Upgrade the existing Land Rover Defenders powertrain

Introduction

The design teams project aim:

To propose a program and new engine design to upgrade the existing Land Rover Defenders powertrain

The following report discusses weaknesses of the current LR Defender along with a subsequent review of advanced engine technologies. Finally the design targets for the new proposed engine are defined with all necessary parameters calculated and justified. The design team have also considered the cost, manufacturability and marketing affects of the new engine as well as the performance requirements.

Benchmarking

The current LR Defender engine will be critically analysed in-relation to the vehicles main competitors in the market place. This will allow the weaknesses of the current engine model to be determined and thus rectified for the new specification engine.

The engine throughout the following section will be critically analysed in the following main areas;

Acceleration & speed performance

Engine power & effectiveness (MEP and power density)

Emissions of harmful gases and particulates

Fuel economy

The potential customer perception of vehicles engine and other performance criterion (including value for money).

The following competitor vehicles were researched and compared to the current LR Defender to establish baseline requirements for the new spec engine;

Jeep Wrangler Unlimited (regarded as the main competitor)

Mercedes-Benz G-Class G550

Nissan Patrol GR

Toyota Land Cruiser 3.0 D

For a table of the above vehicles full specifications, benchmarked with the existing LR Defender, refer to Fig.4.7 in Appendix.1.

Analysis & Discussion

By reviewing the benchmarking table, Fig.4.7 in Appendix.1, it becomes apparent that the current LR Defender is slower than its competitors by taking 14.7 seconds to reach 0 62 mph compared with the Jeep Wranglers time of 11.7 seconds. Although the LR Defender is not expected to have a very fast acceleration time, as it is an off-utility vehicle, potential customers may be discouraged by the Defenders slow acceleration characteristics compared to the Jeep and its other competitors.

The fastest acceleration time is achieved by the Mercedes-Benz G550 which can go 0 62 mph in an impressive 6.1 seconds with its supercharged 5L V8 petrol engine. The cost of the Mercedes-Benz G550 however is 50,250 (compared to 27,610 for the Defender) and has twice the engine capacity of the Defender, so is expected to be far more powerful with its petrol engine compared to the 2.4L turbocharged diesel Defender engine.

The design teams proposal is that the current LR Defender engine should at least match or beat its rival the Jeep Wrangler in areas such as acceleration time from 0-62 mph as the Defender is more expensive than the Jeep and approximately 230kg lighter.

Another issue is that the Defenders top speed is the lowest out of all its competitors. This low top speed has also attracted criticism from motoring journalist Jeremy Clarkson (2006) who stated that the biggest drawback of the LR Defender is its weak engine and thus corresponding slow acceleration and top speed. The new LR Defender engine must eliminate these same criticisms to become more competitive in the market place and thus more appealing to potential customers.

Vehicle

Power per cylinder (kW/cylinder)

Power Density Ratio (kW/L)

LR Defender 2.4L D 4inl

22.50

37.48

Jeep Wrangler 2.8L D 4inl

32.50

46.81

Mercedes Benz G550 5.5L V8 petrol

47.75

70.23

Nissan Patrol GR 3.0 D 4inl

29.50

39.96

Toyota Land Cruiser 3.0D 4inl

30.75

41.25

Fig.1.0 compares the effectiveness of the engines that were benchmarked in detail. It shows that the Defenders current engine has the lowest power density in terms of engine capacity. This is a big weakness as it shows that the when not taking into account the Defenders relatively small engine capacity, its engine is vastly underpowered. The Mercedes Benz G550 as expected displays the best power density as it is a gasoline engine. The Defenders regarded closest rival in terms of target market and image, the Jeep Wrangler, also shows a much greater power density ratio.

Or group propose to increase the engines the power density to be comparable with the Jeep Wrangler with at least a ratio of above 43kW/L.

Vehicle

Torque (Nm)

MEP (Mpa)

LR Defender 2.4L D 4inl

360

1.885

Jeep Wrangler 2.8L D 4inl

400

1.815

Mercedes Benz G-Class G550 5.5L V8 petrol

391

0.903

Nissan Patrol GR 3.0 D 4inl

380

1.619

Toyota Land Cruiser 3.0D 4inl

410

1.729

A weakness of the current Defender however is highlighted test drive of the vehicle by motor journalist Jeremy Clarkson (2006) for the Times Newspaper. He stated that the vehicle simply did not have enough torque to pull a horse trailer behind it at a reasonable speed. This can be backed up by the data seen in Fig.1.1 above which shows that the current Defender engine produces the least amount of torque out of all its competitors. This weakness is heavily compounded by the fact that the vehicle is marketed as an off-road vehicle and thus needs high torque requirements to be-able to navigate through steep and loose terrain effectively. As a consequence customers may be off put by the fact that the current Defender has less torque than its competitors and thus this issue will have to be addressed when upgrading the existing engine.

The new engine must be made more environmentally friendly than the current engine as January 2013 will see the current Euro 5 emissions targets replaced by the Euro 6 legislation and thus stricter emissions targets. The Defender is third best out the five competitors the vehicle was benchmarked against in regards to CO2 emissions, which is an indicator for overall total emission performance.

The current Defender comes second best to the Jeep Wrangler in regards to fuel economy, by achieving a combined fuel economy of 28.3 mpg. However the other diesel engine competitors, the Nissan Patrol and Toyota Land Cruiser are less fuel efficient than the Defender.

When purchasing an off-road vehicle such as the Defender, fuel economy may not be the most important factor but it has become more important to potential customers over the last decade. The same could be said for the greater importance of emissions performance that potential customers may look for. The new Defender engine is likely to be more powerful and thus a better fuel economy target (comparable to 32.8mpg of the Jeep Wrangler) may be out of reach. However through reviewing and selecting technologies, the fuel economy performance could be improved slightly or at least kept the same.

From analysing the Defender, by using the data displayed in Fig.4.7 (Appendix.1), it can be seen that the vehicle is neither the best nor worst performing vehicle in regards to fuel economy, but averagely in relationship to competitors.

The balance of the current Defenders inline four engine configuration, according to Nunney (2006), has perfect primary balance because when one pair of pistons are moving up, the other pair are moving down at the same time. Inline four engines however do not have perfect secondary dynamic balance. This is because piston acceleration varies depending on its vertical position within the cylinder head in relation to the crankshaft that it is connected to. This leads to one pair of pistons moving faster than the other, which creates a secondary imbalance and results in the engine vibrating vertically. Nunney (2006) also explains that at low power configurations the secondary imbalance (vibration) is not too severe but can get considerably worse with increasing size and powerful engines. This may explain why the current inline 4 Defender engine has a lower displacement than its competitors, to reduce secondary imbalance vibration in order to appeal to potential customers and save costs on designing a crankshaft to damp heavy engine vibration.

