2: Geometrical description MS Excel worksheet

1. Engine dimensions

Figure 48. block dimensions. All the units are mm.

Figure 49. Section of the block showing the intake ports.

55

Figure 50. Section of the block showing the exhaust ports.

56 Table 6. Engine geometrical data

Intake half side geometrical data Exhaust half side geometrical data

Bore (D) 34 mm Bore (D) 34 mm Stroke 27 mm Stroke 27 mm Connecting Rod Length (l) 48 mm Connecting Rod Length (l) 48 mm

Piston height (h1) 18.75 mm Piston height (h2) 18.75 mm Connecting rod-crankshaft distance (c) 13.5 mm Connecting rod-crankshaft distance (c) 13.5 mm

Intake Port Beginning-crankshaft distance (b1) 53.75 mm Exhaust Port Beginning-crankshaft distance (b2) 53.75 mm

Intake Port Ending-crankshaft distance (e1) 58.75 mm Exhaust Port Ending-crankshaft distance (e2) 59.75 mm

Number of intake ports (n1) 5 - Number of exhaust ports ( n2) 4 -

Width of each intake port (w1) 5 mm Width of each exhaust port (w2) 6 mm

Radii of the corners of each intake port (r1) 1 mm Radii of the corners of each exhaust port (r2) 2 mm

Intake port angle (a1) 36.5 º Exhaust port angle (a2) 41.34 º

Intake Port Area 266.5 mm2 Exhaust Port Area 280.6 mm2

Other geometrical data of the block

Distance between crankshafts (L) 162.5 mm

3 Sparkplug and pressure sensor cavities volume (Vspark plug & pressure sensor)94mm

57 2. Calculations

As explained in appendix 1:

c·sin(θ ex crank ) EPP = h + c·cos(θ ) + l·cos(arcsin( ) Expression 20 1 ex crank l c·sin(θ ) IPP = h + c·cos(θ ) + l·cos(arcsin( in crank ) = 2 in crank l Expression 21 c·sin(θ − phase) h + c·cos(θ − phase) + l·cos(arcsin( ex crank ) 2 ex crank l

The ITSO (Intake port Total Surface Opened) and ETSO (Exhaust port Total Surface Opened) are calculated according to “if” programming sentences, which can be simplified in the way explained below.

For ITSO: • If IPP> e1→ ITSO = 0

⎡ a1 • If e1>IPP>(e1-r1)→ ITSO = n1·⎢(e1 − IPP)·(π·D· ) − 2·(e1 − IPP)·r1 ⎣ 360 r 2 ⎪⎧⎛ ⎡ ⎛ IPP − ()e − r ⎞⎤⎞ + 2· 1 ·⎨⎜[]arcsin()1 − ⎢arcsin⎜ 1 1 ⎟⎥⎟ 2 ⎜ ⎜ r ⎟ ⎟ ⎩⎪⎝ ⎣ ⎝ 1 ⎠⎦⎠ 1 ⎛ ⎡ ⎛ IPP − ()e − r ⎞⎤⎞⎪⎫⎤ + ⎜[]sin()2·arcsin(1 − ⎢sin(2·arcsin⎜ 1 1 ⎟⎥⎟⎬⎥ Expression 22 2 ⎜ ⎜ r ⎟ ⎟ ⎝ ⎣ ⎝ 1 ⎠⎦⎠⎭⎪⎦⎥ 2 ⎡ a1 2 π·r1 ⎤ • If (e1-r1)>IPP>(b1+r1)→ ITSO = n1·⎢(e1 − IPP)·(π·D· ) − 2·r1 + 2· ⎥ ⎣ 360 4 ⎦ 2 ⎡ a1 2 π·r1 • If (b1+r1)>IPP>b1→ ITSO = n1·⎢(e1 − IPP)·(π·D· ) − 2·r1 + 2· ⎣ 360 4 r 2 ⎧⎛ ⎡ ⎛ IPP − b + r ⎞⎤ ⎞ 1 ⎪⎜ ()1 1 ⎟ − 2·(b1 + r1 − IPP)·r1 + 2· ·⎨ ⎢arcsin⎜ ⎟⎥ − []arcsin()0 2 ⎜ ⎜ r ⎟ ⎟ ⎩⎪⎝ ⎣ ⎝ 1 ⎠⎦ ⎠ 1 ⎛ ⎡ ⎛ IPP − ()b + r ⎞⎤ ⎞⎪⎫⎤ + ⎜ ⎢sin⎜2·arcsin( 1 1 ⎟⎥ − []sin(2·arcsin()0 ⎟⎬⎥ Expression 23 2 ⎜ ⎜ r ⎟ ⎟ ⎝ ⎣ ⎝ 1 ⎠⎦ ⎠⎭⎪⎦⎥ 2 ⎡ a1 2 π·r1 ⎤ • If b1>IPP→ ITSO = n1·⎢(e1 − b1 )·(π·D· ) − 4·r1 + 4· ⎥ Expression 24 ⎣ 360 4 ⎦

