Design, validation, and performance results of a turbocharged ...

Design, validation, and performance results of a turbocharged turbogenerating biogas engine model I.G.M. Thompson, S.W. Spence and C.D. McCartan School of Mechanical and Aerospace Engineering, The Queen’s University of Belfast, Northern Ireland J.M.H. Talbot-Weiss Bowman Power Group Ltd, England

ABSTRACT Turbocompounding is the process of recovering a proportion of an engine’s exhaust energy and adding it to the output power. This was first seen in the 1930s and is carried out by coupling an exhaust gas turbine to the crankshaft of a reciprocating engine. It has been recognised that coupling the power turbine to an electrical generator instead of the crankshaft has the potential to reduce the fuel consumption further with greater flexibility. The electricity generated can be used for automotive applications to assist the crankshaft using a flywheel motor generator or to power ancillaries that would otherwise have run off the crankshaft. In the case of stationary power plants, it can assist the electrical power output. Decoupling the power turbine from the crankshaft and coupling it to a generator allows the power electronics to control the turbine speed independently in order to to optimise the specific fuel consumption for different engine operating conditions. This method of energy recapture is termed 'turbogenerating'. This paper shows the results from turbogenerating a turbocharged reciprocating compression ignition engine for stationary power generation, fuelled by an induced biogas with a small portion of palm oil being injected into the cylinder to initiate combustion. The subject engine demonstrated a fuel consumption improvement of 7.2% due to turbogenerating by comparison with the same conventional nonturbocompounded engine (the baseline engine). The power developed by the turbogenerator was 10.2% of the total system power. This paper also details the development of an experimentally validated engine model using the GT-Power 1D engine modelling software. This paper presents the outcome of a validation exercise for a turbocharged, turbogenerated biogas engine. The model has the ability to predict engine performance and turbocharger and turbogenerator operating characteristics to a satisfactory level of accuracy. NOMENCLATURE p v γ n B θ

QR t Cd BSFC

in cylinder pressure (Pa) cylinder volume (m3) ratio of specific heats polytropic exponent mass fraction burned crankshaft angle (degrees)

PID NA

11

heat released (J) Time (s) Coefficient of Discharge Brake Specific Fuel Consumption (g/kWh) Proportional Integral Derivative Naturally Aspirated

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1. INTRODUCTION 1.1 Historical Background Reference (1) details the background of turbocompounding, which found significant application at the end of World War II when it was employed in aircraft engines. The first two notable examples of this type of engine were the Wright Cyclone and the Napier Nomad (2). Turbocompounding became obsolete though in aircraft applications due to the advent of the gas turbine engine, but later found applications in sea and ground transport. The next major milestone in turbocompounding for vehicle applications was the Differential Compound Engine pioneered by Frank Wallace (3, 4). It was designed to be an integrated engine-transmission system that provided high power output with torque characteristics desirable for traction applications. Throughout the 1970s and 1980s the turbocompound engine was used in various military research programs (5, 6). During the 1980s, manufacturers such as Caterpillar, Cummins and Holset (7-9) carried out extensive research on turbocompounded road going vehicle engines, but Scania were the first company to mass produce the turbocompounded engine for road transport in the early 1990s. During the 1990s, engine manufacturers were developing technologies necessary to meet emissions regulations and mechanical turbocompounding received little attention. However, by the end of the 1990s mechanical turbocompounding was attracting interest again as a means of reducing BSFC, and the availability of high speed electrical machines meant that electrical turbocompounding was also being considered (10-14). Electrical turbocompounding can be achieved in two main ways; either by use of an electric turbocharger or by employing a ‘turbogenerator’. A major challenge with electric turbochargers is that their ability to reject the heat developed within the constrained physical size of the switched reluctance stator of the motor/generator can be insufficient. However, Bowman Power Group (15) has successfully developed a turbogenerator wherein a power turbine is coupled to a high speed permanent magnet electric alternator downstream of the engine’s turbocharger(s). A major advantage of this system is that it is possible to maintain high efficiencies across all wheels, therefore increasing the benefit. Through various research programs, turbocompounding has been shown to reduce BSFCs. Figure 1 shows how it has decreased over time; turbogenerating is the next step in continuing this trend, although this does not indicate the development of the emission regulations against time which tend to increase the BSFC. 1.2 Turbocompounding Theory 1.2.1 Benefits of Turbocharging Figure 2 shows the difference between a naturally aspirated ideal engine cycle, and a turbocharged ideal engine cycle. These two compression ignition engines have the same swept volume and the same maximum cylinder pressure. The difference between them is that one has forced induction at 1 bar gauge while the other is naturally aspirated, consequently they have different clearance volumes in order to maintain the same peak cylinder pressure.