The following strengths are also exhibited by the current engine;

Highest MEP value out of the competitors benchmarked against

Best strength to weight ratio (could be due to basic interior making vehicle lighter)

Potential customers may overlook the apparent power and torque shortfalls as the highly regarded Land Rover brand may persuade people to buy the vehicle anyway.

Summary of existing Defender engine (compared to competitors);

Weaknesses

Strengths

- Slowest acceleration from 0 62 Mph

- Best Mean Effective Pressure of 1.885Mpa, better than competitor vehicles

- Lowest top speed of only 82 mph

- Highest Power to weight ratio (kW/kg)

- More expensive than Jeep Wrangler by >4000 even with poorer speed performance.

- Fuel economy is not the worst

- Lowest Power per cylinder produced (kW/cyl)

- Emission of CO2 is not the worst

- Secondary imbalance of straight inline four engine configuration (rivals also have this weakness)

- Long history of Defender may appeal to potential customers, thus engine shortcomings may be overlooked

- Lowest torque produced out of competitors

- Not the best in either fuel consumption or emissions, even with smallest capacity engine

- Lowest Power Density Ratio (kW/L)

Current Vehicle Performance Trends

Fig 1.3 shows the performance trends for Jaguar engines up to 2010. While exact performance targets cannot be extrapolated from the graph, The design group can see that the new Defender engine needs to have an increase in specific engine power. However with the increase in power, increased emissions and fuel consumption will occur. This in conjunction with Fig 1.4 below from Richardson (2010) shows how CO2 emissions have decreased despite the trend of engine power also increasing.

While the trends from Figs 1.3 and 1.4 show that performance trends increase and emissions trends decrease, the group is concluding that potential customers for a LR Defender will be less likely concerned about the vehicles emissions or fuel economy compared to non-SUV vehicles customers. Thus increasing the torque (and power) of the current engine to match its competitors is prioritised. It is conceded that any improvement in fuel economy and emissions departments will be limited, but in the interests of Land Rovers image, any improvement on these characteristics will be beneficial.

Advanced Engine Technology

Supercharging & Turbo-charging Technology

Superchargers (mechanical drive driven)

This is a device comprising of an air compressor to force more air into the engine. Forcing a greater amount of air (under positive pressure) into the engine provides more oxygen for the combustion process than without a supercharger. As a result more fuel can be thus provided for stoichiometric combustion reaction to occur and allowing more work per a cycle to be done. This thus increases the power output of the engine.

The advantage of supercharging according to Daniels (2001) is that it multiplies the engines BMEP and torque by the amount the air compressor increases the atmospheric pressure into the engine. Supercharged engines also experience better throttle response than naturally aspirated engines.

The disadvantage of using a supercharger is that it is generally less thermally efficient than the more common used turbocharger (which uses energy from otherwise wasted exhaust gas). Another drawback highlighted by Harris (2002) is that supercharging (particularly mechanical-supercharging techniques) puts extra strain on the engine and its components as they are required to withstand extra strains provided by the supercharging boost. This requires the engine to be made stronger, thus thicker, heavier and more expensive. Daniels (2001) also explains how the noise generated by a superchargers mechanical drive components can contribute to extra passenger discomfort.

Turbochargers

These devices consist of a turbine and a compressor and are a type of supercharger. The difference is that instead of mechanically driving a compressor to force more air into the engine, turbochargers uses the engines own exhaust gases (which would have been otherwise wasted). It does this by converting the kinetic energy from exhaust gases into rotational energy to turn a turbine. The turbine is connected to the compressor on the same shaft, thus this powers the compressor to draw in atmospheric air and pump it pressurised into the engine.

The advantages of a turbocharger are same as for a supercharger as previously described of increasing engine BMEP. This is appropriate for the LR Defender which will need the extra power if being used off-road or in mountainous steep roads, which is the market the vehicle is targeted to. Turbochargers are also more thermally efficient than superchargers due to use of the otherwise wasted exhaust gas. This thus decreases exhaust emissions and fumes expelled into the atmosphere. Daniels (2001) also explains that for mainly diesel engines variable geometry turbochargers can maintain an appropriate exhaust gas speed though the turbo turbine when the engine is at low load.

Disadvantages include the need for a cooler to cool exhaust gas before it enters the turbine therefore adding weight and bulk to the engine. During operation turbochargers also experience a turbo lag when the throttle is applied.

As previously explained the LR Defenders competitors (particularly its main rival the Jeep Wrangler) have more powerful engines than the current Defender 2.4litre 4 cylinder engine. If upgrading the engine by increasing its cylinder capacity, more air (particularly oxygen) will need to be supplied to the cylinders for combustion. Thus the use of twin-turbochargers may be required to force more air into the cylinders to make the combustion process stoichiometric. Also the advantage of using two smaller turbochargers (twin-turbo), instead of a larger single turbocharger, is that turbo-lag is reduced. Usually a small turbocharger provides boost at low engine speeds and the second kicks in and supplies boost at higher engine speeds. There are two widely known types of twin-turbochargers called Parallel and Sequential types.

When comparing the advantages and disadvantages of mechanically-driven superchargers and turbochargers it was decided to use turbochargers as they are more environmentally friendly and fuel efficient to run. The current engine for the LR Defender uses a variable geometry turbocharger and it is likely the new spec engine will also be turbocharged by the same type of unit.

Variable Valve Timing Systems

Camless Valve Systems

Autoweek Magazine (2005) states that camless valve systems were tested in 2005 by Valeo on two Peugeot 407s successfully under extreme weather conditions and intensive testing. The valves were controlled by individual actuators and powered through solenoids to open and close valves.

The advantages of camless systems, explained by Daniels (2001), include the following;

Valve timing can be altered to as desired

In theory some cylinders could be shut off (at low load) to allow others to run more efficiently

Valve timing and lift can be matched to the needs of the engine with an estimated saving of up to 20% on fuel saving.

The mechanical design of the engine can be simplified as the usage of a camshaft and other associated valve gear become redundant.

The advantages however are currently overshadowed by the power needs of the camless system and the associated complexity and reliability issues if the vehicle has electrical problems. Peter Brown who is vice president of powertrain engineering and design for Ricardo stated in Autoweek Magazine (2005) It comes down to complexity and cost which sums up why camless systems are still not (although many think they eventually will be) utilised in passenger vehicle engines. For The new LR Defender engine camless systems will not be used for the disadvantages described above.