58 For ETSO: • If EPP> e2→ ETSO = 0

⎡ a2 • If e2>EPP>(e2-r2)→ ETSO = n2·⎢(e2 − IPP)·(π·D· ) − 2·(e2 − IPP)·r2 ⎣ 360 r 2 ⎪⎧⎛ ⎡ ⎛ EPP − ()e − r ⎞⎤⎞ + 2· 2 ·⎨⎜[]arcsin()1 − ⎢arcsin⎜ 2 2 ⎟⎥⎟ 2 ⎜ ⎜ r ⎟ ⎟ ⎩⎪⎝ ⎣ ⎝ 2 ⎠⎦⎠ 1 ⎛ ⎡ ⎛ EPP − ()e − r ⎞⎤⎞⎪⎫⎤ + ⎜[]sin()2·arcsin(1 − ⎢sin(2·arcsin⎜ 2 2 ⎟⎥⎟⎬⎥ Expression 25 2 ⎜ ⎜ r ⎟ ⎟ ⎝ ⎣ ⎝ 2 ⎠⎦⎠⎭⎪⎦⎥ 2 ⎡ a2 2 π·r2 ⎤ • If (e2-r2)>EPP>(b2+r2)→ ETSO = n2·⎢(e2 − EPP)·(π·D· ) − 2·r2 + 2· ⎥ ⎣ 360 4 ⎦ Expression 26 2 ⎡ a2 2 π·r2 • If (b2+r2)>EPP>b2→ ETSO = n2·⎢(e2 − EPP)·(π·D· ) − 2·r2 + 2· ⎣ 360 4 r 2 ⎧⎛ ⎡ ⎛ EPP − b + r ⎞⎤ ⎞ 2 ⎪⎜ ()2 2 ⎟ − 2·(b2 + r2 − EPP)·r2 + 2· ·⎨ ⎢arcsin⎜ ⎟⎥ − []arcsin()0 2 ⎜ ⎜ r ⎟ ⎟ ⎩⎪⎝ ⎣ ⎝ 2 ⎠⎦ ⎠ 1 ⎛ ⎡ ⎛ EPP − ()b + r ⎞⎤ ⎞⎪⎫⎤ + ⎜ ⎢sin⎜2·arcsin( 2 2 ⎟⎥ − []sin(2·arcsin()0 ⎟⎬⎥ Expression 27 2 ⎜ ⎜ r ⎟ ⎟ ⎝ ⎣ ⎝ 2 ⎠⎦ ⎠⎭⎪⎦⎥ 2 ⎡ a2 2 π·r2 ⎤ • If b2>EPP→ ETSO = n1·⎢(e2 − b2 )·(π·D· ) − 4·r2 + 4· ⎥ Expression 28 ⎣ 360 4 ⎦

It can be demonstrated that the error made by taking the circular surface of the ports as planar and not curve is not so big.

Figure 51. Section of the block showing the exhaust ports and the detail of circular surface of the ports.

59 ⎛ ⎛ r ⎞ ⎞ ⎜ 1 ⎟ 2·arcsin⎜ ⎟ ⎜ ⎝ D 2 ⎠ ⎟ π·D·⎜ ⎟ − 2·r1 ⎜ 360 ⎟ ⎜ ⎟ ⎝ ⎠ 2.0011552 − 2 ε = ·100 = ·100 = 0.0577% Expression 29 1 2.0011552 ⎛ ⎛ r1 ⎞ ⎞ ⎜ 2·arcsin⎜ ⎟ ⎟ ⎜ ⎝ D 2 ⎠ ⎟ π·D·⎜ ⎟ ⎜ 360 ⎟ ⎜ ⎟ ⎝ ⎠ ⎛ ⎛ r ⎞ ⎞ ⎜ 2 ⎟ 2·arcsin⎜ ⎟ ⎜ ⎝ D 2 ⎠ ⎟ π·D·⎜ ⎟ − 2·r2 ⎜ 360 ⎟ ⎜ ⎟ ⎝ ⎠ 2.0011552 − 2 ε 2 = ·100 = ·100 = 0.2316% Expression 30 ⎛ ⎛ r2 ⎞ ⎞ 2.0011552 ⎜ 2·arcsin⎜ ⎟ ⎟ ⎜ ⎝ D 2 ⎠ ⎟ π·D·⎜ ⎟ ⎜ 360 ⎟ ⎜ ⎟ ⎝ ⎠