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Cylinder Pressure (Bar)

Figure 1. Specific fuel consumption of turbocompounded engines throughout history 70.0 60.0 50.0 40.0 30.0 20.0 10.0 0.0

Naturally Aspirated Turbocharged

0

100

200

300

400

500

600

Cylinder Volume (cc) P-V Trace Figure 2. Comparison of a turbocharged and a naturally aspirated P-V Turbocharged Cylinder Volum e (cc) combustion cycle

The turbocharged cycle starts at a higher pressure and density, meaning more fuel can be burned during combustion and in turn increases the engine’s power output. This is shown on the diagram by the area enclosed by the turbocharged line being larger than the area enclosed by the naturally aspirated line. 1.2.2 Exhaust Energy Now consider Figure 3. This is a simplified P-V diagram for a turbocharged 4 stroke engine. During the intake stroke 8-1, the air induced into the cylinder is at a higher pressure, Pin, than ambient pressure, Pamb. Compression, 1-2, combustion, 2-4 and expansion occur as normal with the exhaust valve opening at point 5. Note that point 5 is at pressure Pex which is greater than ambient pressure. This produces a positive pressure difference across the turbine with the manifold pressure remaining nearly constant during the exhaust stroke, process 6-7. During the exhaust stroke, the turbine is extracting work from the exhaust gas in the manifold by expanding it from pressure Pex down to pressure Patm. Thus, area 67-12-11-6 on Figure 3 represents the exhaust energy available to the turbine. As is the case for a naturally aspirated engine, the maximum energy theoretically available is considerably greater than this. If the piston were allowed to expand until the pressure in the cylinder was equal to Pex at point 9, area 5-9-6-5 would represent the additional energy that the piston would be able to extract from the exhaust as a result of this ideal expansion. When the piston starts its exhaust stroke, it could be assumed that the manifold pressure falls to ambient, Pamb. If this occurs, the turbine inlet pressure would fall from Pex at point 9 to Pamb at point 10 meaning area 6-9-10-11-6 represents the further additional energy that is available if a complete turbine expansion process is possible.

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3

P

Theoretical exhaust energy available (5-10-11-5)

4

Theoretical exhaust energy available to turbine following a complete turbine expansion using a constant pressure system (6-9-10-11-6)

2

Actual exhaust energy available to turbine using a constant pressure system (6-7-12-11-6)

5 Pin Pex Pamb

1

8 7

9

6

10

12

11

TDC

V

BDC

Figure 3. Ideal cycle of a turbocharged engine It can also be seen in Figure 3 that there are two distinct areas representing the energy available in the exhaust gas. These two areas are the ‘blow down’ energy, area 5-10-11-5 and the exhaust stroke of the piston, area 6-11-12-7-6. Therefore, the maximum available energy to drive a turbine is the sum of these two areas. 1.2.3 Turbogenerating Turbogenerating reciprocating engines is the process of converting wasted energy in the exhaust gasses of turbocharged into electrical power by the use of an additional power turbine which is connected to a small generator. Figure 4 is a simplified T-S diagram for an ideal constant pressure turbocharging process. Process 1-2 is the compressor work and process 4-5 is the turbine work. The work values for these two processes are equal (aside from turbocharger mechanical losses) otherwise the turbocharger would accelerate or decelerate. Exhaust Valve Opening Pressure

2500000

T

3

2000000

Exhaust Manifold Pressure

4

1500000

Turbine Inlet Pressure

Boost Pressure 1000000

5

2 500000

Ambient Pressure

1 0 9

1

1

1

1

1

2

2

2

2

S

2

Figure 4. T-S diagram of an ideal constant pressure turbocharging process Figure 5 is a simplified T-S diagram for a turbocompounded system. Process 1-2 is the compressor work, 4-4a is the turbocharger turbine work and 4a-5 is the power turbine work. Figure 5 shows that by adding another turbine, more of the exhaust energy can be recovered (shown in Figure 6), it can also be designed to provide a pressure differential between intake and exhaust manifold suitable for driving high pressure EGR. The main difference between conventional turbocharging and turbocompounding is shown when comparing Figure 4 and Figure 5. Notice in Figure 5 that the exhaust valve opening pressure is closer to the exhaust manifold pressure meaning that less of the exhaust energy is lost during the blow down when the exhaust valve opens.

