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May 2011

Design And Development of a Variable Geometry Intake for a 4-Stroke 4 Cylinder Engine Wasim Sarwar - Supervisor: Dr. Iain Dupère When optimizing a race engine for a racing formula in which the regulations mandate a restricted airflow, the air intake and cam profiles/timing play a vital role. This paper will look into maximizing the performance of restricted (airflow) engines though development of the air intake system.

School of Mechanical, Aerospace and Civil Engineering, The University of Manchester, Manchester, UK, M13 9PL


Table of Contents Nomenclature ...................................................................................................................... 3 Abstract .................................................................................................................................. 4 1.0 Introduction .................................................................................................................. 4 2.0 Design Objectives ........................................................................................................ 5 2.1 Method to achieve objectives ........................................................................................... 5 2.2 Constraints .............................................................................................................................. 5

3.0 Intake Theory ............................................................................................................... 5 3.1 Effect of restrictor ................................................................................................................ 7

4.0 Simulation Validation ................................................................................................ 8 4.1 Ricardo WAVE ........................................................................................................................ 9 4.1.1 Definition of Engine Parameters........................................................................................... 9 4.1.2 Discretization............................................................................................................................. 12 4.1.3 Intake Geometry Modeling................................................................................................... 13 4.1.4 Exhaust Geometry Modeling ............................................................................................... 19 4.2 Ricardo Vectis ......................................................................................................................22 4.2.1 Purpose of Coupled Simulation .......................................................................................... 23 4.2.2 Process of Coupled Simulation ........................................................................................... 23 4.2.3 Vectis Geometry Preparation .............................................................................................. 24 4.2.4 Mesh Generation ...................................................................................................................... 25 4.2.5 Mesh Density.............................................................................................................................. 26 4.2.6 Vectis Configuration................................................................................................................ 30 4.2.6.3 Physical Parameters & Co-Simulation Linking......................................................... 31 4.2.6.4 Solving Panel .......................................................................................................................... 32 4.2.6.5 Boundary Conditions .......................................................................................................... 32 4.3 Simulation Limitations .....................................................................................................33 4.3.1 1D Simulation (Ricardo WAVE) ......................................................................................... 33 4.3.2 1D-3D Coupled Simulation (Ricardo WAVE with Ricardo Vectis) ....................... 33

5.0 Static Intake ................................................................................................................ 33 5.1 Typical FSAE Intake Solutions .......................................................................................33 5.1.1 Conical Spline............................................................................................................................. 34 5.1.2 Side Entry .................................................................................................................................... 34 5.1.3 Top Center Feed........................................................................................................................ 35 5.2 Component Optimization Method ................................................................................36 5.2.1 Throttle ........................................................................................................................................ 36 5.2.2 Restrictor ..................................................................................................................................... 36 5.2.3 Plenum.......................................................................................................................................... 36 5.2.4 Bellmouth .................................................................................................................................... 36 5.2.5 Runners ........................................................................................................................................ 36 5.2.6 Injectors ....................................................................................................................................... 37 5.3 Component Optimization ................................................................................................37 5.3.1 Throttle ........................................................................................................................................ 37 5.3.2 Restrictor & Nozzle ................................................................................................................. 39 5.3.3 Plenum.......................................................................................................................................... 44 5.3.4 Bellmouth .................................................................................................................................... 48 5.3.5 Intake Runners .......................................................................................................................... 51 5.4 Optimal Conditions ............................................................................................................59 5.4.1 Nozzle............................................................................................................................................ 59 5.4.2 Plenum.......................................................................................................................................... 59

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5.4.3 Bellmouths .................................................................................................................................. 59 5.4.4 Runners ........................................................................................................................................ 60

6.0 Variable Geometry Intake ..................................................................................... 60 6.1 Transient ...............................................................................................................................60 6.2 Proposed System ................................................................................................................60 6.2.1 Drawbacks of Proposed System ......................................................................................... 61 6.3 Feasibility of Use .................................................................................................................61 6.3.1 Cost ................................................................................................................................................ 62 6.3.2 Reliability .................................................................................................................................... 62 6.3.3 Production of Control Systems ........................................................................................... 62 6.3.4 Suitability of a Variable Geometry Intake to a FS Car ............................................... 62

7.0 Conclusion ................................................................................................................... 62 Appendix ............................................................................................................................. 63 Woshni Heat Transfer Model.................................................................................................63 Management Scheme................................................................................................................64

Works Cited ....................................................................................................................... 68

Nomenclature AAD – Average Absolute Deviation AFR- Air to Fuel Ratio

 - Coefficient of discharge

CAN – Controller Area Network CCD – Cylinder-to-Cylinder Distribution CFD – Computational Fluid Dynamics CFL – Courant-Friedrichs-Levy condition ECU – Electronic Control Unit FS – Formula Student IMEP – Indicated Mean Effective Pressure IVO – Inlet Valve Opens PCM – Power Control Module SLS – Selective Laser Sintering WOT – Wide Open Throttle WRC – Wave Ram Charging

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Abstract When optimizing an engine for a racing formula in which the regulations mandate a restricted airflow, the air intake and cam profiles/timing play a vital role. This report will look into maximizing the performance of restricted (airflow) engines though development of the air intake system. The racing formula considered was Formula Student and the associated constraints were applied to all investigations undertaken. Investigations were carried out to determine the optimal conditions for a variable geometry intake. Subsequently the potential performance increase, and feasibility of such a system could be considered. Engine simulation tools Ricardo WAVE and Ricardo Vectis were used to investigate the influence that each constituent of the intake system had upon engine performance. Further, the resources of the University of Manchester Formula Student team were assessed to determine the feasibility of manufacture and implementation.

1.0 Introduction Formula Student (FS) is an international motorsport competition in which universities are challenged to design and build an open wheeled racecar. Students compete in a variety of events where the demonstration of their knowledge is just as important as the performance of their vehicles. The competition is run by the Institute of Mechanical Engineers, is backed by many automotive corporations and boasts the highest profile motorsport engineers amongst its patrons. FS is designed as a role-playing activity, where the students assume the role of engineers who have been tasked to build a prototype track day car for a manufacturing firm. The significance of this is that the students themselves are responsible for the project and thus are empowered to think independently and given creative freedom. As with any form of motorsport, regulations exist. However in FS, they are intentionally kept to a minimum to promote innovation whilst maintaining safety. Given the resources and time restrictions placed upon the University of Manchester’s FS team, boosting engine power by increasing the pressure of the flow of air into the engine appeared to be the most effective avenue to pursue. The use of turbo-chargers and superchargers was avoided as recent changes in the regulations have ensured they are difficult to implement effectively; thus a variable geometry intake was considered. The aim of this report is to determine the feasibility and effectiveness of a variable geometry intake. This will be achieved by using engine simulation tools, Ricardo WAVE and Ricardo Vectis, to investigate the effects of varying individual intake geometry parameters. The optimal intake geometry for each given engine speed will be found, and a system to vary between these geometries suggested. This system will be analyzed to determine the feasibility of implementation.

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2.0 Design Objectives To produce a reasonably flat torque curve for increased racecar drivability, whilst maintaining acceptable transient response and useable bhp/torque.

2.1 Method to achieve objectives In order to maximize the performance we must first understand the basic operating principles of an intake system and further the influence of the restrictor on the airflow. In order to validate ideas and to optimize engine parameters, a means by which to measure any effects that changes had on the engine was required. A combination of a 1-D engine simulation code, (Ricardo Wave), 3-D CFD co-simulated with the aforementioned 1-D engine simulation code (Ricardo Vectis), and physical testing on a dynamometer were utilized. The validity of these methods of measurement for various parameters will be discussed later. Further to intake assembly optimization, a variable geometry system will be investigated and developed.

2.2 Constraints In order to ensure that the intake produced upon completion of this report is practically applicable, it is being designed to comply with the technical regulations for Formula Student. Formula student is a racing formula designed to challenge university student to design and build a single-seat racecar. The technical regulations are tailored towards light, well-balanced, relatively low powered cars to ensure longitudinal velocity is kept to a minimum and hence a greater level of safety maintained. The key excerpts pertaining to the intake from the technical regulations (Society of Automotive Engineers, 2010) can be condensed into the following statement: All air entering the engine must pass through a 20mm restrictor, (regulations B8.1.3, B8.6.1), and the intake system must not make contact with the ground in the event of a rollover, (regulation B.8.4.1). Further, the system must be designed to be compatible with the existing infrastructure of the University of Manchester Formula Student Team to test the implantation potential. The infrastructure which requires compatibility includes the: -

Engine Control Unit (ECU) - forms the basis of the vehicles electronic systems, Controller Area Network (CAN) bus - is the preferred data stream format, Power Control Module (PCM) – provides power to the vehicles systems, Engine – powers the vehicle.

3.0 Intake Theory The purpose of an intake manifold is to evenly distribute the combustion mixture between engine cylinders, to improve the performance and efficiency. In order to maximize the performance of the engine through modification of the intake we must combust a greater mass of fuel in a given period of time. In order to combust a greater mass of fuel, more oxygen is required in the combustion

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chamber. To get a greater amount of air [and hence oxygen] into the combustion chamber we must increase the momentum of the air travelling into the combustion chamber. During the valve overlap period, this increase in momentum will aid in the removal of combustion products from the combustion chamber, and upon conclusion of the intake stroke, a higher air density. Before we can comment on methods by which the momentum of the air will be increased, we must understand the pressure fluctuations occurring within the intake. Every time the Inlet Valve Opens (IVO), the reduction in cylinder pressure creates a negative pressure-wave pulse. This pressure pulse propagates up the intake runners until it reaches the boundary between the intake runners and plenum, rarefaction occurs. [Rarefaction – the density of the air at the entrance of an intake runner suddenly decreases causing the formation of a wake. The surrounding air rushes into the runner to fill this wake causing the formation of a positive pressure-wave pulse (propagating down the intake runners towards the valves). Thus the pressure has been reflected upon contact with a large change in cross sectional area. With each reflection, the amplitude decays.] The reflected positive pressure pulse can be tuned to increase the momentum of the air at a given engine speed. This phenomenon is known as wave ram charging (WRC), and can increase volumetric efficiency to a peak of 125%, (Harrison & Dunkley, 2004).