The volume of the cylinder contained between the two piston faces can be calculated, as said on Appendix 1, with the expression 31.

⎛ D 2 ⎞ ⎜ ⎟ Expression 31 V = (π·⎜ ⎟)·(L − EPP − IPP) +Vspark plug & pressure sensor ⎝ 4 ⎠

Trapping Compression Ratio is calculated with excel programming, by looking for the minimum volume and the volume when both ports are closed in the compression stroke, expression 32.

Vtrapped TCR = Expression 32 Vmin imum

Many other calculations can be easily done over IPP and EPP with MS excel programming, such as the angular period where one port is open, the angle interval during which both ports are open, etc.

60 3. Graphs

Figure 52. Trapping and Geometric Compression Ratio vs Phase Angle

Figure 53. Exhaust and Intake ports Total Surface Opened, Pistons´ height, and Volume between pistons vs Intake side Crankshaft Angle

61 Appendix 3: Exhaust Pipe designer MS Excel worksheet

The process of design has been described on section 4.2.

Figure 54. Design process

Needed data was calculated as explained: • Speed for engine´s peak power was decided at 6000 rpm. Criteria were based on the experience with the engine, which shows stable behavior and lack of mechanical failures from 4500 to 6500 rpm. • Exhaust port opening angle interval can be taken from the Geometrical Description worksheet, by using excel programming. • Temperature of the exhaust flow is taken according to measurements as 280ºC. R and γ are function of this temperature and the mix fractions (see section 2.1.6 in Ref [5]). They were took as γ=1.362 and R=286.4 J/kg·K according to table 2.1.3 on the previous reference.

Once the tuned length and the equivalent initial diameters are found (expression 6), the design is based on empirical expressions for calculating lengths and internal diameters of each stretch.

62

Figure 55. Nomenclature of the dimensions of the designed exhaust pipe

Lengths are obtained according to these coefficients applied to the tuned length (see table 6). Table 7. Calculations of the stretches’ lengths KL1 KL2 KL3 KL4 KL5 KL6 KL7 0.1 0.275 0.183 0.092 0.11 0.24 0.24

Diameters D1, D4 and D7 have to be calculated by coefficients applied to the equivalent diameter of the exhaust port area (D0) (see table 7).

Table 8. D1, D4 and D7 calculations Broadly tuned, enduro type 1.125 engines KD 1 Road-racing with the 1.05 highest specific output. Broadly tuned, enduro type 2.125 engines KD 4 Road-racing with the 3.25 highest specific output. 0.6 Track racing (11 bar bmep) KD7 0.65 Track racing (9 bar bmep) 0.7 Track racing (8 bar bmep)

While diameters D2 and D3 are calculated by coefficients defined in a more complex way, multiplying D1 (consult table 8).

Table 9. D2 and D3 calculations L + L ⎛ d ⎞ x12 2 3 kh ⎜ 4 ⎟ D2 d 2 = d1·e with x13 = ( ) ·log e ⎜ ⎟ L2 + L3 + L4 ⎝ d1 ⎠ L ⎛ d ⎞ x12 2 kh ⎜ 4 ⎟ D3 d 3 = d1·e with x12 = ( ) ·log e ⎜ ⎟ L2 + L3 + L4 ⎝ d1 ⎠

63

With 1.25

Selected constants for diameter calculations are shown on table 9:

Table 10. Design constants

Ktailp pipe 0.7

K1 1.125

K2 2.22

Kh 1.5

Final results of the design are shown on table 10:

Table 11. Designed exhaust pipe Tuned Pipe Length 867.4 mm

L1 86.7 mm d1 21.3 mm L2 238.5 mm d2 35.4 mm L3 158.7 mm d3 35.6 mm L4 79.8 mm d4 42.0 mm L5 95.4 mm d7 13.2 mm L6 208.2 mm L7 208.2 mm

64 Appendix 4: Calibration of the Optrand C33233-Q pressure transducer

The description of Optrand C33233-Q pressure sensor can be found on section 3.2.3 in Ref [2] (notice that the old code was C32233-Q). Since it is a dynamic sensor, the static calibration procedure needs of some particularities that will be commented: • Prepare the sensor with a standard connection, following the scheme on figure 56 and table 12.