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10.1243/17547164C0012010002 2500000

Exhaust Valve Opening P6 (exhaust manifold pressure)

T

3 4

2000000

P6a (inter turbine pressure) 1500000

4a

P1 (boost pressure) 1000000

2

5

500000

P11 (ambient pressure) 1

0 9

1

1

1

1

1

2

2

2

2

S

2

Figure 5. Simplified T-S diagram of the turbomachinery of a turbocompounded engine system 3

P

Theoretical exhaust energy available (5-10-11-5)

4

Theoretical exhaust energy available to turbine following a complete turbine expansion using a constant pressure system (6-9-10-11-6)

2

Actual exhaust energy available to turbine(s) using a constant pressure system (6-7-12-11-6)

5

Pex Pin

7 8

Pamb

12 TDC

6

1

9 10

11 V

BDC

Figure 6. Ideal cycle of a turbocompounded engine Figure 6 is a simplified (P-V) diagram for a turbocompounded engine system. Comparing Figure 6 to 3 shows again that by adding an additional turbine, less of the exhaust energy has been wasted during the exhaust process (area 5-9-6-5 in Figure 6 is smaller than area 5-9-6-5 in Figure 3), thus reducing the amount of lost energy. Turbocompounding can increase the engine thermal efficiency by increasing and optimising the exhaust manifold pressure so that when the exhaust valve opens, the amount of energy lost from the blow down pulse is reduced. While this increased exhaust manifold pressure results in more work than is necessary for driving the turbocharger compressor, the surplus is exploited through the high efficiency power turbine of the turbocompounder. The additional pumping work is therefore offset by the recovered power from the turbogenerator, hence, overall thermal efficiency increases. 2. CASE STUDY The engine that is the subject of this investigation is an 11 litre, straight 6 turbocharged, turbogenerated compression ignition engine, which is used to drive an electric generator at a medium engine load using biogas as its main fuel source. Biogas is induced with air prior to the turbocharger compressor and is combusted in the cylinder with the assistance of palm oil being injected at an advanced crank angle (compared to diesel due to its combustion characteristics). The burnt gasses are then exhausted and expanded through the turbocharger turbine followed by the turbogenerator power turbine. 15

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Turbogenerating has been shown to reduce BSFCs by over 10% (14). The subject engine of this paper is already in use in the field and is performing well. Turbogenerting has reduced the dual fuel engine’s fuel consumption by 7.2% when compared to it being mechanically turbocompounded. The overall aim of this research program is to identify the system parameters that affect the overall efficiency of a turbogenerated engine, and to study how the performance can be maximised. To achieve this, a validated model of a turbogenerated engine has been developed. 2.1 The engine tests Tests were carried out on the candidate engine in a fully instrumented test cell. Over 30 temperature readings and 15 pressure readings were recorded, along with engine power, various mass flow rates and turbomachinery measurements. Figure 7 is a schematic of the test cell. The performance of the system was measured at 1500 rev/min, which is the engine’s design operating speed for exporting electricity to the grid at 50Hz. Turbogenerator Electrical Power measurement

Exhaust

T5 P5 Engine

T4 P4

Generator

T3 P3 T2 P2 T1 P1

Pressure and Temperature measurement

Turbocharger Intake Air

Figure 7. Engine test cell schematic Figure 8 shows the variation of in-cylinder pressure over one complete cycle. From this the heat release and Mass Fraction Burned (MFB) curves that are needed for the GT-Power combustion model were calculated as shown below. The heat release curve was obtained from following the equation, as described in reference 16. The in-cylinder pressure results were used as p1, p2, V1 and V2 with n being the gradient of the LOG P-LOG V graph’s combustion line, shown in figure 8, and gamma beign the ratio of specific heats of combustion. n ⎧⎪ ⎛ V1 ⎞ ⎫⎪⎛ V2 V − V1 ⎞ δQR = ⎨ p2 − p1 ⎜⎜ ⎟⎟ ⎬⎜⎜ + 2 ⎟ 2 ⎟⎠ ⎪⎩ ⎝ V2 ⎠ ⎪⎭⎝ γ − 1

16

(1)

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L OG n C ylinder P res s ure (B ar)

Gradient = 1.2727 6 5 4 3 2 1 0 0.00

2.00

4.00

6.00

8.00

10.00

L OG n C ylinde r V olum e (c c )

Figure 8. Log diagram of cylinder volume versus cylinder pressure The mass fraction burned curve, shown in Figure 9, was then found using the following equation from reference 16. The desired data, such as burn delay and burn duration were then input to the GT-Power model. θ