Variable Valve Timing Technology

Mechadyne International (2006) states that that the use of variable valve train systems can substantially reduce both fuel consumption and exhaust emissions. The amount by which the variable valve train systems reduce fuel consumption and emissions is going to be approximated to 10%. This is because, as the Bosch Automotive Handbook (2007) states, BMWs VALVETRONIC system reduces fuel consumption and exhaust emissions by over 12%.

According to the Bosch Automotive Handbook (2007) the following types of variable valve timing technology are available;

Camshaft phase adjustment

Camshaft-lobe control

Fully variable valve timing with camshaft

Fully variable valve timing without camshaft

Camshaft Phase Adjustment

This type of variable valve timing adjusts the phase that the cams are in contact with the levers that open and close the valves. To change the phase of the camshaft small adjustments are made, by electrically controlled actuators, to the camshaft as a function of engine speed. Typically the camshaft can only be controlled to move to two pre-calculated extreme positions.

Advantages include greater power, torque and efficiency being experienced for a wider range of engine speeds. Disadvantages to other valve timing methods include the limited range in which the valves timings can actually be altered.

Fully Variable Valve Timing with Camshaft

These types of systems can vary both valve lift and timing. The lobes on the camshaft have a curved profile which in conjunction with the camshaft being able to move freely laterally, this enables the valve lift and timing to also be varied independent to each other, which is an obvious advantage to the previously limited valve control systems mentioned above.

Fully Variable Valve Timing without Camshaft

These types of systems are very different, to the previously mentioned, as it replaces the use of a camshaft with either the following types of control methods solenoid (electromagnetic) or electro-hydraulic actuators.

The biggest advantage of these systems are that operate independently from the crankshaft and thus this allows the valves to be opened at any time period of the engines cycle. This, as stated by the Bosch Automotive Handbook (2007), offers the greatest degree of freedom for valve timing and thus the greatest potential for reducing fuel consumption. Also deactivation of certain cylinders can be achieved thus allowing the active cylinders to work more efficiently at lower engine speeds. Disadvantages are however that superchargers cannot be installed (without very expensive and complicated design), and while space is saved from not using a camshaft, electrical components can be bulky and hazardous. Also the cost of fully variable valve timing systems means it is unlikely they will be incorporated into The new engine design.

Camshaft-Lobe Control

In these types of systems it becomes possible for a valves timing to be controlled by three separate camshaft lobes depending on the engine speed. According to the Bosch Automotive Handbook (2007) the one lobes profile is tailored so that valve timing and lift is optimised for the lower to mid engine speed range. Another lobes profile is optimised for higher engine speeds by maximising valve lift and opening times. Systems such as Hondas VTEC and Toyotas WTI use camshaft-lobe control method. Camshaft-lobe shifting types of variable valve timing also share similar advantages and disadvantages to the camshaft phase adjustment method.

This type of variable timing (camshaft-lobe control) will be used for the new engine design. This is because it doesnt cost as much (or weigh as much) as the other variable valve timing systems while still being hugely advantageous in terms of performance, fuel economy and emissions control gain.

Fuel Injection Systems

Common Rail Fuel Injection

These fuel systems consist of a common rail tubing system maintained at constant high pressure via a pump. Injectors for each cylinder in the engine are in turn connected to the common rail tubing. The injectors have solenoid valves which are electronically controlled via an engine ECU (Electronic Control Unit) to open and close at the desired timings as explained in detail by DENSO (2005).

An advantage of common rail fuel injection is that control of fuel injection (according to Daniels, 2001) is at the injector itself and not at the pump which is the case with other fuel injection systems. Higher pressures can also be achieved thus more fuel can be injected into the cylinder in a shorter amount of time with better fuel atomisation, as described by DENSO (2005), leading to high combustion efficiency and a reduction in emissions. This is important as new emissions targets will have to be met in 2014 with the Euro 6 legislation when the vehicle will be on the market.

The main disadvantage of this type of injection technology according to Daniels (2001) is that the injectors are expensive to manufacture and inherently complicated in design.

Piezoelectric Injectors (For Common Rail Systems)

Instead of using solenoid valves which are more frequently used in common rail fuel injection system, piezoelectric injectors can be used in higher performance engines. These injectors work by using piezoelectric crystals that expand when supplied with an electrical charge and thus opening and closing fuel injection valves. The following attributes of piezoelectric type injectors are common;

Greater compact dimensions than solenoid valve injectors.

More accurate control over injection timing and fuel volume.

Piezoelectric injectors can be used with Accelerometer Pilot Control (APC) to minimise diesel engine vibration at low engine speeds. This is achieved by injecting a small quantity of fuel before the main injection quantity.

Piezoelectric injectors can also operate faster with more frequency than solenoid valves (approximately five times faster), which allows greater control over fuel consumption and emissions.

The Bosch Automotive Handbook (2007) states that the use of piezo-injectors for common rail fuel systems can reduce emissions by up to 20%.

Emissions Reduction Technologies

Stanton (2009) from roadtransport.com explains how the European Parliament (EP) and European Commission (EC) have agreed new targets for comply with Euro-6 emission legislation. The new Euro-6 targets will have to be met by vehicle manufacturers and thus the new spec LR Defender by 1st January 2013. This is before the new LR Defender model will reach Job 1 (mid to late 2013). It is therefore important that new and existing technologies are reviewed in Emissions control to meet these targets. In recent years the environmental performance of vehicles influences potential customers more than ever in their buying decision. It is therefore important we maintain Jaguars highly regarded brand image and compete with competitors by meeting the existing (Euro-5) and future Euro-6 emissions targets.

Diesel Particulate Filters (DPF)

This is a device which is responsible for removing small particulate particles and soot from the exhaust gas of a diesel engine. A DPF is not 100% but is normally found to be over 50% efficient most of the time. A good feature of a DPF is that its function according to torquecars.com (2008) is independent to a catalytic converter thus ensuring a fault in the DPF will not affect overall emissions critically.

The advantages of particulate filters are much publicised including removing dangerous small particles from an engines emissions. The two types of DPF, active and passive, have their own advantages and disadvantages.

The main disadvantages of DPF, explained by torquecars.com (2008), are highlighted below;

The filters can get very hot causing a possible fire safety hazard.

To remove a DPF very technical changes have to be made to the affected ECUs to change the sensitivity of sensors in the vehicles engine and exhaust.

A DPF can decrease engine performance by at most 10% Bhp.

Other Technologies

Accelerometer Pilot Control (APC)

Diesel engines are known to display harsh chugging and vibration at low engine speed, which can now be minimised through technology called Accelerometer Pilot Control (APC). An APC system, described by Delphi (2008) consists of an accelerometer (microphone) attached to the engine block which listens to the nature of the combustion which may have caused vibrations occur throughout the engine block. An engine management system then minimises the unwanted vibrations and noise by optimising the amount of fuel pilot injected for combustion, in a closed loop system, until acceptable noise and vibration levels are reached.