2 2 3 3 1 1 4 4

Figure 56. Scheme of connections on the sensor side

Table 12. Wires, its functions, and the pins of the connector where they should be attached Color Function Pin number Red Power Supply 1 White Signal 2 Bare Case ground-shield 3 Black Ground 4 Green Diagnostic-static calibration -

The potentiometer should be a 2 kΩ twenty turn type.

• Couple correctly the sensor to the calibration circuit by fitting the seating ring of the sensor (5.8 mm diameter). Ensure that there isn´t (too much) air between the sensor membrane and the oil level, by taking the last to an adequate height. On this purpose, the use of the calibration circuit should be learnt: valve 1 allows flow from the tank, valve 2 permits the oil to reach the open-end of the circuit (that would be closed by the sensor), while the intermediate volume can be filled or emptied by turning the bolt. So close the open-end valve, open the tank valve, fill the intermediate volume by unscrewing the bolt, close the tank valve, open the open-end valve, and screw the bolt to making the oil of the intermediate volume to reach and adequate level on the open end. If more oil is needed, repeat the operation.

65

Platform on a cushion of oil 2) Open-end valve

Oil Tank

1) Tank valve

Bolt

Figure 57. Scheme of connections

• By using the potentiometer, adjust the output signal of the sensor to 0.6 V in order to achieve static calibration. This value should be constant during 20 minutes before starting calibration.

• Load and unload pressures (weights) in the rotating platform on a cushion of oil, by screwing clockwise and unscrewing anticlockwise the bolt (ensure the open- end valve is open and the tank valve is closed). Loads should follow always 0 bar intercalated series, like the one shown on table 12. Ensure 0 bar pressure by opening the valve 1. While following the process, take the times for each load and take also a slow PDAQ log is with the PDAQslowlog application.

66 Table 12. Applied pressure and time when it was applied Time Pressure (bar) 0 0 1:23 25 1:39 0 2:16 50 2:35 0 3:42 60 4:05 0 4:52 85 5:10 0 5:52 105 6:07 0

• Open the log file with MS Excel and look for stable points for each pressure.

Figure 58. Cylinder pressure sensor output voltage vs Time graph on the load and unload stages.

• Use Pkalz program to make an analysis of the log and obtain the calibration coefficients.

• Get the constant variation with temperature, and extrapolate the value of the calibration constant at the correct working temperature (for the engine operation, about 150ºC inside the combustion chamber), by using the slope between the manufacturer values.

67

Figure 59. Sensor data on its label.

Figure 60. Sensor constant vs Temperature, for manufacturer and calibration values.

68 Appendix 5: Calibration of the Gems 2200AGA4001A2UA pressure transducer

The description of Gems R114684 pressure sensor can be found on section 3.2.

The calibration procedure is based on static load, as described: • Join the pressure sensor to the PV411A hand pump with DPI 705 manometer attached to it, ensuring a good sealing of the joint with the pertinent o-ring. • Change the channel of pressure measurement on Cell4, so the mV signal is showed. This can be done by changing the “Zero Reduction” to 0, and the “Scale Factor” to 1. • Take a time log while loading the sensor under known pressures, increasing them from 0 bar to almost the higher limit the pressure sensor can handle (4 bar for this pressure sensor model). Operation of the hand pump can be understood by watching figure 53.

Vacuum- Pressure valve

DPI 705 PV411A manometer hand pump

Pump actuator

Figure 61. PV411A hand pump and DPI 705 manometer

69 • Open the log file with MS Excel and look for stable points for each load.

Figure 62. Intake pressure signal vs Time graph on the load and load stages, with its correspondent trendlines.

• Build a Pressure vs Pressure sensor signal table or graph, and calculate the trendline parameters, slope and intercept.

Figure 63. Pressure vs Intake pressure sensor signal graph on the load stage with its correspondent trendline.