Mas s F rac tion B urned

Bθ =

∑ δQ 0

Rθ

(2)

QR

1.2 1.0 0.8 0.6 0.4 0.2 0.0 -10

0

10

20 C rank Ang le (deg)

30

40

50

Figure 9. Mass fraction burned Curve Reliable values of discharge coefficients for the inlet and exhaust valves are important to ensure the accuracy of the model. Reference (17) shows that incorrect coefficient of discharge values can produce errors in the model of up to 10%. Coefficients of discharge for the inlet and exhaust ports were obtained for flow in both directions and over a range of pressure and area ratios as described in (18). Figure 10 shows the test apparatus used. Valve lift profiles and fuel injection timings were also obtained and input to the model. Diffuser Section

Vacuum Tank

Cylinder Head

Settling Tank Flow Control Valve

Data Acquisition System

Mass Flow Meter

Figure 10. Cd Test apparatus

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10.1243/17547164C0012010002 2.2 The engine model 2.2.1 Combustion The model developed was a 1-dimensional time based solution, using a semipredictive combustion sub model. This is more efficient to implement than a fully predictive combustion model and it is reasonable to assume that combustion is not significantly affected when investigating parameters such as intake tuning/design, exhaust tuning design and cam timing etc., particularly since this engine was operating at a fixed speed. The combustion process in this specific engine is not common and full combustion modelling would have required special attention. Since combustion data was available it was deemed adequate to use a semi-predictive combustion model. 2.2.2 Modelling the fuels The intake gas composition by volume is shown in table 1. Table 1. Intake gas composition Constituent

Percentage by volume

Methane Carbon Dioxide Oxygen Nitrogen Hydrogen

54.695 44.100 0.200 1.000 0.005

In order for the model to utilise the biogas, the composition was defined as a mass fraction. A PID throttle controller was employed to regulate the biogas mass flow rate prior to mixing with air at the engine intake. The biogas and air mixture was then compressed by the turbocharger compressor and cooled through a water-toair charge air cooler before entering the intake manifold. Having induced the biogas and air mixture into the cylinder, it was then compressed by the piston with combustion being initiated by the injection of palm oil several degrees before top dead centre. In the model, the palm oil liquid was imposed as a user defined component with a user specified palm oil vapour. This was executed by defining the appropriate attribute values, in particular, the lower heating value. Once the intake gas and palm oil had been defined and added to the model, a check was carried out to confirm that the correct amount of energy was being fed into the system and released during combustion. 3 COMPARISON OF RESULTS This section compares the simulation results obtained from the model to actual test results. 3.1 Combustion The in-cylinder pressure trace from the operating condition with the highest overall system efficiency and the predicted pressure trace from the corresponding simulation are shown in Figure 11. The simulation predicts a higher maximum cylinder pressure than is measured. However, on average, the predicted pressures were within 2% and the model was considered sufficiently accurate.

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(Bar) InInCCylinder ylin d er Pressure P res s u re (B ar

10.1243/17547164C0012010002 140 120

Meas ured

100

S imulation

80 60 40 20 0 -360

-260

-160

-60

40

140

240

340

C rank Angle (deg )

Figure 11. In-cylinder pressure trace Figure 12 compares the measured heat release curve to the one obtained from the simulation. The two lines match well and demonstrate that combustion is initiated at the same point in the simulation as it is in the experimental engine. H e a t R e le a s e R a te (J /d e g )

250

0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0

Measured S imulation

200 150 100 50 0

-40

-20

0

20 C rank Angle (deg)

40

60

80

Figure 12. Heat release curve 3.2 Mass Flow Rates The model mixed the biogas and intake air with the code predicting the resulting gas composition. Table 2 shows the measured flow rates, predicted flow rates and their percentage differences. Comparison of the air flow rate values shows that the PID controller worked well with the code accurately predicting the air flow rate and matching the turbomachinery operating conditions. The palm oil flow rate is a model input, and therefore has no associated error. Table 2. Mass flow rate comparisons Measured Biogas Mass Flow Rate (kg/s)

0.0367

Biogas/Air Mixture Mass Flow Rate (kg/s) Palm Oil Mass Flow Rate (kg/s)

Predicted

% Difference

0.0365

-0.50

0.3345

0.3317

-0.84

0.00186*

0.00186*

*Denotes model input 3.3 Engine performance Figure 13 shows the measured P-V diagram and the simulated P-V diagram of the engine. The measured output power from the engine was 230 kW, while the simulation predicted 231.7 kW, demonstrating agreement within 0.7%.