Figure 1: Rarefaction – (Left-Right, (a)-(e)) – (a) Negative pressure-wave pulse propagating through intake runner towards plenum, (b) Negative pressure-wave pulse reaches intake-plenum boundary, (c) Higher density air in plenum fills the wake caused by the negative pressure-wave pulse, (d) Smaller positive pressure-wave pulse is formed, (e) Positive pressure wave pulse travels towards intake valve

We can calculate the period of time required for the reflection of a pressure pulse, as we know a pressure pulse will propagate at the speed of sound:  

2       

1.

However the time period itself is not a very useful; knowledge of the position in the engine cycle that the time period corresponds to is required; given the valve profiles, we require knowledge of the crankshaft displacement,  , to find the position in the cycle:       

2.

6


  Where:

2 360

 ! 60

3.

 = Crankshaft angular displacement (degrees) (deg [Angular displacement for one reflection]  = Runner Length (m)

 = Engine crankshaft speed (revolution/minute)

Intake runner length for maximum wave ram effect (mm)

Studies have shown (Hiesler, 1996), 1996) that the optimal  is between 80 and 90. 90 Taking   85, figure 2 shows the optimal runner length for a given engine speed. 1400 1200 1000 800 600 400 200 0 0

2000

4000

6000

8000

10000

12000

14000

Engine Speed (RPM)

Figure 2:: Relationship between runner length for maximum wave ram effect and engine speed

Clearly, the ideal runner length (that which will produce maximum volumetric efficiency) will vary as engine speed varies. An engine air intake that could utilize WRC across the engines practical rev range, range, rather than at a single engine speed could substantially increase engine efficiency (as demonstrated by figure 2) with a smaller increase in complexity, mass and volume than other devices which aim to boost performance by increasing the amount of air delivered to the cylinders, (namely turbo or superchargers). supercharger

3.1 Effect of restrictor The air restrictor will influence the performance of the engine throughout the rev range by increasing pumping losses. Figure 3 below demonstrates that the restrictor is not chocked (experiencing sonic flow) at at any point during the engine cycle.

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Figure 3: Mach number through restrictor plotted against engine crankangle at 12,000rpm

The significance of the pumping losses is illustrated in figure 4. They occur as a result of overcoming the pressure difference between the air intake and atmosphere. At its peak the pressure difference is 13%, which subsequently leads to a performance drop of approximately 15% at 12,000rpm, (though this figure will vary based on the effectiveness of the restricted intake system).

Figure 4: Pressure plotted against engine crankangle at 12,000rpm – Pink line - pressure at the restrictor, blue line - pressure at location of pink line with the restrictor removed

4.0 Simulation Validation In order for us to have confidence in the simulations so they can be used as a design tool, it is important for us to validate the results.

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4.1 Ricardo WAVE Ricardo WAVE is a 1D engine & gas dynamics simulation software package developed by Ricardo Software to analyze the time dependent fluid dynamics dyn and thermodynamics of pressure waves, mass flows and energy losses in ducts, plenums and the manifolds of various systems and machines, machines (Ricardo, 2009). 2009) WAVE is employed throughout the engine design process, from the inception stages to detailed investigations of existing production engines; engines; and therefore contains a broad and comprehensive range of features to reflect this. this Ricardo WAVE is to bee used to model an engine with complex intake and exhaust manifolds. Multiple intake ake assembly configurations will be considered to determine the required conditions to maximize the volumetric efficiency over a range of engine speeds.

Figure 5: Graphical display of 1D engine model

Figure 5 shows a graphical representation of the engine modeled in Ricardo’s 1D engine simulation package, WAVE. The validity of the model is determined almost entirely upon the quality of the input data; justification will be provided for the data input. 4.1.1 Definition efinition of Engine Parameters To produce data from our 1D engine simulation that can be validated, we must ensure all input data matches that of the exact physical engine the data will be validated against. 4.1.1.1 Engine Configuration The engine modeled is a 2004 600cc Yamaha R6 engine. The Yamaha R6 was used because it complies with the Formula Student technical regulations, and the engine is available for use in physical testing (on a dynamometer).. All parameters specified in figure 6 were taken from the Yamaha R6 service manual, (Yamaha, 2004).

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Figure 6: Engine configuration parameters

4.1.1.2 Engine - Friction Mean Effective Pressure The data given under the friction correlation panel is used in the Chen-Flynn friction correlation model to calculate the Friction Mean Effective Pressure (FMEP), which is calculated in wave by the following equation: %&'(

 % ) *% +(,-. / ) % 0

12 12 5 3 ) 4% 0

3 2 2

4.

Figure 7: Engine friction correlation

The ACF term is the constant portion of the term and thus can be used to enter the friction value directly, (by setting the other values to zero). The BCF term varies linearly throughout the rev range to reflect the changes in the maximum cylinder pressure, (,-. , which can be related to the frictional losses. The CCF term varies linearly with piston speed to account for hydrodynamic friction which occurs during the power stroke. The QCF term varies quadratically with piston speed to account for windage losses during the power stroke.

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4.1.1.3 Combustion Profile

Figure 8: Engine combustion profile

Ricardo WAVE does not use a predictive combustion model; instead it models the heat release of combustion against time as displayed by the graph in figure 8, (Ricardo, 2006). The SI Wiebe model is used to model combustion against time and it is a simple S-curve function, which represents experimentally observed cumulative heat release “well for most situations�, (Ricardo, 2006). The first derivative with respect to time of the S-curve function will give the rate of heat release. The values associated with the calculation of the S-curve function were taken from a similar, inline 4-cylinder engine of displacement under 1000cc. For the purpose of performance optimization, approximate values are adequate. 4.1.1.4 Engine Heat Transfer

Figure 9: Engine Heat Transfer parameters

The Woshni model of heat-transfer assumes the flow has a constant heat flow coefficient and velocity over all surfaces of the engine cylinder, (Ricardo, 2009). The standard model has been modified to account for varying levels of Indicated 11


Mean Effective Pressure. Details of the Woshni heat transfer model can be found in the appendix. 4.1.2 Discretization Discretization is the process of taking a large volume and splitting it into smaller sub-volumes in order to obtain better a better resolution of the changes occurring within a fluid. In each of these smaller sub-volumes, equations will be solved to ensure mass and energy are conserved. Momentum will be conserved, and vector quantities such as mass-flow and velocity will be stored at the boundaries of these subvolumes; scalar quantities such as temperature and pressure will be stored at the center of the sub-volumes, (Ricardo, 2009). Larger discretization will lead to a greater resolution and hence greater accuracy; however it is accompanied by a corresponding slowing of the simulation. Figure 10 demonstrates the greater level of accuracy provided by a higher level of discretization using the example of a travelling pressure wave.

Figure 10: Pressure wave travelling in a pipe shown at various levels of discretization – Source (Ricardo, 2009).

Clearly smaller sub-volumes will provide more accurate results. However WAVE is an explicit solver with a variable time-step. The Courant-Friedrichs-Levy condition (CFL) is used to determine the maximum allowable time-step for every sub-volume and can be written as: 1  DA;-A-A>@E; ;F>> @G ;@EA H |DA;-A-A>@E; B-; J>K@<:L| 678 . 9:;<=>:?-:@A 8>ABC

5.

Where CFL is a user-imposed multiplier between 0 and 1 to reduce time-step size and stability. As the time-step is linearly related to discretization, higher discretization will lead to smaller elements, which will result in a smaller time-step, and therefore more time-steps will be required in order to complete the simulation. Further, the equations for the conservation of mass and energy are completed in each sub-volume, thus more sub-volumes means more calculations and hence computational time required, (Ricardo, 2009). As acoustic modeling may be carried out, the discretization should be high enough to resolve the highest frequency of interest in a particular study. A

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discretization length, (â&#x2C6;&#x2020;

), ), which is 10% of the wavelength of the highest frequency is found to produce good results, (Ricardo, 2009), (Jebasinski & Eberspacher, 2009). â&#x2C6;&#x2020; 

  1  10 %,-.

6.

Where %,-. is highest frequency of interest. The simulation should resolve several harmonics above the engine firing frequency at the highest engine speed, therefore %,-. can be given as: %,-. 

N-.:,E, >A:BA> ;F>> +RN/ . SE,T>= @G C-=,@A:<; -T@J> G:=:AB G=>UEA<L

7.

VW

For high temperature flows, â&#x2C6;&#x2020; should be modified based on the following relationship: O

â&#x2C6;&#x2020; â&#x2C6;&#x2020;

P1  

XY 5

C@

O

â&#x2C6;&#x2020;

P1  

XY 5

8. <@K

4.1.3 Intake Geometry Modeling

Figure 11:: Graphical representation of the air intake sub-assembly sub

The modeled air intake assembly is almost identical to that employed on the 2009 University of Manchester Formula Student Car, which was subsequently utilized on the racing teams dynamometer. This particular intake was modeled, as it was simplistic in design, design, and allowed the validity of the Ricardo WAVE results to be compared with those obtained from the dynamometer.

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4.1.3.1 Throttle

Figure 12: Throttle parameter definition

Although the air intake assembly used on the dynamometer utilized a single barrel throttle, a simpler butterfly valve was deemed to be an adequate approximation where transient response simulations were not carried out. 4.1.3.2 Nozzle

Figure 13: Converging (Left) & Diverging (Right) sections of nozzle

The two sections combine to form form a throat of 20mm, as mandated by regulation B8.6.1 of the FS regulations, regulations (Society of Automotive Engineers, 2010). 2010) This throat will lead to significant changes to the flow downstream of the throat, and thus a sufficiently small discretization must be utilized to ensure an acceptable resolution,, and hence accuracy is maintained throughout the nozzle, (to prevent large error propagation). ). 4.1.3.2.1 Nozzle Discretization Upstream of the throat Ricardo recommend a discretization of 0.45 multiplied by the engine bore diameter: â&#x2C6;&#x2020;  0.45 * \  0.45 65.5   29.475 

9.