ECU Remapping

According to mobilechiptune.com (2007), when we remap an engine ECU we are fine tuning the program that deals with engine performance. Remapping or upgrading an ECU could therefore potentially increase the available engine power and torque. Mobilechiptune.com (2007) also states that remapping a diesel turbo engine ECU will produce 30 - 50% BHP on exact the specification, where diesel engines give the most impressive power and torque gains available. A remap of the ECU will definitely be required be a twin-turbo (or other technologies) are added to the new engine, however the ECU itself is only likely to achieve small gains in efficiency, fuel economy and emissions.

Summary of Chosen Technologies

Fig1.5 below shows the selected technologies the group is proposing for inclusion into the new Defenders engine. Fig1.5 also shows estimates of the expected improvement over engine performance, emissions and fuel economy. Also see Section.3 for justification to estimates below.

Feature

Selected Technology

Twin-Turbo (reused exhaust gas

Variable Valve train

Diesel Particulate Filter

Piezo Injectors*

Improved ECU Mapping

APC

Performance, BHP

+ 20%

+ 10%

- 10%

+ 5%

+ 2.5%

+ 2.5%

Emissions, CO2 g/km

+ 5%

- 10%

~ 0%

- 10%

- 2.5%

- 2.5%

Fuel Economy, mpg

- 10%

+ 10%

~ 0%

+10%

+ 2.5%

+ 2.5%

*Piezo injectors as opposed to solenoid controlled injectors in a common rail fuel injection system.

3. Selection of Engine Arrangements

Modified engine parameters:

Total engine capacity 3000 cc.

Capacity per each cylinder 500 cc.

Number of cylinders 6

Type of engine Diesel engine

The target is to improve engine performance (mainly torque) by increasing the number of cylinders from 4 to 6. Although there is a reduction of capacity per cylinder, a net increase in total engine capacity of 600 cc will not only compensate it, but also increases total horsepower produced. Kayne (2009) states that 6 cylinder engines are more suited to towing, off-road, hilly and mountainous areas while experiencing greater throttle response. Bore size is thus reduced from 89.9 mm to 82 mm while retaining the same stroke length. Bore/stroke ratio is 1.15, which is within the range of 1-1.3 for diesel engine. The weight of the current engine is estimated as being 180kg taken from a BMW 2.5L inline 4 diesel engine (plus weight added for turbo) from data compiled by Williams (2006), which is a similar spec to the current Defenders 2.4L turbo inline 4. The new engine is estimated as being 25% larger thus heavier by the same margin, and an additional 50kg for the additional technologies added. The new engine weight is thus taken as approximately 300kg.

4. Determination of Design Targets

This section of the report provides estimations for the new engines power, torque, fuel economy and emissions characteristics. Below Fig1.7 Shows modifications to the Defenders current engine will affect the new engines power performance.

Performance Estimation

Feature

Estimated affect on engines Performance

BHP (%) affect from current Defenders 121 BHP engine

Increasing engine capacity by 600cc

+ 25 %

+ 30 BHP

Upgrading current Turbocharger to a Twin-turbo charger

+ 20 %

+ 24 BHP

Installing a Variable valve train system Camshaft Lobe Control

+ 10 %

+ 12 BHP

Decreasing the bore from 89.9mm to 82mm

- 10 %

- 12 BHP

Adding a Diesel Particulate Filter

- 10 %

- 12 BHP

Piezo-electric injectors (instead of solenoid valves) in common-rail system

+ 5%

+ 6 BHP

Miscellaneous;

-Accelerometer Pilot Control (APC)

-Improved ECU Mapping

-Improved intake air flow

+ 5 %

+ 6 BHP

Total affect in BHP =

+ 50 %

60 BHP Increase

Given the maximum power for previous engine is 121 bhp. Therefore, the new engines maximum power is:

Power = (121 + 30 + 24 + 12 - 12 - 12 + 6 + 6)bhp

= 181 bhp

= 135kW

Torque and Power at 3 operating conditions:

T = 368.5 Nm @ max power (3500rpm)

T = 400.0 Nm @ max torque (2000rpm)

T = 120.0 Nm @ idle (1000rpm)

The Torque at various engine speeds were calculated via using the following equation:

Engine power: Pe=2*? *N* T

Figure 1.8 shows the estimated power and torque curves for the vehicle.

Justification of Targets & Estimations

While the decision has been taken to increase the engines capacity, increasing the engines power to increase the vehicles acceleration and torque characteristics, the fuel economy and emissions of the engine also has to improve. This is due to more stringent legislation and targets, as well as the expectations from potential customers who expect the engine to improve in every department.

It may be said that that increasing the engines capacity from 2.4L to 3L means that the targets of decreasing the fuel consumption and emissions will be difficult. The group would argue however that the current Defenders engine is underpowered compared to its competitors and was consequently the recipient of bad reviews from motor journalists (such as Jeremy Clarkson, 2006).

The Defenders potential customer market also may not require huge improvements in fuel consumption and emissions. This is because the Defender is going to be utilised for and marketed as an off-field vehicle with specialist applications such as towing and rough terrain excursions. These categories of vehicles are expected by customers to have poorer fuel economy and emissions than other smaller vehicle types. These customer expectations will therefore be beneficial when designing the engine as while emissions and fuel economy is targeted to at least stay the same, the issue of increasing the Defenders torque can be prioritised.

The increase in engine capacity naturally means the emissions and fuel consumption will increase. To overcome this advanced engine technology will be utilised in order to decrease the emissions and fuel consumption. Estimations will be made regarding how much saving (in terms of percentage) the addition of new engine technology will have on emissions and fuel consumption. These savings from selected technologies are justified below.

Effect of increasing engine capacity

Old Defender engine = 2.4L, new proposed engine = 3L

Therefore by comparing the chemical (fuel) energy now available;

Increase in performance, fuel consumption and emissions = (3L-2.4L)/2.4L x 100=25%

Therefore it is assumed that increasing the capacity by 600cc has a 25% negative effect on fuel economy and emissions. It was also assumed that the engine performance increases by 25%.

Upgrade from turbo to twin turbo

The fuel consumption and emissions will increase when upgrading to a twin-turbocharger. However the Bosch Automotive Handbook(2007) states that the reuse of exhaust gas to work the twin-turbo can minimise the amount of emissions experienced. Therefore as more exhaust gas will be re-used by the twin turbo (as opposed to a single turbocharger) it is assumed that only a 5% increase in emissions will occur. It is also assumed that a twin-turbo will have a 10% negative effect on fuel economy as the greater fuel volume will be used to correct the air/fuel mix ratio.