70

SUMMARY OUTPUT

Regression Statistics Multiple R 1.0000 R Square 1.0000 Adjusted R Square 1.0000 Standard Error 0.0037 Observations 6

Coefficient Standar Lower Upper Lower Upper s d Error t Stat P-value 95% 95% 95.0% 95.0% Intercept -0.0021 0.0031 -0.6916 0.5272 -0.0107 0.0064 -0.0107 0.0064 X Variable 1 0.3984 0.0006 642.22 3.5E-11 0.3966 0.4001 0.3966 0.4001 Figure 64. Results from Regression analysis on Pressure vs Intake pressure sensor signal by MS Excel Regression Tool.

• Introduce the results in the Pressure channel of Cell4. Remember the correct way to introduce them (expression 33).

Pr essure = Scale _ Factor·()Signal − Zero _ Re duction = ⎛ Intercept ⎞ Expression 33 = Slope·⎜ Signal − ⎟ ⎝ Slope ⎠

71 Appendix 6: Calibration of the torque sensor

The torque sensor is based on a commercial Burster 8523-50 S/N 307215 force sensor attached to an arm. Its descriptions can be found on section 3.7 and on (see section 3.2.4 in Ref[2]).

The calibration procedure is based on static loads and unloads, as described: • Change the channel of torque measurement on Cell4, so the mV signal is showed. This can be done by changing the “Zero Reduction” to 0, and the “Scale Factor” to 1. • Load known weights on the arm, with a known distance from the center of the shaft holding the oil pump to the loads (it should be the bigger possible, in order to get an accurate measurement even with light loads). Take a time log while increasing the weights from 0 grams to the higher limit, and then also in the unload process. *Suggestion: a basket or any other homemade item where to place the loads is a solution that will make the calibration easier and faster. Of course the weight of the basket should be added to all the loads, even with 0 grams. • Open the log file with MS Excel and look for stable points for each weight.

Figure 65. Force sensor signal vs Time graph on the load and load stages.

• Calculate the moment due to each load by simply using expression 34.

Moment = Mass·g·Length Expression 34

72 where g is the gravitational acceleration. • Build a Moment vs Force Sensor Signal table or graph, and calculate the trendline parameters, slope and intercept.

Figure 66. Moment vs Force Sensor signal graph on the load and load stages, with its correspondent trendlines.

SUMMARY OUTPUT

Regression Statistics Multiple R 0.9999 R Square 0.9998 Adjusted R Square 0.9998 Standard Error 0.0087 Observations 15

Standar Lower Upper Lower Upper Coefficients d Error t Stat P-value 95% 95% 95.0% 95.0% Intercept -1.6115 0.0098 -165.05 5.6E-23 -1.6325 -1.5904 -1.6325 -1.5904 X Variable 1 0.0020 0.0000 286.20 4.4E-26 0.0019 0.0020 0.0019 0.0020 Figure 67. Results from Regression analysis on Troque vs Force sensor signal by MS Excel Regression Tool.

73

• Introduce the results in the Torque channel of Cell4. Remember the correct way to introduce them, as in expression 35.

Torque = Scale _ Factor·()Signal − Zero _ Re duction = ⎛ Intercept ⎞ Expression 35 = Slope·⎜ Signal − ⎟ ⎝ Slope ⎠

74 Appendix 7: Calibration of the air flow meter

The air flow meter is a differential pressure based on, type Setra 265 (complete code 26512RWD2BT1C), installed on an exhaust gases catalyst the intake circuit.

Calibration of the sensor is made by using a BIOS DC-Lite flow calibrator: • Assembly the BIOS DC-Lite flow calibrator in series with the intake circuit. According to the placement, the closer to the catalyst the more accuracy in its measurements, while the closer to the inlet port of the engine the less influence of leakages along stabilizing tanks, air heater and other elements (be aware that any change in the circuit after

calibrations will modify air leakage and so the measurement).

Figure 68. BIOS DC-Lite flow calibrator

• Change the channel of air flow measurement on Cell4, so the mV signal is showed. This can be done by changing the “Zero Reduction” to 0, and the “Scale Factor” to 1. • Start operating the valve in order to open the circuit and deliver flows to the BIOS DC-Lite flow calibrator. Deliver a series of flows like shown on the table. Be aware of the output ranges of both calibrator and flow meter: lower limit of the output signal of the Setra 265 differential pressure flow meter is so high compared to lower band of the range of flows of BIOS DC-Lite calibrator that doesn’t have enough sensibility to detect any pressure drop (see figure 61 compared to table 14 to realize it); while DC-Lite calibrator has the smallest range of flows and so establishes the higher flow of the calibration procedure (20 l/min). BIOS DC-Lite has the possibility of automatically repeat 10 measurements and take their average (consult manual). An adequate procedure would be to take 2 rounds of 20 measurements for each flow: during the first one the valve is operated and the flow changes, while flow is stabilized at the beginning of the second round and measurement would be reliable.