Recalling equation 7, Ricardo recommends taking the 5th harmonic: 14,000 + (&/ 5 7000  `a 30 3 Therefore, re, substituting into equation 9: 9 %,-. 

10.

14


343  14.7  7000 10

3 We therefore used a discretization length of 14.7mm in the nozzle. ∆ 

11.

4.1.3.2.2 Nozzle Taper Above a Reynolds number of 2000, flow separation occurs at a divergence angle of 7°, (Sparrow, Abraham, & Minkowycz, 2009), 2009) (Kolin, Markov, Sukhanov, Trifonova, & Shukhoztsov, 2008). 2008) 1D calculationss cannot predict the pressure loss due to flow separation, as the flow will will remain attached at all points; WAVE however will use a tabulated value of pressure drop. drop 4.1.3.3 Plenum The plenum has been modeled as cylindrical in shape using complex Y-Junctions Y and mass-less less ducts. Figure 12 shows the modeled geometry of complex YY Junction 6, which is connected to the remainder of the plenum using ducts 127 and 128, and connected to the engine through duct 32 (intake runner).

Figure 14:: Graphical representation of plenum, where the blue arrow highlights Complex Y-Junction Y 6 (Left) – Physical geometry definition of Complex Y-Junction Y 6 (Right)

Complex Y-Junction 6 is indicated in figure 14.. The heat transfer/skin friction area is defined as the internal surface area of the shape, and the initial conditions are reasonable estimates. 4.1.3.3.1 Complex Y-Junction Junction Definition To explain the geometric definitions used to define the physical physical geometry of complex Y-Junction Junction 6 in figure 14, 14 the arbitrary system em in figure 15 is considered.

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Figure 15: Modeled Air-Filter – WAVE representation (Left), 3D representation (Right) – Source (Ricardo, 2006)

The DELX value is also known as the characteristic length, and is the length between the duct connection point and the wall across the volume as seen in figure 16 (a). The DIAB value is also known as the expansion diameter, and is the maximum area that the gas can expand into perpendicular to the duct entrance, as seen in figure 16 (b).

Figure 16: 3D Representation of (a) DELX (Left) and (b) DIAB (Right), (Ricardo, 2006)

The “Thick” value was set to zero as it has no influence upon performance simulations.

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Figure 17: Complex Y-Junction 6 definition

Figure 16 defines the remaining physical geometry; wall friction and heat transfer coefficients are set to a recommended value of 1, where a value of zero will ill turn off wall friction and heat transfer entirely. 4.1.3.4 Runners

Figure 18: 1 WAVE representation of Intake Runner Network

In WAVE, the intake runners were modeled as a series of orifices connected by ducts. Multiple ducts were used rather than a single one to increase the flexibility allowed by WAVE to model the runners as seen in figure 19.

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Figure 19:: WAVE definition of Duct 39 (highlighted above in figure 18)

Each intake runner is curved slightly downstream of the plenum-runner plenum runner interface, and as such a method is required to accurately define the geometry in WAVE. Although WAVE cannot model flow separation (as it is a 1D simulation), a tabulated pressure loss coefficient coefficient angle based on angle and length is applied, (Ricardo, 2009). The duct-orifice duct network applied allows us to apply the bend angle/pressure /pressure drop to duct 39, whilst still modeling ducts 38, 40, and 41 as straight. 4.1.3.5 Injectors Complex Y-Junctions Junctions were created to model injector mounts, to which injectors were added as seen in figure 20. 20

Figure 20: WAVE Injector mount definition

The physical geometry was defined as before, and once again recommended values weree used for wall friction and heat transfer multipliers.

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Figure 21: 2 Injector (highlighted in figure 19) definition

A proportional characteristic is utilized in which adequate fuel is injected into the fluid stream to match the targeted air-fuel air ratio, (Ricardo, 2006). 2006) Although the exact behavior of the injectors can be mimicked through the use of actuators, to reduce complexity it was concluded that in this 1D model such action was not required as the result conformed to expectations. 4.1.4 Exhaust Geometry Modeling The exhaust system comprises ducting, collectors collectors and a resonator (silencer), which is graphically displayed in figure 22. 2

Figure 22: 22 WAVE representation of exhaust system

4.1.4.1 Exhaust Duct Network The exhaust ducting is defined in WAVE WAV in a similar manner to the runner network employed in the intake system as shown in figure 22 above. Experimentally obtained gas and wall temperatures were used (1100K and 750K

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respectively) to define the initial conditions in conjunction with a recommended pressure of 1.05 bar. 4.1.4.2 Collectors Correctly modeling the collectors was essential in order to obtain the correct power curve from the simulation. The collectors are responsible for flow effects such as scavenging and will therefore heavily influence the volumetric efficiency and in turn the profile of the torque curve produced.

Figure 23: WAVE collector definition

As before, geometric measurements could be taken of the existing system, and wall and gas temperatures measured. Once again the recommended pressure value was utilized. 4.1.4.3 Resonator (Silencer) In order to conform to the Formula Student regulations, the maximum sound level must remain below 110 dBA, (Society of Automotive Engineers, 2010). We therefore require a device to reduce the noise levels output; we have used a resonator to do this, (which works on the principle of superposition to create destructive interference). The resonator comprises of a perforated tube inside a larger (concentric) tube. The resonator acts as a muffler and dampens a range of frequencies in the exhaust stream. This is modeled in WAVE as a series of complex Y-Junctions connected by massless ducts, (Ricardo, 2009).

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Figure 24: Resonator representation in WAVE â&#x20AC;&#x201C; Source (Ricardo, 2009)

Beginning with the perforated tube, we must first decide how many complex YY Junctions are required to model the tube. To determine the number of consecutive Y-Junctions Junctions required, required we use equation 12:  c  d e f ! 1 



   1  '  1 1!a

12.

300   7.5 40 

Therefore 7 complex Y-Junctions Junctions per tube will suffice and the Y-Junctions Y Junctions are defined as per figure 25.. The volume is obtained from the knowledge that each YY Junction has a length of 300mm/7, 300mm/7, and the Heat Transfer/Skin Friction area is the total area less the area of perforations. perforations Note we use a resonator thickness of 1(mm) and discharge coefficient of 0.6 to define the perforation.

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Figure 25: Perforated tube complex Y-Junction Y definition (left), Geometry Definition (Right)

Within the perforated tube, the massless ducts used to connect the complex YY Junctions are similar to those used in the plenum and are simply defined as straight ducts with no taper with a length of zero. The massless ducts that connect the perforated tube to the outer can, however, employ a â&#x20AC;&#x153;countâ&#x20AC;? value greater than the default 1; the â&#x20AC;&#x153;countâ&#x20AC;? value is defined as the number of fluid flows from the left hand side of the duct to the right.

Figure 26:: Massless duct within perforated tube (Left), (L between perforated tube and surrounding can (Right)

The count value is obtain as:   g  



 c   1   c  c   d e f ! 1

13.

200  28.57 7

4.2 Ricardo Vectis Ricardo Vectis is a 3D fluid dynamics program designed specifically to address fluid flow simulations for engines and vehicles. Vectis is capable of modeling inin cylinder fluid motion, spray dynamics, intake and exhaust system fluid behavior and vehicle thermal management. Ricardo Vectis is to be used to model the air intake where WAVE produces insufficient data. Ricardo Vectis will be coupled with WAVE to produce detailed three-dimensional dimensional flow data; data we will focus upon the dynamic flow field data

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produced by this coupled simulation to investigate the pressure waves and pulses occurring within the intake. intake 4.2.1 Purpose of Coupled Simulation The purpose of 1D-3D 3D coupled simulation between Ricardo WAVE and Vectis is to predict engine performance whilst obtaining detailed three-dimensional three dimensional flow information about a given component. As WAVE and Vectis will communicate, coupled simulation has the added advantage of producing results of greater accuracy. This is achieved by using the three-dimensional, dimensional, detailed flow information produced by Vectis to continually feedback into the one-dimensional dimensional ducts used by WAVE; therefore as the simulation progresses, the boundary conditions will recurrently update. Figure 27 demonstrates how the results produced by WAVE and Vectis differ.

Figure 27: Graphical representation of results produced in WAVE (left) and Vectis (Right) â&#x20AC;&#x201C; Source (Ricardo, 2009)

WAVE will predict perfect mixing and hence hence assumes that the junction will produce a homogenous mixture, however Vectis will predict species transportation and mixing and we can therefore see how the spatial mass fraction varies across the mixing zone. 4.2.2 Process of Coupled Simulation In order for WAVE and Vectis to communicate, we must set-up set boundaries, i.e. locations where the 1D and 3D codes can communicate. In WAVE, boundaries are set-up up simply by inserting external junctions that are later matched to the associated Vectis boundary. We will will discuss how this is accomplished in Vectis in subsequent sections.

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Figure 28: Modified graphical representation of engine in WAVE to include external junctions

Coupled analysis occurs in three stages: 1. The 1D WAVE code is the only code that is run, and it is run until converged. This solution will be used as the initial flow field conditions for the 3D Vectis code. 2. One-way way coupling, (WAVE will pass data to Vectis, but Vectis will not pass information back ck to WAVE), will commence for 180 Crank Angle Degrees (CAD). This will allow the Vectis flow field to initialize prior to two-way two coupling commencing. 3. Two-way coupling will commence, and WAVE E and Vectis will pass data between each other. 4.2.3 Vectis Geometry ometry Preparation The intake from the 2009 Manchester Formula Student car was the intake modeled; it is the intake currently mounted upon upon the university dynamometer, thus some validation n of Vectis results is possible. possible To be modeled, the intake was modified modified to simplify the problem. Although Vectis allows sensor modeling, all sensors and sensor mounts were removed as sensor positioning is beyond the scope of this investigation. Further, the fuel injectors were removed as for the majority of cases; they induce induce unnecessary additional processing requirements.

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Figure 29: Simplified CAD representation of 2009 Intake (left) – STL export of geometry (right)

The Computer Aided Design (CAD) intake is exported as an STL file and imported into the Vectis geometry preparation tool. The triangles of the STL files are then defined to reflect their purpose as seen in figure 30.