Installing variable valve train (camshaft lobe control method)

According to the Bosch Automotive Handbook (2007), the BMW VALVETRONIC achieves more than 12% saving in fuel economy and emissions. Therefore we have decided to estimate The fuel economy and emissions saving as 10%. This is due to the BMWs system being more flexible than the variable camshaft lobe system the group are proposing to use. A fully variable valve system such as BMWs is not used due to its high complexity and cost, which are not suitable for the price range of the vehicle.

Piezo electric injectors

According to the Bosch Automotive Handbook (2007) up to 20% savings (likely in extreme cases) can be achieved with piezo-injectors instead of solenoid valve injectors in a common rail fuel injection system. The group has thus conservatively estimated fuel and emission savings at 10%.

Miscellaneous factors (e.g improved ECU mapping and APC) have also been factored in albeit these technologies contribute to very little performance gain, at approximately 5% in total.

Fuel Consumption Estimation

The current Defenders combined fuel consumption is 28.3mpg. This figure will be used as baseline for any changes that will be adopted for the new Defenders engine. Fig 1.9 below estimates the change in fuel consumption by using the estimations from Fig. 1.5 (Section. 2).

Feature

Estimated affect on Fuel Economy

mpg (%) affect from current Defenders 28.3mpg economy

Increasing engine capacity by 600cc

- 25 %

- 7.075 mpg

Upgrading current Turbocharger to a Twin-turbo charger

- 10 %

- 2.83 mpg

Installing a Variable valve train system

+ 10 %

+ 2.83 mpg

Decreasing the bore from 89.9mm to 82mm

+ 10 %

+ 2.83 mpg

Adding a Diesel Particulate Filter

~ 0 %

~ 0 mpg

Piezo-electric injectors (instead of solenoid valves) in common-rail system

+10%

+ 2.83 mpg

Miscellaneous;

- Accelerometer Pilot Control (APC)

- Improved ECU Mapping

+ 5 %

+ 1.415 mpg

Total affect on mpg =

0% change

0 mpg change

The current Defenders combined fuel economy is 28.3mpg. Therefore the new engines combined fuel economy is;

Fuel economy = 28.3 mpg

The table below compares the new engines fuel economy to competitors;

Vehicle

Combined Fuel Economy (mpg)

Current Land Rover Defender

28.30

Jeep Wrangler

32.80

Mercedes-Benz G-Class G550

12.00

Nissan Patrol GR

26.20

Toyota Land Cruiser

25.02

NEW DEFENDER

28.30

As shown by Fig. 2.0 above, the new proposed defender still has the second best fuel economy out of its main competitors. Only the Jeep Wrangler is more efficient. A net no change in fuel economy is very good considering the power of the engine has been increased. For the vehicles potential market of off-road vehicles, the team has decided that not getting any improvement on fuel economy is acceptable as the current Defender is already more fuel efficient than the majority of its competitors.

CO2 Emissions Estimation

The current Defenders CO2 emissions are recorded as being 266 g/km. This figure will be used as a baseline to calculate the new Defenders engine emissions. Fig. 2.1 below estimates the change in engine emissions by using the estimations from Fig. 1.5 (Section. 2).

Feature

Estimated affect on CO2 engine emissions

g/km (%) affect from current Defenders 266g/km CO2 emission

Increasing engine capacity by 600cc

+ 25 %

+ 53.2 g/km

Upgrading current Turbocharger to a Twin-turbo (exhaust gas re-use) charger

+ 5 %

+ 26.6 g/km

Installing a Variable valve train system

- 10 %

- 26.6 g/km

Decreasing the bore from 89.9mm to 82mm

- 10 %

- 26.6 g/km

Adding a Diesel Particulate Filter

~ 0 %

~ 0 g/km

Piezo-electric injectors (instead of solenoid valves) in common-rail system

- 10 %

- 26.6 g/km

Miscellaneous;

- Accelerometer Pilot Control (APC)

- Improved ECU Mapping

- 5 %

- 13.3 g/km

Total affect on CO2 emissions =

- 5 %

- 13.3 g/km Decrease

The current Defenders CO2 recorded emission is 266g/km. Therefore the new engines calculated estimated emission is;

Engine emission = 266 13.3 g/km = 252.7 g/km ~ 253 g/km

Vehicle

Engine Emissions (CO2 g/km)

Current Land Rover Defender

266

Jeep Wrangler

227

Mercedes-Benz G-Class G550

322

Nissan Patrol GR

288

Toyota Land Cruiser

214

NEW DEFENDER

253

Fig. 2.2 above shows that the new proposed defender remains third best out its competitors regarding emissions, even with a 5% reduction in CO2 emissions. The addition of a Diesel Particulate Filter (DPF) however will vastly decrease the emissions of harmful small particulates and also NOx emissions. The DPF will not however effect the emission of CO2 according to www.trolleycoalition.org (2002), which is why in estimations of CO2 output, the DPF is considered to have ~ 0% effect of CO2 g/km output. According to the Bosch Automotive Handbook (2007) the DPF can be up to 97% effective, thus it can be assumed that the emission of harmful particulates (and soot) will decrease by up to 95% in the new engine with the addition of the DPF.

5. Determination of Engine Design Parameters

Brake mean effective pressure (BMEP) = 4?T/ V

Idle 5.03 bar

Maximum Torque 16.76 bar

Maximum Power 15.41 bar

Indicated mean effective pressure (IMEP) = BMEP/ nm

We assume mechanical efficiency nm = 0.85, which yields

Idle 5.91 bar

Maximum Torque 19.71 bar

Maximum Power 18.13 bar

Friction mean effective pressure (FMEP) = IMEP - BMEP

Idle 0.89 bar

Maximum Torque 2.96 bar

Maximum Power 2.72 bar

To determine compression ratio, CR:

CR = VBDC / VTDC

= 500cc / 28.52cc

= 17.53

The other design parameters at the three power conditions are summarised in the table below.