75 Table 14. Average flow over ten measurements and time when average was taken Time (min:sec) Flow (l/min) 0:00 0 1:33 2.093 2:10 6.499 22:40 10.02 3:08 12.56 3:32 14.39 4:00 17.3

• Open the log file with MS Excel and look for stable points for each output signal step. Notice that Setra 265 differential pressure sensor has not sensibility working under 6 l/min when working as flow meter. Therefore, that working points should be avoided or skipped (in the table above, all the flows under 6 L/min should be taken as 0 l/min, since sensor will deliver voltage equivalent to this flow).

Figure 69. Setra 265 differential pressure sensor signal vs Time graph.

76

• Build a Flow vs Differential pressure sensor signal table or graph, and calculate the trendline parameters, slope and intercept.

Figure 70. Flow vs Setra 265 differential-pressure-meter signal, with its correspondent trendlines.

SUMMARY OUTPUT

Regression Statistics Multiple R 0.9995 R Square 0.9990 Adjusted R Square 0.9987 Standard Error 0.1471 Observations 5

Standar Lower Upper Lower Upper Coefficients d Error t Stat P-value 95% 95% 95.0% 95.0% Intercept 8.0484 0.0984 81.79 0.0 7.7352 8.3615 7.7352 8.3615 X Variable 1 0.0767 0.0014 56.10 0.0 0.0724 0.0811 0.0724 0.0811 Figure 71. Results from Regression analysis on Moment vs Force sensor signal by MS Excel Regression Tool.

77 • Introduce the results in the Air flow channel of Cell4. Remember the correct way to introduce them, as in expression 36.

Air flow = Scale _ Factor·(Signal − Zero _ Re duction) = ⎛ Intercept ⎞ Expression 36 = Slope·⎜ Signal − ⎟ ⎝ Slope ⎠

• For other expressions, like Mass flow or Delivery Ratio, create the adequate channels on Cell4 and follow expressions 37 and 38. ⎛ g ⎞ ⎛ l ⎞ ⎛ g ⎞ Mass flow ⎜ ⎟ = Volumetric flow ⎜ ⎟·ρ ⎜ ⎟ = ⎝ min ⎠ ⎝ min ⎠ ⎝ l ⎠ 101325 Pa Intake Absolute Pr essureChannel Cell 4 ()bar · ⎛ l ⎞ 1 bar 1 m3 Volumetric flowChannel Cell 4 ⎜ ⎟· · = ⎝ min ⎠ ⎛ J ⎞ 1 kg 1000 l R ⎜ ⎟· ·Room Absolute TemperatureChannel Cell 4 ()K ⎝ kg·K ⎠ 1000 g

⎛ l ⎞ Intake Absolute Pr essureChannel Cell 4 ()bar Volumetric flowChannel Cell 4 ⎜ ⎟· ·353.05 ⎝ min ⎠ Room Absolute TemperatureChannel Cell 4 ()K Expression 37

mass of delivered fresh air(kg) DR = = ⎛ kg ⎞ displacement volume()m3 * ambient density⎜ ⎟ ⎝ m3 ⎠ ⎛ g ⎞ Mass flowChannel Cell 4 ⎜ ⎟ ⎝ min ⎠ 3 · 3 m ⎛ rev ⎞ displacement volume()cm · ·Angular SpeedChannel Cell 4 ⎜ ⎟ 106 cm3 ⎝ min ⎠

1 = Atmospheric Pr essure ()Pa · ⎛ J ⎞ 1 kg R ⎜ ⎟· ·Room Absolute TemperatureChannel Cell 4 ()K ⎝ kg·K ⎠ 1000 g

⎛ g ⎞ Mass flow ⎜ ⎟ Channel Cell 4 min ⎝ ⎠ ·0.0578 ⎛ rev ⎞ 1 Angular SpeedChannel Cell 4 ⎜ ⎟· ⎝ min ⎠ Room Absolute TemperatureChannel Cell 4 ()K Expression 38

78

Appendix 8: 12 V power electric map

79