Figure 30: Boundary definitions

Fluid will only flow into and out of an “Inlet/Outlet” boundary, and therefore these boundaries are placed to mate with the external junctions placed in WAVE. 4.2.4 Mesh Generation A three-dimensional mesh is created with local refinement blocks where a greater mesh density is required. The local refinement blocks have been placed in areas of particular interest (figure 31 – right), and areas of significant flow effect activity (figure 31 – left).

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Figure 31: Vectis mesh definition with local refinement around restrictor and nozzle (left) plenum runner interface (right)

The level of refinement is determined by the “DEEP” and “FORCE” values. The DEEP value determines the level of refinement of the cells at the surface whilst the FORCE value determines the level of refinement applied to the internal cells. The refinement given by DEEP and FORCE is defined as the cell split into 29hhR or 27i6h and it is recommended that the values of DEEP and FORCE should remain in a 2:1 ratio, thus the DEEP value was set to 4, and the FORCE value is set to 2. \''( j 2k  16

%l ' j 25  4

The local refinement is 16 times finer than the global mesh at the surface, and 4 times finer than the global mesh internally. The mesh structure is generated specifying fluid properties, boundary conditions and the calculation parameters required for each cell to begin CFD calculation. 4.2.5 Mesh Density A variation of mesh density is likely to have the greatest effect in the vicinity of the restrictor, where activity is the highest, therefore a cross-section is taken through the center of the restrictor and examined.

26


Figure 32: Cell Size â&#x20AC;&#x201C; Large mesh (Left), small mesh (Right)

The two meshes shown in figure 32 are of differing cell size and for simplicity will be referred to as the large and small meshes respectively. Mesh Size Maximum Cell Size Large 3.2237 mm Small 1.854 mm

Variance Minimum Cell Size 73.89% 0.20785 mm 0.12159 mm

Variance 70.94% -

Table 1: Cell Size variation

An uneven mesh is generated with greater resolution in the vicinity of the walls. This refinement will ensure a smaller cell size is used in the region where a boundary layer is likely to form, thus providing greater accuracy without resorting to computationally expensive, high-resolution uniform meshes. 4.2.5.1 Effect upon Mach Number & Temperature

Figure 33: Mach number at 570 Crank Angle Degrees â&#x20AC;&#x201C; large mesh, (Left), small mesh (Right)

The flow velocity is considered at 570 CAD where activity within the system is high. A short period prior to this point spitback was seen from the restrictor, and the flow is now just beginning to move in the positive direction, (through the restrictor towards the plenum). The variation in the maximum and minimum velocities between the meshes is very small thus the large mesh is sufficient to provide a reasonable account of the general behavior occurring at any given period during the engine cycle. Figure 33 however displays slight differences in the flow behavior. At point A, we clearly see flow separation in the small mesh, but not in the large mesh. Similarly, at point B, the small mesh shows greater spitback form the runners than the 27


large mesh, however significantly, the large mesh does still show this behavior. The behavior at point B contributes to the different radius of curvature of the circulating flow at point C, where the small mesh produces a smaller radius of curvature. At point D we see the non-circulating flow from the intake runners, and again, the behavior differs. In the large mesh this non-circulating flow is larger than that seen in the small mesh, as the circulating flow at point C is smaller in the large mesh. Point E displays a lower velocity in the small mesh, however this is less pronounced in the large mesh. This difference can likely be attributed to the damping effect that a large mesh (low mesh density) has. Cell Size Large Small

Maximum Mach Number 0.045971 0.043538

Variance 5.58% -

Table 2: Mach number variation

A similar variation in behavior can be observed when viewing the temperature contours.

Figure 34: Temperature at 710 Crank Angle Degrees â&#x20AC;&#x201C; large mesh, (Left), small mesh (Right)

4.2.5.2 Effect upon Turbulence

Figure 35: Turbulent Length at 15 Crank Angle Degrees â&#x20AC;&#x201C; large mesh, (Left), small mesh (Right)

Cell Size Large Small

Maximum Turbulent Length 537.90 m 543.25 m

Variance 0.985% -

Table 3: Turbulent Length variation

The turbulent length is a measure of the length of turbulent eddies and is shown above in figure 35. The smaller mesh produces a maximum value slightly higher

28


than that produced in the large mesh. Figure 35 shows that the smaller mesh produces a more defined turbulent length.

Figure 36: Turbulent Time Scale at 10 Crank Angle Degrees â&#x20AC;&#x201C; large mesh, (Left), small mesh (Right)

Cell Size Large Small

Maximum Turbulent Time 25.50 ms 25.35 ms

Variance 0.592% -

Table 4: Turbulent time variation

The turbulent time scale is a measure of the turbulent period. The variation of behavior of the turbulent time, plotted in figure 36 mirrors that of the turbulent length, plotted in figure 35. The effects seen in figures 35 and 36, are similar to those seen in section 4.5.1 above. Once again, although the variation in maximum value is small, the behavior observed shows a marked difference. 4.2.5.3 Mesh Selection Quite clearly the small mesh is producing superior results to those produced by the large mesh. As we are simply using our WAVE-Vectis coupled models to determine relationships, and therefore build up a generalized picture of the behavior, the small mesh should produce adequate results. Ideally, a mesh with a greater density than the small mesh would have been considered; the results of this smaller mesh would have been used to check if the solution produced by the small mesh is fully converged. A lack of time and computational power however, meant that this was unfeasible. A solution for the large mesh took approximately 500 hours to produce, and approximately 700 hours were required to produce a solution for the small mesh. Insufficient time remained to complete a coupled simulation using a mesh with greater density.

29


4.2.6 Vectis Configuration 4.2.6.1 Timing, algorithm and iteration parameters

Figure 37: Vectis timing, algorithm and convergence settings

The PISO algorithm is used rather than the SIMPLE algorithm because of the transient nature of the simulation; the SIMPLE algorithm would have been used for a steady state problem. A post processing interval of 5 Crank Angle Degrees (CAD) was selected as it is sufficiently small to produce good quality animations, whilst being large enough to maintain a reasonable post processing file size. To ensure the solution is stable and accurate, the time-step length must be smaller than the time-step length in WAVE: 1  720p 11  !n!  1    1 

o5W okV

 0.97

14.

Typically a time-step of 0.25 CAD is used; here 0.12 CAD is used for greater accuracy. An end time of 7200 CAD equates to 10 full engine cycles, and is specified as the end condition. 4.2.6.2 Simulation Controls As coupled simulations can require in excess of 300 hours for completion, simulation restart parameters were setup such that if the simulation were interrupted it could be restarted. Further Vectis was instructed to write postprocessing files for all parameters of interest such as density, pressure, flow velocity, turbulent kinetic energy and shear stress.

30


Figure 38: Restart controls (left) â&#x20AC;&#x201C; Monitoring Point Panel (right)

Monitoring points are used by Vectis to monitor the convergence of the solution. Monitoring points have been setup in all significant areas of the intake, with one in each intake runner, one in the plenum, one in the diffusor of the nozzle and one in the restrictor. 4.2.6.3 Physical Parameters & Co-Simulation Linking

Figure 39: Models definition screen (left) â&#x20AC;&#x201C; Species data (right)

The species data option (seen in figure 39) is used to define the properties of the four species present in our simulations; fresh air, burnt air, fresh fuel and burnt fuel. The k-q turbulence model is used, as it is the recommended model when coupling WAVE with Vectis. Whilst using it, we must understand its limitations 31


and weaknesses. The k-q turbulence module can have trouble predicting boundary layer transition and separation. Further, in a highly compressible flow, as seen at high rpmâ&#x20AC;&#x2122;s, dilation effects can be seen on turbulence production. The works of Yianneskis et al show the variation in produced results between the different available turbulence models to be negligible, (Chen, Lee, Yianneskis, & Ganti, 1995). Here the external junctions that were setup previously (4.2.2) link WAVE to Vectis. The boundary numbers seen in figure 30 are simply matched to the number of the external junctions in WAVE. 4.2.6.4 Solving Panel The solving panel is used to inform Vectis which equations require solving. Since we wish to examine the flow behavior in the intake we require all equation except those associated with species two (burnt air) and species three (burnt fuel). Species two and three are not present in the air intake and therefore Vectis will not attempt to calculate their associated equations.

Figure 40: Solving Panel

4.2.6.5 Boundary Conditions Although WAVE will provide the boundary conditions in a coupled simulation, boundary conditions should still be defined as though Vectis was running independently. Zero Dimensional data is setup as in figure 35. In a coupled simulation the use of E the Zero dimensional data is confined to Er and the length to predict the turbulent boundary conditions, (WAVE does not predict turbulence and cannot therefore communicate turbulent boundary conditions).

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Figure 41: Boundary conditions â&#x20AC;&#x201C; Zero Dimensional Data

4.3 Simulation Limitations 4.3.1 1D Simulation (Ricardo WAVE) Ricardo WAVE cannot model 3D flows and therefore cannot model turbulence. WAVE attempts to compensate for a lack of flow field and turbulence modeling by using tabulated values for pressure drops. These values are used to account for flow separation at large angles and pressure drops due to bends in the geometry. The 3D flow field cannot be examined; to model the 3D flow field, WAVE must be coupled with a CFD package. WAVE overestimates the heat transfer from the walls to the air, modifications have been made to the standard Woschni model to compensate for varying levels of IMEP, however a coupled simulation should be run when temperature must be modeled accurately. As atmospheric velocity is defined as zero, ram-air effects cannot be directly modeled. 4.3.2 1D-3D Coupled Simulation (Ricardo WAVE with Ricardo Vectis) Coupled simulation is extremely computationally expensive. Simulations using an acceptable mesh density took on average 18 days (~430 hours) to achieve converged solutions for 5 complete engine cycles. Should one wish to undertake extensive simulation, large scale parallel computing or supercomputer use would be required.

5.0 Static Intake Before a variable geometry is considered, we consider a regular, static intake to optimize the static components and to ascertain the optimal geometry across the rev range.

5.1 Typical FSAE Intake Solutions Typically in FSAE, 3 types of intake are used. They are known as the Conical Spline, Side Entry and Top Center Feed.

33


5.1.1 Conical Spline The conical spline intake aims to treat the plenum like a continuation of the diverging section of the nozzle, and as such has the ability to provide a very even distribution off the flow between the cylinders.