Design Parameters

General

Idling

Max. Torque

Max. Power

New Engine Weight approx. (kg)

300

Weight to Power (max) Ratio, Mo (kg/kW)

2.3

Dimension Power Ratio, Vo

0.000023

Compression Ratio, CR

1.15

Power, Pen (kW)

12.57

83

131

Torque, T (Nm)

120

400

245.6

Engine Rotation Speed (rpm)

1000

2000

3500

Mean piston speed, Vm (m/s)

3.15

6.31

11.04

IMEP (bar)

5.91

19.71

18.13

BMEP (bar)

5.03

16.76

15.41

FMEP (bar)

0.89

2.96

2.72

GMEP (bar)

9.91

24.71

24.13

Power Density (in terms of capacity), PL (kW/L)

4

35

44

Power Density (in terms of piston area), PF (kW/cm2)

397

3282

4134

The equations used in all calculations are shown below:

Swept Volume per cylinder, Vs = (?*B2*S)/4

Piston Area: Ap = Vs/S or it can be calculated as Ap= (?*B)2/4,

Engine Capacity, Ce = Vs*Nc where Nc denotes the no. of cylinders

Engine power: Pe=2*? *N* T

Brake Mean Effective Pressure, BMEP = 2*?*T/ Vs

Mechanical efficiency, ?m =BMEP/IMEP

Mean Piston Speed, Vm = (S/30)*n

S = engine speed

n = number of strokes

Friction Mean Effective Pressure, FMEP = IMEP BMEP

Gross Mean Effective Pressure, GMEP = IMEP-PMEP

Power Density (using Capacity), PL = Pen /Vt (kW/L)

PE = Engine power

Vt = Total engine capacity (Litres)

Power Density (using Piston Area), PF = Pen /(iAp) (kW/cm2)

PE = Engine power

i = Number of Cylinders

Assumptions made during calculation:

T = 368.5 Nm @ max power (3500rpm)

T = 400.0 Nm @ max torque (2000rpm)

T = 120.0 Nm @ idle (1000rpm)

To determine clearance height, hc:

Given compression ratio for previous engine is 17.5, capacity per cylinder is 600cc and bore diameter, B = 89.90mm.

CR = VBDC/ VTDC

17.5 = 600/ VTDC

VTDC = 600/ 17.5

= 34.285 cc

VTDC = (?*(B)2/ 4)*hc

34.285 = ?*(8.99)2hc

hc = 0.54 cm = 5.4mm

Below the P-V and P-??are plotted for the 3 operating conditions of the engine;

6. Engine Operating Dynamics

Load Flow Diagram

Forces that act upon piston and conrod can be split into normalized pressure load, fg and normalized inertial force, fj. The combination of these two forces yields the real resultant force. The formulae for the two forces are:

f_j=(m_j r?^2)/((?B^2)/4) (cos?+?cos2?)

And

f_g=(P/V)^1.4=22.5/[A_pr(1-cos?+0.075cos2?)]^1.4

Finally, the resultant force is f=f_j ?+f?_g

The following parameters are required to calculate resultant forces.

Reciprocating mass is given by:

m_j=m_p+m_l=m_p+(m_l (l-l^'))/l=0.45+(0.6(157.7-105))/127.7=0.65 kg

Piston Area is given by:

A_p=(?B^2)/4=(??(0.082)?^2)/4=0.00528 m^2

And radius of crankshaft, r = 47.3 mm, length of connecting rod = 142mm.

The following load flow diagrams were thus produced for three engine conditions;

Speed Diagrams of Main Components

The speed diagram for piston, conrod and crankshaft are drawn using acceleration and velocity relationships with respect to crank angle, ?. The formulae are as follow:

Velocity: v=sin??+?/2 sin2?

Acceleration: a=cos?+?cos2?

? is ratio of crankshaft radius over conrod length which is assumed to be 0.333. The following speed diagram is thus acquired using Matlab:

Load on Engine Body (Head/bottom/block)

To get the maximum load possible, maximum power parameters were applied where the rpm would be 3500rpm. Below are the equations used to calculate reciprocating inertial load on cylinder block:

R_j1=F_j/2+T_kj/b=F_j/2+F_jr/b=F_j (1/2+r/b)=-m_j r?^2 (cos?+?cos2?)(1/2+r/b)

And

R_j2=F_j/2-T_kj/b=F_j/2-F_jr/b=F_j (1/2-r/b)=-m_j r?^2 (cos?+?cos2?)(1/2-r/b)

Rj1 and Rj2 represent the reaction forces at 2 bottom ends of the cylinder block.

Figure 8:

The following graph displays a comparison of loads at different operating conditions, namely idling (1000 rpm), maximum torque (2000 rpm) and maximum power (3500 rpm).

7. Mass Distribution and Engine Balance

As the team has decided to use an inline 6 Engine configuration to replace the current Defenders inline 4 which will require the crankshaft to be re-designed and balanced. An inline 6 configuration was chosen because it is the best balanced engine configuration. In an inline 6, there are three pairs of pistons which move in the same direction, being 120 degrees apart from the other pairs.

The reciprocating and rotating masses of engine components produce large inertial forces. These forces generate vibrations leading to greater frictional wear of the engines moving components. To see the extent of any imbalances, the forces generated from the rotating and reciprocating masses were broken down into x-y coordinate directions.

The inline 6 crankshaft is essentially the combination of two 3 cylinder crankshafts joined together. Thus the analysis was carried out on half the inline 6 crankshaft to make calculations simpler.

Selected Firing Sequence 1 5 3 6 2 4

Balance of Reciprocating Mass

Force analysis

The result of the force balance above indicates that there is complete balance of the engines reciprocating masses.

Moment analysis

When considering the reciprocating local moments, we get the following results;

? M?_X=M_l cos60-M_r cos60

= am_j r?^2 [3/4 cos?+?(?3)/4 sin???-3?/4 cos2?-?(3?3?)/4 sin??2?]

? M?_y=M_l cos60+M_r cos60

= am_j r?^2 [3/4 cos?-?(?3)/4 sin???+3?/4 cos2?+?(?3?)/4 sin??2?]

It is calculated that the reciprocating mass is not fully balanced, and cannot be balanced by the crankshaft or engine. Support engine mounts are thus used, to cancel out the unbalanced reciprocating masses, preventing the engine from turning over.

Balance of Rotating Masses

Force analysis

X-axis:

Y-axis:

The force analysis shows that the rotating forces are automatically balanced by the crankshaft itself.

Moment analysis

The results of analysing the rotating masses on the crankshaft are summarised below;

X-axis:

M_x=Fr2acos30+Fracos30=(3?3)/2 F_r a

Y-axis:

M_y=F_r2a+Fr2asin30-Frasin30-F_ra = ( 3)/2 F_r a

Total Moment?

Thus the moment is unbalanced.

Therefore from the calculations above it can be seen that there is a resultant rotating moment on the crankshaft due to the crank pin. To balance this rotating motion, web balancers are required. A summary of the calculations to determine the mass of the web balancers required is described below.