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In C y lin d er P res s u re (B ar

10.1243/17547164C0012010002

140 120

Meas ured

100

S imulation

80 60 40 20 0 0

0.2

0.4

0.6

0.8

1

Volume/Vmax

Figure 13. Pressure – Volume diagram As can be seen from Figure 13, there is a slight difference in the position of the two traces. This difference may be attributed to the slight difference between the model’s manifold geometries and that of the real engine; however, this difference is minimal and was accepted. One other possible source of error could be the heat transfer within the cylinder. Incylinder heat transfer has a significant effect on the predicted IMEP. Unfortunately heat transfer is difficult to measure in reality and therefore difficult to predict with confidence in the simulation. The common method used to overcome this problem is to use a heat transfer model when calculating the apparent burn rate and then re-use this model within the whole engine simulation. This is achieved in GT-Power using the Woschni heat transfer model. The effects of this model are most significant during combustion but experience among engine simulation researchers (19) notes that this Woschni model can underpredict heat transfer in many situations. However, it is possible to alter the model through various multipliers to adjust the heat transfer. This was required within the model, with the adjustments being minor and within the recommended range, they were therefore deemed acceptable. 3.4 Test data and model data comparison Table 3 compares the main engine test data measurements to the simulation output data for the various turbomachinery components and displays their respective percentage difference. All of the simulation values are shown to be within 2.5% of the measured test values. Turbogenerating improved overall efficiency by 3.3% to 45.3% (this is compared to the same baseline engine). The model predicts an efficiency of 45.7%, this is a difference of only 0.8%. It must be noted that the predicted turbogenerator power difference could be due to the manufacturing tolerances and the assembly procedure of the turbogenerator and power electronics as well as the theorectical models of the losses being different to those in reality (e.g. losses due to bearings, windage, alternator, power electronics and cables). Again, theses inaccuracies are within the limits for the model to be considered sufficiently accurate, but reducing these errors could be the basis for further work.

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Table 3. Turbomachinery test data comparison Main engine electrical output Power (kW) System efficiency Turbogenerator electrical output power (kW) Turbocharger speed (krpm) Compressor inlet temperature (deg K) (T1) Compressor inlet pressure (kPa (abs)) (P1) Compressor outlet pressure (kPa (abs)) (P2) Compressor outlet temperature (deg K) (T2) Compressor reduced mass flow (kgs-1K0.5MPa-1) Intake manifold temperature (deg K) Exhaust manifold pressure (kPa (abs)) (P3) Exhaust manifold temperature (deg K) (T3) Turbocharger Turbine reduced mass flow (kgs-1K0.5MPa-1) Turbogenerator inlet pressure (kPa (abs)) (P4) Turbogenerator inlet temperature (deg K) (T4) Turbogenerator outlet pressure (kPa (abs)) (P5) Turbogenerator outlet temperature (deg K) (T5) Turbogenerator reduced mass flow (kgs-1K0.5MPa-1) Turbogenerator speed (krpm)

Measured 230 45.3 27.1 81.6 293.8 94.18 240.1 401.3 60.87 314.6 308.9 912.8 61.22 163.7 796.7 99.9 709.6 67.89 49*

Predicted 231.7 45.7 27.5 82.8 292.7 93.4 235.0 401.7 60.76 315.8 303.2 910.2 61.70 167.1 801.5 100.0 717.7 67.90 49*

% Difference 0.7 0.8 1.48 1.5 -0.4 -0.8 -2.1 0.1 -0.18 0.4 -1.8 -0.3 0.78 2.07 0.6 0.1 1.1 0.015

* Denotes model input 4. CONCLUSIONS The engine simulation package, GT-Power, has been shown to accurately predict the operation and performance characteristics of a turbocharged, turbogenerating biogas engine. On average, the model predicted the performance parameters with an accuracy of better than 2%. Through actual engine tests, turbogenerating has been shown to reduce fuel consumption by 7.2%. This equates to an overall efficiency of 45.3%. The model calculates an overall thermal efficiency with only a 0.8% difference. However, it is important to know the model’s limitations, it uses semi-predictive combustion modelling which is imposed from a measured in-cylinder pressure trace. This is sufficient for investigating turbomachinery, manifold and valve timing effects but is not considered useful when investigating emissions. If emissions are to be considered, a fully predictive combustion model will have to be used. 5. FURTHER WORK Now that an accurate model has been constructed and validated, focus will turn to the optimisation of the system. Turbogenerating in this application has already been shown to reduce fuel consumption by 7.2% when compared to the baseline engine. The aim now is to improve this further. This will be done by considering different valve timings, injection timings and the impact of compression ratio changes with the findings being presented in subsequent papers. Investigations will be carried out at a broad spectrum of speed and load conditions, which is especially relevant in automotive applications where a large range of engine speeds and loads is encountered. Initial improvements can already be seen in (14) where turbogenerating has been shown to improve BSFC by over 10% over the majority of the engine speeds. © Authors 2010