Figure 42:: Conical Spline Intake â&#x20AC;&#x201C; UAS Ravensburg 09 (Left), Model for Simulation â&#x20AC;&#x201C; Source: (Stockburger, Claywell, & Horkeimer, 2006) (Right)

Although the conical spline intake will provide the most even flow distribution to the cylinders, (Stockburger, Claywell, & Horkeimer, 2006), 2006), manufacturing the runners at an equal length will prove very challenging. Further, Further, the conical spline shape does nott provide a good platform to build a variable geometry intake due to the challenges presented by the taper of the plenum and the curve of the runners; the conical spline intake therefore is nott an ideal platform for a variable vari geometry intake. 5.1.2 Side Entry The side entry intake is also commonly known as a non-symmetric non symmetric intake manifold as the flow enters the plenum perpendicular to the intake runners, and said plenum is non-symmetric symmetric to encourage even flow distribution, (Tun & Ling, 2006).

34


Figure 43: Non-Symmetric Symmetric Side Entry Intake – TU Dresden 09 (Left), Model for Simulation – Source: (Stockburger, Claywell, & Horkeimer, 2006) (Right)

The side entry ntry plenum helps maintain a lower center of gravity, is easier to package and can very simply be made to utilize ram induction (utilizing the forward velocity of the vehicle to increase the momentum of the air travelling into the cylinder). Variation of the plenum geometry is particularly challenging due to its non-symmetric symmetric nature and the sensitivity of flow distribution between the cylinders to the plenum taper. As the mountings of the plenum must be fixed, and significant runner protrusion into the plenum plenum is undesirable in this case, variation of the runner lengths is also very challenging, thus, the side entry plenum is not suitable for a variable geometry intake. 5.1.3 Top Center Feed The top center feed intake feeds air to the center of a symmetrical cal plenum, which mimics the cylinder arrangement for the runner outlets.

Figure 44: Top Center Feed Intake – University of Bath 09 (Left), Model for Simulation – Source: (Stockburger, Claywell, & Horkeimer, 2006) (Right)

Although the top center feed intake provides a greater volumetric imbalance between the cylinders than the conical spline intake, (though a smaller imbalance than the side--entry intake, (Stockburger, Claywell, & Horkeimer,

35


2006)), it is far simpler to manufacture. Packaging problems can occur if an emphasis is placed upon low-end torque, (as longer intake runners will be required). The plenum can be easily modified to accept both a variation in plenum and runner geometryâ&#x20AC;&#x2122;s, therefore the top center feed intake would provide a good platform upon which to build a variable geometry intake. The type of intake used will heavily influence the cylinder-to-cylinder distribution of the air, and is a major consideration when attempting to package the intake into the given constraints.

5.2 Component Optimization Method The components of the intake will be investigated independently. Subsequently, where a comprise must be reached between different parameters, a further investigation will be completed. 5.2.1 Throttle The two potential primary weaknesses of throttle bodies on restricted intakes are pressure drops under partial load, and obstruction to flow at Wide Open Throttle (WOT). A system alleviating the aforementioned potential weaknesses, whilst keeping cost at a minimum, will be developed. 5.2.2 Restrictor The angle of divergence from the nozzle is the primary influencing factor in volumetric efficiency maximization, (Claywell & Horkheimer, 2006). Although a smaller angle of divergence will prove beneficial when attempting to improve volumetric efficiency, it is accompanied by a corresponding increase in length of the diverging section of the nozzle, presenting challenges when attempting to package the intake system within the regulations. The aforementioned regulation B8.4.1 states no part of the intake must come in contact with the ground in the case of a car rollover, (Society of Automotive Engineers, 2010). Where the need for compromise occurs, effects will be investigated. 5.2.3 Plenum An ideal comprise between achieving greater engine power (through a large plenum) and acceptable transient response (through a small plenum) is essential in order to maximize engine performance. An investigation will be run in order to gauge the sensitivity of transient response and power output to establish an acceptable compromise. 5.2.4 Bellmouth A pipe entry will be designed using the findings from (Blair & Cahoon, 2006) to maximize the coefficient of discharge ( ), and hence maximize the mass flow rate through the pipe. 5.2.5 Runners As the purpose of the runners is to transfer air between the plenum and the intake ports of the engine, the ideal geometry would contain little to no radius, and have a smooth internal surface. An investigation will be run to determine the optimal runner length (to maximize performance) over a range of engine speeds using WAVE.

36


5.2.6 Injectors The injectors must maximize the atomization of the fuel in order to ensure the levels of incomplete combustion are kept to a minimum, and hence efficiency and power output is maximized. Mixing is primarily influenced by the speed of the flow, as sufficient turbulence is produced in most systems to aid mixing. When the flow is relatively slow, (at low rpms), better mixing and hence performance is achieved by placing the injectors further away from the intake valve. When the flow is fast moving, (at high engine speeds), higher performance/efficiency is produced with the injectors placed close to the inlet valve. As adjusting the diameter of the intake runners is varying the speed of the flow, any variation of the injectors would not be to the primary influencing factor, thus it was decided that the significant computational expensive and additional complexity could not be justified.

5.3 Component Optimization 5.3.1 Throttle Data logged at the 2009 Formula Student competition formed the basis of the requirements of the throttle. Analysis of the data obtained whilst the car was competing in the endurance event and thus running around the Silverstone track showed that 60% of the lap was completed at full throttle, and [as a result of the traction control] 96% was completed at either full throttle or the throttles idle position. A system was developed that would not obstruct the airflow at WOT; little consideration was given to pressure drops at partial throttle as these conditions accounted for under 5% of use during the endurance event. Partial throttle will be used during the sprint event, however as the throttle level is maintained at a relatively constant level, pressure drops under partial throttle will not significantly affect performance; accurate flow metering is not a high priority. A barrel style throttle system fulfills the requirements and thus was developed to meet the regulations of the Formula Student competition.

37


Figure 45: Barrel Throttle Concept

The system simply comprises a rotating barrel with a central bored hole contained within a body that contains a corresponding bored hole. When these holes align perfectly, full throttle is achieved. The idle position and full throttle positions can be seen below in figure 46.

Figure 46: Barrel Throttle Top View â&#x20AC;&#x201C; Idle Position (Left), Full Throttle (Right)

Clearly, at WOT there are no restrictions to flow. However as seen in figure 47, at partial throttle, a pressure drop will be observed as the flow must travel through large angles to pass through the throttle.

38


Figure 47: Section view of barrel throttle at partial conditions; red arrows indicate bulk motion of flow

A double barrel system was considered to aid flow metering, reduce partial throttle pressure drops and maintain an unobstructed flow at WOT.

Figure 48: Double Barrel Throttle – Partial Throttle (Left), Full Throttle (Right)

The system was dismissed as it provided little benefit whilst increasing mass, complexity and cost. 5.3.2 Restrictor & Nozzle The work of (Claywell & Horkheimer, 2006) indicates the most influential nozzle parameter upon volumetric efficiency is the angle of divergence, thus the effect of angles of divergence of 7°, 5° and 3° are investigated.

39


Figure 46: Mass Flow through restrictor for a range of nozzle divergence angles

Figure 46 shows how the mass flow varies through the restrictors tested. The data is not shown for the 7° restrictor as the data was incomplete, and insufficient time remained to re-run the simulation. The average mass flow through the 3° restrictor is 7.8858 x 10sV t 1 sX , whilst through the 5° restrictor it is 7.7544 x 10sV t 1 sX , 1.7% less. From the partial data of the 7° restrictor, it appears this behavior continues. The peaks and troughs are not aligned as the nozzle length varies therefore the time period for a pressure pulse to reach the restrictor varies. Angles greater than 7° should be avoided as flow separation occurs, reducing the mass flow through the restrictor, (Kolin, Markov, Sukhanov, Trifonova, & Shukhoztsov, 2008). It can be seen (in figures 47-49), that the pressure pulses caused by the opening of the inlet valves have an influence at the restrictor. Larger angles of nozzle divergence lead to stronger rarefaction as the nozzle is shorter, and thus the pressure pulse has less time to weaken. In fact, with the 3° nozzle, the bulk velocity at maximum spitback remains above zero, whereas it is (slightly) negative for the larger angled nozzles.

40


Figure 47: Velocity Vectors around the restrictor, with a 7째 angle of nozzle divergence at 400 CAD

Figure 48: Velocity Vectors around the restrictor, with a 5째 angle of nozzle divergence at 400 CAD

41


Figure 49: Velocity Vectors around the restrictor, with a 3° angle of nozzle divergence at 400 CAD

Prior to rarefaction occurring, the pressure pulses from m the intake valves will cause a drop in plenum pressure. This in turn will cause the formation of a flow circulation, where the air from the atmosphere is being drawn into the system by the momentum of the air, and the air from the plenum is travelling out ou of the system due to the pressure difference. Figures 50-52 show how this flow circulation develops and figure 53 shows how the pressure difference eventually overcomes the momentum of the incoming air. This will eventually lead to rarefaction as seen in figure 47. The effect’s shown in figures 50-53 53 are for a nozzle which will diverge at 5°. 5 At 3° the effect is much weaker, whilst at 7° the effect is stronger. We must note that the discussed effects occur within 140 milliseconds.