The Application of Web Balancers

To balance the rotating mass, the design team will use 8 web balancers as shown by the configuration in Fig. 3.3 which illustrates the locations for only half the crankshaft. This however is not the best method of balancing an inline 6 engine, which is to use 12 web balancers in total. However to save weight on the engine a compromise was chosen to use 8 web balancers instead.

The following equations and method was used to calculate the mass of the balancers;

m_r=m_c+m_2=m_c+ml'/l = 1.6kg thus,

Therefore the weight of each mass web balancer must equal 0.69kg

Engine Torque fluctuation

Torque due to reciprocating mass

If it assumed that T1 is a the torque resulting from one of the reciprocating masses, we get;

T_1=F_jrsin?(?+?)/cos?(?)

Where ? denotes the angle between the upward direction and the conrod,

?=47.3/141.9=0.3333

r (length of the crank) =47.3mm

Fj is the reciprocating force on piston

? F?_j=(m_j )?^2r(cos?+?cos2?) =13429.7(cos?+0.3cos2?)

A_p=(?B^2)/4=(??(0.082)?^2)/4=0.00528 m^2

Torque due to gas pressure

It is assumed that T2 is the torque in one cylinder due to gas pressure acting on it:

T2=PApr(sin?+0.15sin2?)

f_g=(P/V)^1.4=22.5/[A_pr(1-cos?+0.075cos2?)]^(-1.4)

Torque Fluctuation Graph

The torque due to the gas pressure is small compared with the reciprocating mass. The total torque is therefore;

The torque fluctuation in a single cylinder is thus given by the following

Flywheel Design

The equation to calculate the inertia of the flywheel is given below;

If = (900/?2) *[(? ??)/ (? n2)]

Where;

I_(f ) - Inertial moment of flywheel (kg m^2)

? Fraction of rotating mass of the total engine (assumed to be 0.8)

?? - The net work that can be stored in the flywheel and is 280 NM in the design.

? - Engine speed unevenness, assumed as 1/50 in this design.

I_(f )=(900/?^2 )*[0.8*280/(?2000?^2*1/50)]=0.255

To calculate the inertia of the flywheel we also have the following relationship:

Where;

G flywheel mass

D flywheel diameter

The flywheel for initial calculation is assumed to be a thick disc of uniform shape. The material chosen is a Titanium Alloy with a density of 4420kg/m^3.

Where;

? - Density

h - Height of flywheel and is assumed to be 25mm,

r - Radius of flywheel

The equation yields that the flywheel radius, r =165mm.

The flywheels design however will mean that it is thinner in its centre than outer diameter, as many modern flywheels are designed. The mass distribution of the flywheel was thus recalculated using the following formulae;

We get the following dimensions for the flywheel; R2=126mm, R1=195mm, r=165mm

G= ?h?r^2=4420 "x" 0.025 "x" 3.142 "x" ?(0.165)?^2=9.9kg

Therefore the mass of the flywheel is 9.9kg.

A SolidWorks representation of the flywheel is shown over the page with an engineering drawing.

8. Design Sketches

9. Valve Train Design

For new engine: 3.0L Diesel 6 cylinders in line

Camshaft arrangement type: DOHC

Number of valve each cylinder: 4

Valve type: poppet

Design Arrangement

Double over head camshaft (DOHC)

DOHC, compared to SOHC engines according to angelfire (2007), run cooler, more smoothly, quietly and efficiently, because the engine has twice as many intake and exhaust valves as a SOHC engine. The disadvantages, however, are that DOHC engines cost more for repairs. To ensure against expensive engine repairs, the engines timing belt must be changed about every 60,000 miles.

Chain drive

Chain drive is able to take high load and is quite durable and reliable (the max torque of the new engine is 400Nm).

Four valves per cylinder.

Using mushroom-shaped poppet valve, it has very good characteristics on cost, fluid flow, sealing, lubrication, and heat transfer to the cylinder head stated by Stone (1999).

Variable Valve Timing (VVT)

The new engine arrangement will have camshaft lobe control with 3 different lobe profiles available for valve timing to be optimised at different conditions.

Design Requirements

Valve overlap time

In terms of the effect of the valve overlap on engine operation, large turbocharged engines for specific operating conditions can have a valve overlap of about 150. At a full-load operating condition the boost pressure from the compressor will be greater than the back-pressure from the turbine. So the large valve overlap allows a positive flow of air though the engine. This explained by Stone (1999) ensures excellent scavenging and cooling of valves, turbine and combustion chamber with high thermal loadings.

Valve seat inserts are required (especially for Al. alloy cylinder block to ensure minimum wear).

The centre of the cam is offset from valve axis so that the poppet valves can rotate in order to even out any wear, and to maintain good seating

Valve dimensions (from Stone, 1999)

The diameter of inlet valves is larger than that of exhaust valves.

For a flat four-valve head, each inlet valve could be 39% of the bore diameter and each exhaust valve could be 35% of the bore diameter. (Gives about a 60% increase in total circumference, or 30% increase in curtain area, for same Lv/Dv)

At the inlet side the division between the two valve ports should have a well-rounded nose. For the exhaust side division wall cam taper out to a sharp edge.

Variable valve timing (VVT)

The greater valve overlap takes fuller advantage of the pulse effects that can be particularly beneficial at high engine speeds (also for turbocharged diesel engines).

Although there is some limitation for compression ignition engines:

At top dead centre: the piston to cylinder-head clearance

At bottom dead centre: the inlet valve has to close soon after BDC, otherwise the reduction in compression ratio may make cold starting too difficult

The range of valve leads and lags (from Xu, 2009)

For best engine performance

Intake Open: 10 30 lead

Intake Close: 45 75 lag

Exhaust Open: 45-75 lead

Exhaust Close: 10-30 lag

For low exhaust emissions

Intake Open: 2 10 lead

Intake Close: 30 45 lag

Exhaust Open: 30- 45 lead

Exhaust Close: 2-10 lag

Valve timing maps for the upgraded engine

According to the design requirements above, the degrees of the leads and lags are defined based on the engine speed. The faster engine operation, more air need for combustion so lead and lag of the intake valve should increase with the increase of the engine speed. In addition, the more exhaust flow, the more power generated by the engine turbo (because the new engine is twin-turbo charged).

For idling at 1000 rpm:

At the idling speed, engine does not need too much air for combustion, so the valve overlaps of TDC and BDC are defined as 10 and 60 respectively for the engine efficiency improvement, as shown in Fig. 4.3.

Intake Open 5 lead

Intake Close 30 lag

Exhaust Open 30 lead

Exhaust Close 5 lag

For the max torque at 2000 rpm:

The maximum torque is reached when the engine speed is at 2000 rpm, the overlaps increase to 30 and 100 respectively, see Fig. 4.4.