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10.1243/17547164C0012010002 REFERENCE LIST 1

2 3 4

5 6 7 8 9 10 11

12

13 14 15 16 17 18 19

Thompson, I.G.M., Spence, S.W. and McCartan, C.D, “A Review of Turbocompounding and the Current State of the Art”, 3rd International Conference on Sustainable Energy and Environmental Protection, Dublin, Ireland, August 12-15, 2009, Paper No 101219 Sammons H. and Chatterton E., “Napier Nomad Aircraft Diesel Engine” , SAE Trans. 63, No. 107, 1954. This paper was presented at the SAE Summer Meeting, Atlantic City, June 10, 1954 Wallace, F.J. “The Differential Compound Engine”, Proc. IMechE, 1973, Vol 187, 48/73, p548 Wallace, F.J. and Winkler, G. “Very High Output Diesel Engines - A Critical Comparison of Two Stage Turbocharged Hyperbar, and Differential Compound Engines”, International Off-Highway & Powerplant Congress & Exposition, Milwaukee, Wisconsin, Sept 1977 SAE Paper 770756 Black, J.W., Fox, L.D., French, P.B. and Schwarz, E.E., “Uprate of Cummins V903 Diesel Engine to 1000 bhp for Military Application” SAE Paper 830505, 1983 Kamo, R and Bryzik, W., TARADCOM “Adiabatic Turbocompound Engine Program, SAE 810070, Detroit 1981 Tennant, D. W. H. and Walsham, B. E., “The Turbocompound Diesel Engine”, International Congress and Exposition, Detroit, Michigan, Feb 27-Mar3, 1989 SAE Paper 890647 Brands, M. C., Werner, J. R., Hoehne, J. L. and Kramer, S., “Vehicle Testing of Cummins Turbocompound Diesel Engine, SAE Paper 810073, 1981 Wilson, D. E., “The Design of a Low Specific Fuel Consumption Turbocompound Engine”, International Congress and Exposition Detroit, Michigan, Feb 24-28, 1986, SAE Paper 860072 Leontopoulos, C., Robb, D.A. and Besant, C. B. “Vibration Analysis for the Design of a High-Speed Generator for a Turbo-Electric Hybrid Vehicle”, Proc Instn Mech Engrs Vol 212 Part D, Paper D01497, 1998 Panting, J., Pullen, K. R. and Martinez-Botas, R. F., “Turbocharger MotorGenerator for Improvement of Transient Performance in an Internal Combustion Engine”, Proc Instn Mech Engrs Vol 215 Part D, Paper D04500, 2001 Fieweger, K., Paffrath, H. and Schorn, N. “Drivability Assessment of an HSDI Diesel Engine with Electrically Assisted Boosting”, Seventh International Conference on Turbochargers and Turbocharging, 14-14 May 2002, Paper No C602/009/2002 Hopmann, U. and Algrain, C, “Diesel Engine Electric Turbo Compound Technology”, Future Transportation Technology Conference, Costa Mesa, California, June 23-25, 2003 SAE Paper 2003-01-2294 Vuk, C., “John Deere - Electric Turbo Compounding technology update”, 13th Diesel Engine-Efficiency and Emissions Research Conference, Detroit, Michigan, 2007 Talbot-Weiss, J., “SFC Improvements from Turbo-Generating Heavy-Duty Diesel Engines”, PhD Thesis, University of Sussex, 2009 Blair, G.P., “Design and Simulation of Four-Stroke Engines”, Society of Automotive Engineers, ISBN 0-7680-0440-3, p418-420 Blair, G.P. and Drouin, F.M.M., “Relationship Between Discharge Coefficients and Accuracy of Engine Simulation”, SAE Paper 962527, 1996 Stevenson, P.M., “A Study of factors affecting the coefficient of discharge of twinned poppet-valves”, PhD Thesis, The Queen’s University of Belfast, 1999 GT-Power v6.2 User’s Manual, Gamma Technologies Inc, Westmont, IL, USA, Chapter 3, p23 & 74

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