42


Figure 50: Velocity Vector at restrictor for 5° 5 angle of nozzle divergence – 240 CAD (Left), 245 CAD (Right)

Figure 51: Velocity Vector at restrictor for 5° 5 angle of nozzle divergence – 250 CAD (Left), 255 CAD (Right)

Figure 52: Velocity Vector at restrictor for 5° angle of nozzle divergence – 260 CAD (Left), 270 CAD (Right)

43


Figure 53: Velocity Vector at restrictor for 5° 5 angle of nozzle divergence â&#x20AC;&#x201C; 310 CAD (Left), 330 CAD (Right)

Although a smaller angle of divergence for the nozzle will lead to higher mass flow through the restrictor, the angle of divergence divergence will likely be dictated by the packaging constraints. This brief investigation indicated that if the angle is kept below 7°,, the variation in mass flow will be small. Further work should focus on the the effects upon CCD and mass flow when keeping the length of the nozzle equal, and varying the end diameter. This is complicated by the fact that a variation in nozzle divergence angle whilst keeping the length equal will lead to a difference of volume of the system. A detailed investigation is required. The use of swirl vanes to increase mass flow through the restrictor should be investigated. It should be noted that the shape used for the nozzle was dictated by the limited manufacturing facilities available. avail 5.3.3 Plenum Given the geometry of the plenum will differ a great deal based upon the method of variation chosen for the intake system, general relationships will be determined to form design guidance. 5.3.3.1 Plenum Volume Above 3 liters, plenum volume had little effect upon volumetric efficiency; in fact doubling the plenum volume increased the volumetric efficiency by just 1%, (Claywell & Horkheimer, 2006). 2006). Investigations were carried out between 3-8 3 liters, which forms a range that should allow reasonable flexibility in design. Transient testing, to determine throttle response, should be carried out on a dynamometer. Investigations could not be carried out out due to time limitations. Further, Helmhotlz resonators should be investigated to determine the effectiveness and result of damping the pressure pulses.

44


5.3.3.2 Plenum Shape The plenum shape determines the flow distribution between the cylinders, hereafter referred to as Cylinder-to-Cylinder Distribution (CCD). A small CCD is beneficial as a single lambda sensor is used in the exhaust. This sensor will detect the amount of oxygen within the exhaust gas, allowing the ECU to change the amount of fuel injected to meet the target Air-to-Fuel Ratio (AFR), and hence produce maximum efficiency/power. If the CCD is large, at a given period, one cylinder may require more fuel to meet the target AFR, and another cylinder require less. The single lambda sensor will not be able to make this distinction and therefore the information relayed to the ECU will not allow for optimal fuel injection. As a result, the level of fuel injected is likely to fluctuate around the target AFR and so efficiency/power will be reduced. The works of Stockburger et al use the Average Absolute Deviation (AAD) to asses the degree of CCD imbalance. The AAD is given as, S

1 \  u| : e v | 

15.

:wX

The three types of intake discussed in section 5.1 were investigated, with the results shown below in figure 54.

45


Figure 54: AAD for the main intake concepts â&#x20AC;&#x201C; Source: (Stockburger, Claywell, & Horkeimer, 2006)

Significantly, the use of a conical spline intake results in a CCD a magnitude smaller than that seen in the other two systems. The conical spline intake however must use bent runners to be mounted onto a conventional inline fourcylinder engine. As these runners must be bent to differing angles to be of equal length, the CCD comparisons should utilize the results produce when bent runners are used.

46


Figure 55: AAD for the Conical Spline Intake, with and without bent runners â&#x20AC;&#x201C; Source: (Stockburger, Claywell, & Horkeimer, 2006)

Figure 55 shows that at very high engine speeds, the bent runners produce significantly higher CCD imbalance, however as the engine will usually be operating below 11,000 rpm, the bent runners do not have a significant adverse affect. From figure 54 it can be seen that the conical spline intake has the lowest levels of AAD, and the side entry the highest. The effect of this can be seen when instructing WAVE to maintain an average AFR.

47


Figure 56: AFR distribution amongst cylinders for the three intake design concepts â&#x20AC;&#x201C; Source: (Stockburger, Claywell, & Horkeimer, 2006)

Although the conical spline intake produces significantly smaller CCD imbalance than the other two designs, it is also likely to be the most difficult to develop a variable geometry intake with. Peak variation with the top center feed intake is x3%, as opposed to the x1% seen in the conical spline intake. Although the average standard deviation is larger, it should still form a suitable base upon which to build a variable geometry intake. 5.3.4 Bellmouth Entries and exits of pipes within the intake system can produce significant head loss; the use of bellmouths is investigated to minimize this loss. The effectiveness

48


of flow transfer (through the end of a pipe) can be expressed numerically as the Coefficient of Discharge, ( ).

Typically the  is found using a flow bench, and thus for steady state conditions; this data is not particularly useful when considering the unsteady flow conditions present within the intake system. When flow travels through the pipe of area F , a “vena contracta” of area < will form and  can be found as   < ⁄F which is illustrated in figure 57 below.

Figure 57: Equipment for measurement of  , with example of “vena contracta” shown (Left), Fluent Velocity flow profile into a pipe (Right) – Source: (Blair & Cahoon, 2006)

The works of (Lau, 1995) demonstrate a significant (27%) increase in the  with the addition of a simple radius to a pipe. Lau concluded  is primarily a function of the entry diameter of the bellmouth rather than bellmouth length or profile. Further, Lau concluded optimal performance is achieved by using a short and fat bellmouth; more specifically, bellmouth length should be equal to the pipe diameter, the intake diameter of the bellmouth should be 2.13 times the diameter of the pipe, and the corner radius of the bellmouth should be 0.08 times the bellmouth intake diameter. Three systems were considered; -

Plain ended pipe – control, Simple radius - for manufacturing simplicity and space saving within the system, Elliptical profile - the system that Lau concluded would produce optimal performance.

49


Figure 58: Considered bellmouths, plain-ended pipe (Left), pipe with simple radius (Center), pipe with elliptical profile (Right)

System

Pipe Diameter Plain ended Pipe 40mm Simple Radius 40mm Elliptical Profile 40mm

Bellmouth Length

Bellmouth Intake Diameter

Bellmouth Radius

80mm 40mm

60.8mm 85.2mm

5.2mm 6.8mm

Table 5: Pn  ! 1 c 1

The performance advantage of the considered elliptical profile over the simple radius is shown below in figure 59. The performance advantage of using a simple radius over a plain-ended pipe is 27%, and the advantage of the elliptical profile is 4% over the simple radius.

Figure 59:  plotted against pressure ratio for the three considered bellmouths

To determine the reasoning behind this, both Lau and Blair and Cahoon conducted further experimentation. The mach number through a pipe was observed at a constant pressure ratio.

50


Figure 60: Mach number contours in, plain-ended pipe (Left), pipe with simple radius (Center), pipe with elliptical profile (Right) – Source: (Blair & Cahoon, 2006)

On the left of figure 60, a very strong vena contracta is evident in the plain-ended pipe. In the center, we can see that the size of the vena contract has reduced significantly and table 6 below shows the dramatic effect this has upon  (and therefore upon the mass flow rate, z). The vena contracta has reduced further with the elliptical profiled bellmouth used on the right of figure 60; once again the consequential rise in  can be observed in table 6 below. System Plain ended Pipe Simple Radius Elliptical Profile

 0.5672 0.7190 0.7430

Table 6:  for considered systems

At given crankangles in the engine cycle, reverse flow is seen and commonly referred to as “spitback”. The strength of this reverse flow seems to be inversely proportional to the bellmouths  during regular flow.

Figure 61: Maximum spitback in, plain-ended pipe (Left), pipe with simple radius (Center), pipe with elliptical profile (Right) – Source: (Blair & Cahoon, 2006)

The spitback is so strong in the case of the plain-ended pipe that a torodial vortex has formed. The strength of spitback reduces with a simple radius and is almost negligible when using an elliptical profile. The elliptical profile produces the greatest bellmouth performance, a plainended pipe the worst. The addition of a simple radius produces a significant advantage of a plain-ended pipe. Where the elliptical profile cannot be used due to packaging constraints, the use of a simple radius is acceptable as the difference in effectiveness is 4%; a plain-ended pipe should be avoided (as it is 27% less effective than a simple radius). 5.3.5 Intake Runners In order to understand the behavior of geometric changes to the intake runners, changes were made in isolation, and subsequently combined to determine the

51


cumulative effects. Sweep plots were created in WAVE to determine the optimal geometry of the runners. 5.3.5.1 Runner Length Runner length where stated is given as the value above minimum, (where minimum is the length between the engine head and the inlet valve).

Figure 62: Sweep plot showing performance variance with runner length (shown as length above minimum required)

Figure 62 agrees with the theory discussed in section 3.1 and shown in figure 2 in that to produce peak power, the runner length should be approximately inversely proportional to engine speed. Although one may expect a smooth band of colour mimicking the runner length scale at a given engine speed, such a result would be incorrect. To discuss the phenomenon that occurs in reality, we will take the example of the color band at 12,000 rpm. Peak power is delivered when the shortest runners are utilized. However somewhat curiously, we move to the longest runner length band to deliver a level of power not much less than that delivered at peak power rather than moving to the medium length band. With short runners we will see good WRC from the system. The inlet valve will open and close very quickly at 12,000rpm, thus a shorter distance to travel in the runners is advantageous because the pressure wave, which will travel at a constant speed, will re-enter the combustion chamber at the correct time.

52


Figure 63:: Pressure wave propagation in short intake runners â&#x20AC;&#x201C; ideal WRC

With medium length runners, the worst performance is achieved as the pressure wave is either still present as a negative wave wave (low pressure moving away from the inlet valve), or has just undergone rarefaction at the time when the inlet valve is closed. The pressure wave reaches the inlet valve when it is fully shut.

Figure 64:: Pressure wave propagation in medium length intake in runners â&#x20AC;&#x201C; No/Negative WRC

The performance at 12,000rpm with long runner is better than that with medium runners, but worse than that which results from short runners. WRC occurs, however the previous cycle rather than the current one generates the pressure ssure pulse. The performance is nott as good as that achieved with short runners because the pressure wave is weaker as a result of the greater frictional effects associated with longer runners.

Figure 65:: Pressure wave propagation in long intake runners runne â&#x20AC;&#x201C; Weaker WRC

To achieve peak power the runner length should be shortened with respect to engine speed. A key finding is that when looking to achieve a compromise between runner length variation and performance; simply choosing a length as close to the ideal deal length as practical does nott always produce the greatest compromise. At times, increasing the length of the runners with time can produce a beneficial compromise. Should the incorrect runner length be used,

53


performance drop by as much as 18 bhp, therefore ensuring the correct runner length is used is essential. 5.3.5.2 Runner Diameter The velocity of the flow and therefore momentum of the air entering the combustion chamber at different engine speeds can be adjusted by varying the diameter of the intake runners.