Intake Open 15 lead

Intake Close 50 lag

Exhaust Open 50 lead

Exhaust Close 15 lag

For the max power at 3500 rpm:

At this moment, the valve overlaps arrive at maximum (40 and 120 respectively) because of the max power output, as shown in Fig. 4.5, and also the emissions will be the highest.

Intake Open 20 lead;

Intake Close 60 lag

Exhaust Open 60 lead

Exhaust Close 20 lag

10. Conclusion

To conclude the design team feels that the engine upgrade described in this report improves on the performance and thus marketability of the current Land Rover Defender engine.

The fuel economy stayed the same (even though the power increase of the engine) and emissions output has been improved although not to a great extent as vehicles such as the Defender are not expected to have the best fuel economy or emissions standards by potential customers. However keeping at least the same fuel economy and emissions is good for the Land Rover company image while there is an increase in power.

The following main parameters define the engine;

The following advanced technologies were incorporated;

Twin-turbocharger

Variable valve train (Camshaft lobe control system)

Diesel Particulate Filter

Piezo-injectors for common rail fuel injection system (old engine used solenoid valves)

Improved ECU mapping (tighter control)

Acceleration Pilot Control

The following compares the performance figures for the new engine against the old engine;

Performance Area

NEW ENGINE

OLD ENGINE

Percentage Improvement (%)

Power, BHP, (kW)

181 (135)

121 (90)

50%

Torque, Nm

400

360

11.1%

Fuel Economy, mpg

28.3

28.3

0%

Engine Emissions, CO2, g/km

253

266

5%

The following points below have to be taken into consideration;

Increased weight of engine from 200kg to 300kg. However the current Defender model is its lightest among competitors and still will be with the new engine.

A larger engine means greater manufacturing costs, however the increased vehicle performance and, as Abeles (2004) explains, the fact that new car prices rise yearly means the increased manufacturing costs could be incorporated more easily.

Appendix 1 Full Benchmarking Table

Engine Parameters

Land Rover Defender

Jeep Wrangler Unlimited

Mercedes-Benz G-Class G550

Nissan Patrol GR

Toyota Land Cruiser 3.0 D D-4D C

Layout

4 cyls In Line , 2401 cc diesel

4 cyls In Line , 2777 cc, turbocharged diesel

5439cc V8 petrol, supercharged

2953cc di longitudinal inline 4 turbocharged

2982cc Longitudinal inline 4, diesel, Turbocharged

Max power

121hp/90kw@3500

174hp/130kw @3800

512hp/382kw @ 6000 RPM

158hp/118kw @3600

170hp/123kw @3400

Max torque

360nm@2000

400nm@2000

391nm @ 2800 RPM

380 nm@2000

410nm@2000

Installation

F Longitudinal

F Longitudinal

Longitudinal

F Longitudinal

longitudinal inline 4

Bore/stoke

89.9x94.6 mm

94.0x100.0 mm

97 X 92 mm

96.0x102.0 mm

96.0x103.0 mm

Compression ratio

17.5:1

17.5:1

10.7:1

17.9:1

17.9:1

Ignition and fuel

Common-rail injection. Variable geometry turbocharger. , Diesel

common rail direct injection

indirect injection

indirect injection

Valves

dohc 4 valves per cylinder

dohc 4 valves per cylinder

dohc 3 valves per cylinder

dohc 4 valves per cylinder

dohc 4 valves per cylinder

Power to Weight Ratio

22.38:1 kW/kg 30hp/kg

16.19:1 kW/kg 21.7hp/kg

18.68:1 kW/kg

0.0872 PS/kg

Other Parameters

Price

27,610

23,250

50,250

26,040

29,000

0-62 mph

14.7 sec

11.7 sec

6.1 sec

15.2 sec

12.7

Max Speed

82 mph

112 mph

131 mph

99 mph

109mph

Body Dimensions (L*W*H)

3894*1790*2021 mm

4751*2030*1840

4680*1864*1931

5030*1940*1855

4410*1800*1860

Weight

1797 kg

2030 kg

2456

2495 kg

2125 kg

Wheelbase

2360 mm

2947 mm

2850 mm

2970 mm

2460 mm

Fuel Tank

60 L

80 L

96 L

135 L

87 L

Combined/Urban

28.3 / 22.6 mpg

32.8 mpg/??

12/11 mpg

26.2 mpg/??

25.02 mpg/??

CO2 Emissions

266 g/km

227 g/km

322 g/km

288 g/km

214 g/lm

Suspensions

live beam axle, single rate coil springs, telescopic hydraulic dampers

live axle, leading arms, track bar, coil springs, gas shocks, stabiliser bar

rigid axle, coils, shocks, sports suspension

solid axle, 5 link, coil springs, stabiliser bar

rigid axle, coil springs, stabiliser bar, panhard rod

Steering Type

power assisted, worm and roller

recirculating ball, power assisted

recirculating ball power assisted

recirculating ball, power assisted

recirculating ball, power assisted

Turns Lock to Lock

3.4

3.1

3.4

3.1

Front Brakes

298mm solid discs

ventilated discs 302mm

ventilated discs

ventilated discs

ventilated discs

Rear Brakes

298mm solid discs

discs 316mm

ventilated discs

ventilated discs

ventilated drums-in-discs

Front Wheel & Tyres

235/85 R16,5.5J x 16 in

255/75 r17

285/55 r18

265/70 r16

265/70 r16 c

Rear Wheel & Tyres

235/85 R16,5.5J x 16 in

255/75 r17

285/55 r18

265/70 r16

265/70 r16 lt

Appendix 2 - Nomenclature

S Stroke (mm)

B Bore size (mm)

r Radius of crank (mm)

? Angular speed (rad/s)

? Crank angle (o)

n speed of crank (rpm)

Vs Swept volume (m3)

CE Engine Capacity (m3)

PE Engine Power (kW)

hc Clearance Height (mm)

IMEP Indicated Mean Effective Pressure (bar)

BMEP Brake Mean Effective Pressure (bar)

GMEP Gross Mean Effective Pressure (bar)

FMEP Friction Mean Effective Pressure (bar)

?m Mechanical Efficiency

Vm Mean Piston Speed (m/s)

a Acceleration (m/s2)

v Velocity (m/s)

mj Reciprocating mass (kg)

mp Piston mass (kg)

ml Conrod mass (kg)

fj Normalised Inertial Force (N)

fg Normalised Load (pressure) (N)

f Sum of gas pressure and inertial load (N)

Rj1, Rj2 Reciprocating load (reaction forces) on cylinder block on both sides (N)

bhp Brake horse Power (kW)

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Team 12 Powertrain Engineering Project Stage 1 - Engine

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