Figure 66: Sweep plot showing performance variance with diameter

A small diameter runner will be beneficial at low engine speeds, as it will increase the velocity of the flow when the mass flow rate remains the same. This higher velocity will increase the momentum of the air, thus increasing the density of the air in the combustion chamber. The mass flow rate will remain relatively constant until the point where the flow becomes chocked, that is the mass flow is limited by sonic flow occurring at some point in the runner network. In a restricted system this may occur at higher engine speeds. Figure 66 is a sweep plot generated in WAVE which shows how the power output of the engine varies with the engine speed as the diameter of the intake runners is varied. It shows the runner diameter required for peak performance varies proportionally with engine speed. At 9,000rpm, a rapid transition occurs from an optimum diameter of around 40mm to 54mm; the system tested achieves optimal WRC conditions at this engine speed, thus with a greater mass flow being delivered at this point, a larger runner diameter is beneficial. To achieve optimal performance, a diameter variation between 30mm and 40mm should suffice. Although at very low, and very high engine speeds, a

54


runner diameter outside this range would be beneficial, the magnitude of the benefit is, at most, 2.6 bhp for a very short period, thus the additional complexity of a system with 100% greater variation range cannot be justified. 5.3.5.3 Runner Taper A small angle of convergence applied to the intake runners can help accelerate the flow, thus increasing the momentum of the air entering the combustion chamber. Three taper angles were investigated, 0.5, 1, and 3 degrees.

Figure 67: Power produced by the engine across a range of engine speeds for a range of intake runner taper angles

A greater taper angle results in greater performance around the ideal WRC conditions. The intake runners tapered at 3 degrees produce an additional 5bhp (note the unit of the scale of figure 67) at 9,000rpm, the ideal WRC engine speed for this system. The effect is at its largest where WRC occurs and some effect is seen at high engine speeds. 5.3.5.4 Influential Factors Runner length, diameter and taper have varying degrees of influence upon the engine performance. The runner length and diameter are the most influential factors and their influence on performance is shown in figure 68 below.

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Figure 68: Power produced by the engine across a range of engine speeds for a range of intake runner geometry variations

For clarity, the potential power output of the engine when varying the length of the intake runners (black line) is shown against the peak power output when varying the diameter (pink line). At the lower engine speeds, variation of the diameter is more influential, whilst at higher engine speeds the runner length is more influential. Clearly, variation of both the diameter and runner length lead to a significant increase in engine performance over the control. At low engine speeds, a variation of diameter is more influential than variation of length. When a large length is used, a high Reynolds number will cause a pressure drop of up to 10 millibars, diminishing the effects of WRC, (WAVE uses tabulated values of pressure drop due to Reynolds number and pipe length); conversely when the diameter is varied, the increase of flow velocity at low engine speeds will lead to an increase in flow momentum thus increasing the density of the air within the combustion chamber. Flow with a greater velocity, as a consequence of smaller diameter intake runners will promote superior mixing between the air and the fuel, thus producing more complete combustion and therefore greater engine power. This simulation, completed with WAVE (1D flow) cannot model turbulence, thus does not account for this phenomenon.

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Case Control Variable Length & diameter

Length of Runners (mm) | Diameter of Runners (mm) at engine speed (RPM) 12,000 10,000 9000 8000 7000 6000 5000 4000 285 | 285 | 285 | 285 | 285 | 285 | 285 | 285 | 50.8 50.8 50.8 50.8 50.8 50.8 50.8 50.8 145 | 43 225 | 50 275 | 280 | 285 | 325 | 445 | 485 | 52 40 33 30 30 30

3000 285 | 50.8 535 | 30

Table 7: Runner Lengths and Diameters for optimal conditions over a range of engine speeds

To confirm the variations made to the geometry are resulting in the ideal conditions for WRC, (and thus ensuring the maximum potential performance has been achieved), we examine the charging efficiency.

Figure 69: Charging efficiency across a range of engine speeds for varying intake runner geometry

Charging efficiency (CHRGEF) is given as:   '! !n 

&11        !c 1 !c \1!  \ 1n  !c 1   

Where the displacement is equal to the engine stroke multiplied by bore area, and the density is averaged over a cycle at the intake valve of cylinder 1. The control in figure 69 clearly shows one distinct point where charging efficiency exceeds one; this is the point where WRC occurs. With the exception of the value at 3,000 rpm, CHRGEF for the engine remains above one with use of a variable diameter and length intake system, thus we can conclude that the system is successfully achieving WRC throughout the useful power band. The plot of CHRGEF against engine speed for the variable length and diameter intake runners is not simply a straight horizontal line as the intake runners are

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simply the primary influencing factor of CHRGEF rather than the sole influencing factor. In this case, exhaust scavenging is likely to be the second most influential factor; thus we see a peak in charging efficiency in this system at 9,000rpm (the engine speed at which the control is designed to produce peak power).

Figure 70: Power produced across a range of engine speeds for varying intake runner geometry, (Indicate taper angle)

As indicated in section 5.3.5.3, the use of a taper will slightly increase the performance over a small range of engine speeds, and figure 70 confirms the behavior stands when combined with the effects of varying the length and diameter.

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Figure 71: Charging efficiency across a range of engine speeds with and without intake taper

The full effect of the addition of a taper can be seen upon examination of the CHRGEF (in figure 71). Clearly a change takes place, however the effect is negligible when compared to that of varying the length or diameter. When the length and diameter of the intake runners is varied with engine speed, a significant increase in performance is observed. Tapering of the intake runners does not add a significant amount of performance. Variation of the diameter has a greater influence at low engine speeds; variation of length has a greater influence at high engine speeds.

5.4 Optimal Conditions 5.4.1 Nozzle The angle of divergence will be dictated by the packaging constraints. The use of smaller angles of divergence for the nozzle will lead to the greater mass flow through the restrictor. Given the constraints, the minimum angle of divergence should be used. 5.4.2 Plenum Above 3 liters, an increase in plenum volume does not lead to an increase in power output. An investigation was not carried out at volumes smaller than 3 liters. A conical spline style plenum produces the best CCD, however a top center feed intake should also be considered. 5.4.3 Bellmouths The elliptical profile detailed in table 5 produced the best performance, however should packaging become an issue, a simple radius would prove acceptable.

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5.4.4 Runners The optimal length and diameter of runners will change at different engine speeds as seen in table 7, (section 5.3.5.4). These lengths will be influenced by factors external to the intake such as cam timings, the valve overlap period and exhaust geometry. To determine optimal conditions for a given system, the entire system must be simulated; it is not sufficient to merely state shorter runners will benefit performance at high engine speeds and longer runners at low engine speeds as the optimal conditions at any given engine speed fall within a relatively small window.

6.0 Variable Geometry Intake 6.1 Transient In order to operate effectively, any variation that takes place within the intake system must do so at a sufficient pace. This pace corresponds to the speed at which the engine accelerates from low rpmâ&#x20AC;&#x2122;s to high, and thus corresponds to the application of the engine. For a FS car, as considered in this report, we will consider the rate of acceleration in first gear, as maximum acceleration will occur. As the rate of acceleration is non-linear, further experimentation is required.

6.2 Proposed System Variation of diameter is not easily achieved given the constraints faced. It can be achieved through the use of: -

-

An aperture system, which would control the diameter of the intake runners at the engine head and have a flexible lining inside the runners to smoothly transition from the diameter of the runners near the plenum, to the diameter of the runners at the engine head. A lining inside the runners made from a flexible polymer, which will reduce in diameter as the length of the runnersâ&#x20AC;&#x2122; increases. This system however would not be able to accommodate the bends required in the intake to fit within the required area on the FS car.

Therefore the proposed system does not include variable diameter runners. They do provide a significant increase in performance, particularly at low engine speeds, therefore if a cost effective system could be devised, their use would be advisable. The use of variable length runners however will also produce a significant increase in performance as seen in figure 68. To reduce complexity and cost [of electronic and mechanical control systems], at low engine speeds, the length which corresponds to peak performance will not be used up to 6,000rpm as the additional complexity required in the system does not justify the power increase over the control. The system was proposed and can be viewed in figure 72. It features intake runners bent at 90 degrees, which contain a single level of telescopic tubing to allow a variation in length. Figure 72 shows relatively long runners prior to the 60


area of adjustability, however this is simply for illustrative purposes and the dimensions would be adjusted to suit the situation. The curved plenum transitions into the nozzle, whose angle of divergence from the restrictor will be at the mercy of the space available.

Figure 72: Section view of proposed system mounted on Yamaha R6 Engine

The single step telescopic tubes would be controlled by two servos, such as the Hitec HS7940TH (which with a capability of turning through 60° in 0.06 seconds, should be sufficiently quick), with each servo moving two intake runners. To ensure variation of length occurs at the same rate, all four tubes would be connected by a rigid joint, (thus minimizing CCD). The servos would be attached to these joints by means of ball head bolts and corresponding adapters. 6.2.1 Drawbacks of Proposed System Many flaws can be seen with this proposed system: -

-

-

The proposed system cannot vary the diameter of the runners, nor can it achieve very long runner lengths, thus sacrificing potential power output for simplicity. When the runners are at their longest, the plenum volume is reduced. Although this isnâ&#x20AC;&#x2122;t likely to be a problem directly in terms of power output, it is likely to affect the CCD. System would require an expensive Selective Laser Sintered (SLS) plenum. Fast, powerful servos could double the price of the system.

6.3 Feasibility of Use This report will focus on the feasibility of use of the proposed system. The proposed system allows the engine to produce significantly more than when using the control, however prior to its implementation, multiple factors must be considered.

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6.3.1 Cost Over a full aluminum construction, the proposed system would add significant expense. An SLS plenum would be required, which is likely to cost in the region of £300-500. Two high speed, titanium geared servos would add a further £150200, and the associated electronic control systems required in a FS environment would add an additional £200. An additional cost of £650-900 appears to be good value for the additional power output the system brings. 6.3.2 Reliability Although the edge of the interface between the stationary section and moving section of intake runners is within the plenum, it must not become prone to airleaks. The use of a flexible outer layer to ensure air leaks do not occur must be tested extensively. Although the servos are known for being rugged, the vast majority of electronic failures on racecars are as a result of poor connections between components, therefore adding additional connections increases this risk, however the cost stated above for electronic control systems incorporates the price of motorsport standard connectors. Should a fault occur however, a mechanical failsafe must be introduced, however this is easily achievable through the use of springs. 6.3.3 Production of Control Systems The production of a sufficient quality electrical control system would be a challenge within a FS team. A system must be produced which can minimize lag, identify false data provided by external components, and be bug free. A large undertaking, however given sufficient time, it shouldn’t pose a barrier to implementation. 6.3.4 Suitability of a Variable Geometry Intake to a FS Car A variable geometry intake will add significant performance to the engine of a FS vehicle without adding a significant cost. The time of both a mechanical engineer and electrical engineer would be required to produce the system. Given the basics of the vehicle have been well developed, a variable geometry intake would be a time and cost effective feature to add to the vehicle.

7.0 Conclusion The author investigated the feasibility of the use of a variable geometry intake in a FS vehicle. Models were developed in Ricardo WAVE and Ricardo Vectis to understand how the intake manifold influenced the performance of an internal combustion engine. WAVE allowed quick, 1D simulations, whereas Vectis was used to gather detailed flow data. Works were carried out to determine the optimal geometry of the individual intake constituents over a range of engine speeds such that the required conditions for a variable geometry intake could be obtained.

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It was found that the mass flow through the restrictor was the highest when the angle of divergence of the nozzle was at its lowest. Although further work is required to determine the optimal shape and exit diameter, the works of Claywell et al demonstrate that the angle of divergence is key parameter when maximizing the performance through a nozzle. The plenum will influence power and CCD. It was found that above 3L additional volume had little effect on engine power. A concical spline shaped intake produced the greatest CCD, however given the manufacturing difficulties posed by this system, a top center feed shaped intake should also be considered.

A 27% increase in the  was seen when adding a simple radius to a pipe. The optimal bellmouth increased the  by 4%. The packaging constraints will limit the flexibility of design of the bellmouth, however, it is advised to always use a simple radius bellmouth at the very least. The length and diameter of the intake runners had the greatest influence upon engine performance. The optimal conditions were found over a range of engine speeds, and although generalities such as runner length should be inversely proportional to engine speed are largely true, the design of the system should not be carried out solely upon these generalities. The exhaust system and camshafts will heavily influence the design of the intake runners.

To achieve peak performance, the intake runners should reduce in length as the engine speed increases, however should the reduction lie outside of a relatively narrow and specific window, it could produce significantly worse performance than increasing the length of the runners. Similar behavior occurs with the variation of runner diameter at high engine speeds. A system was proposed to understand the challenges of implementation. Given the performance of the engine, the expenditure is relatively small, however for effective operation, a robust control system must be designed.

Appendix Woshni Heat Transfer Model The Woshni model of heat-transfer assumes the flow has a constant heat flow coefficient and velocity over all surfaces of the engine cylinder. This co-efficient is given as: {|w}.}~€‚}. ƒ}.€ „‚}.…† ‡ˆ}.€ ‰Š‹{Œ

Where D represent the cylinder bore, P the cylinder pressure, T the temperature, ‡ˆ the characteristic velocity and ‰Š‹{Œ the user entered multiplier. The characteristic velocity is found using the following equation: <   Ž!X , )

g9 =>G g< 5 +( e (,@ / ‘!X , ’1 ) 2 “ ” •&'(s.W5 –—˜ (=>G g=>G g 63


Where: â&#x20AC;˘ â&#x20AC;˘ â&#x20AC;˘ â&#x20AC;˘ â&#x20AC;˘ â&#x20AC;˘ â&#x20AC;˘ â&#x20AC;˘ â&#x20AC;˘ â&#x20AC;˘ â&#x20AC;˘

Â?, = mean piston speed, g9 = cylinder displacement, =>G = reference temperature, (=>G = reference pressure, g=>G = reference volume, (,@ = motored cylinder pressure, g< = Clearance volume, Â&#x2022;&'(= Indicated Mean Effective Pressure, J !X= 6.18 + 0.417 Â&#x2122; during scavenging, !X= 2.28 + 0.308 JÂ&#x161;

are closed. Â?;  Â&#x203A;  1  1Â&#x153; !n  c

!5  3.24 10

sV

RN

JÂ&#x2122;

JÂ&#x161;

when valves

Â?W

Â&#x17E;;.Â&#x;  during combustion, zero otherwise. ,

Management Scheme The management of this project proved difficult as the scope was not clearly defined. As a result, the scope covered a range of topics that could not be adequately be covered in the given time. This resulted in:

-

-

-

A reliance on the findings of similar works, where an independent investigation would have been preferential. Some key areas were not discussed, and left as further work where the reliability of findings of others was questionable. No use of the dynamometer to validate results. Although the validation was done against the manufacturers dynamometer plots, conversations with industry professionals have led me to believe the model should be tailored to a specific engine, rather than its specifications. At the time of submission, the dynamometer with the modeled engine and intake system appears to be a fortnight away from its first operation. A poor proposed variable geometry intake. Detail is lacking, and the system lacks creativity, inspiration and does not produce close to the maximum potential of a variable geometry intake.

The project was started in the summer of 2010, and a Gantt Chart was produced, (and can be seen at the end of this section). This proved very difficult to adhere to however, as in the early stages of working with Ricardo Vectis, it proved difficult to produce working simulations which generated sensible results. Given I had more work to do than I could feasibly fit into the time period, I allocated 3 days a week to my project and listed the â&#x20AC;&#x153;critical itemsâ&#x20AC;? that would have occurred in updated Gantt charts. Once these tasks had been completed, all other tasks were listed, with an approximate time of completion and priority. Having allocated time to allow for unexpected difficulties, to scope was refined.

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The project was hampered by a lack of computational power. Although 5 computers were obtained to run simulations by the start of March (with many thanks to Robert Wroe), this simply proved inadequate. As each coupled simulation took on average three to four weeks to complete, parallel computing or the use of a supercomputer would have helped greatly. Although much effort was put into gaining such computational power, the efforts were ultimately in vain. Although the project was largely completed independently, weekly meetings with Dr. Dupère proved extremely useful in assessing progress, and determining realistic targets. In hindsight, it would have been beneficial to ask Ricardo Technical assistance for help sooner than I did with the problems I encountered with Vectis. Further, a greater emphasis would have been placed on obtaining computational power as the lack of computational power significantly hindered my progress.

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Works Cited Wikipedia. (2010, November 12). Rarefaction. Retrieved November 16, 2010, from http://en.wikipedia.org/wiki/Rarefaction Yamaha. (2004). R6 Service Manual.

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Blair, G., & Cahoon, M. (2006). Design of an Intake Bellmouth. Prof Blair & Associates. Chen, A., Lee, K. C., Yianneskis, M., & Ganti, G. (1995). Velocity characteristics of steady flow through a straight generic inlet port. Internation Journal for Numerical Methods in Fluids , 21 (7), 571-590. Claywell, M., & Horkheimer, D. (2006). Improvement of Intake Restrictor Performance for a Formula SAE Race Car through 1D & Coupled 1D/3D Analysis Methods. SAE International . Dupere, I. D., & Dowling, A. (2005). The Use of Helmholtz Resonators in a Practical Combustor. Journal of Engineering for Gas Turbines and Power , 127, 268-275. Gilkes, Mishra, Rao, & Fieldhouse. (2007). TRANSIENT RESPONSE OF TURBOCHARGED DIESEL ENGINE FOR TRANSIENT OPERATION USING AIR INJECTION ASSIST SYSTEM. University of Huddersfield. Harrison, M., & Dunkley, A. (2004). THE ACOUSTICS OF RACING ENGINE INTAKE SYSTEMS (Vol. 271). Journal of Sound and Vibration, Issues 3-5. Hiesler, H. (1996). Advanced Engine Technology. Edward Arnold. Jebasinski, R., & Eberspacher, J. (2009). Calculation of the tail-pipe noise of exhaust systems using WAVE. Kolin, I., Markov, V., Sukhanov, V., Trifonova, T., & Shukhoztsov, D. V. (2008). Investigation of the development of unsteady flow separation from a model swept wing. Fluid Dynamics , 44 (5), 680-686. Lau, H. (1995). Coefficients of Discharge at the Aperatures of Engines. Queens University of Belfast. SAE. Mishra, R., & Gilkes, O. (2006). COMPARISON OF PASSIVE AND ACTIVE METHODS OF IMPROVING TRANSIENT PERFORMANCE OF TURBOCHARGED ENGINE SYSTEMS. University of Huddersfield. Society of Automotive Engineers. (2010, January 1). Formula SAE Rules. Retrieved 10 03, 2010, from SAE: http://students.sae.org/competitions/formulaseries/rules/ Sparrow, E., Abraham, J., & Minkowycz, W. (2009). Flow separation in a diverging conical duct: Effect of Reynolds number and divergence angle. International Journal of Heat and Mass Transfer , 52 (13-14), 3079-3083. Stockburger, G., Claywell, M., & Horkeimer, D. (2006). Investigation of Intake Concepts for a Formula SAE Four-Cylinder Engine Using 1D/3D (Ricardo WAVEVECTIS) Coupled Modeling Techniques. Minnesota: SAE International. Ricardo. (2009, 01 01). Using WAVE: Sub-Models: Woschni Heat Transfer. Retrieved 12 12, 2010, from CAD Family:

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http://ol.cadfamily.com/wave8/wave_help_system/help/wave/using_wave/sub -models/woschni_heat_transfer_model.htm Ricardo. (2009, 01 01). WAVE Knowledge Center. Retrieved 12 15, 2010, from CADFamily: http://ol.cadfamily.com/wave8/wave_help_system/wkc_new.htm#help/wave/u sing_wave/sub-models/iris_valve_port_geometry_model.htm Ricardo. (2006). Ricardo Wave Knowledge Center. Tun, L., & Ling, J. (2006). CFD Analysis of Non-Symmetrical Intake Manifold for Formula SAE Car. SAE Internatinal .

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Design and Devlopment of a Variable Geometry Intake for a 4 Cylinder 4 Stroke Engine