1. Analysis of Turbocharged V6
Miller Cycle Gasoline Engine
Dylan N. Alesio May 6, 2014
Independent Research
by: Dylan N. Alesio
Advised by:
Dr. Richard Cohen
Temple University
College of Engineering
1947 North 12th Street
Philadelphia, PA 19122
2. Independent Research in Mechanical Engineering: MEE 4191
Analysis of Turbocharged V6 Miller Cycle Gasoline Engine
Dylan N. Alesio
Examiner
Dr. Richard Cohen
Advisor
Dr. Richard Cohen
Contact person
Dylan N. Alesio
May 6, 2014
1. Abstract
In 1993 Mazda introduced a supercharged Miller cycle gasoline engine into the market as an answer to
satisfy low emission, high fuel economy requirements. Fuel economy is enhanced by artificially
downsizing displacement resulting in reduced "pumping losses." The Miller cycle reduces knocking by
lowering the effective compression ratio through late closing of the intake valves and reduces the intake
charge displacement, while retaining high expansion ratio for improved efficiency. Utilizing the Miller
cycle process is an effective way to improve fuel economy and reduce emissions.
In order to realize the potential efficiency of the Mazda developed Miller Cycle engine, the following
modifications were implemented for this research:
1. Replaced the Lysholm supercharger with a properly matched turbocharger to achieve a
reduction of the parasitic drag and improve efficiency by utilizing otherwise wasted exhaust gas
energy.
2. Reduced "pumping losses" by improving intake system flow and removed the dual intercoolers.
3. Enhanced volumetric efficiency via porting and polishing intake and exhaust ports, and
improving exhaust system flow to achieve excellent "acceleration and mid-to-top-end feel."
4. Developed and installed direct ported water/alcohol injection to promote "in-cylinder"
temperature cooling, provide dense air charging, and suppress knocking.
5. Recalibrated fuel and ignition tuning to optimize efficiency and power output.
6. Upgraded fuel injectors with modern high-atomization injectors to improve efficiency.
The Miller cycle engine is a practical method for gasoline engines to improve thermal efficiency. The
turbocharged V6 Miller cycle gasoline engine can be an alternative to the large displace naturally
aspirated (NA) engine because of its equivalent torque performance and lower fuel consumption via the
effects of smaller displacement and reduced "pumping losses." In order to effectively implement the
turbocharged Miller cycle engine, drivability, durability, low cost, high efficiency, and high power all
must be retained. Turbocharging can achieve better performance and fuel economy over supercharging
while the Miller cycle process provides an additional efficiency gain over the Otto cycle process.
Temple University
College of Engineering
3. Table of Contents
1. Abstract............................................................................................................................................2
2. Introduction .....................................................................................................................................5
3. Mazda's Miller Cycle Engine ............................................................................................................6
3.1. Background ...........................................................................................................................6
3.2. Late Closing Intake Valve Timing ..........................................................................................8
3.3. Intake Air System ..................................................................................................................8
3.4. Thermal Load Reduction.......................................................................................................9
3.5. Combustion Improvements ..................................................................................................9
3.6. Performance..........................................................................................................................9
4. Problem Statement........................................................................................................................10
4.1. Limitations...........................................................................................................................10
5. Approach: Modification to Further Boost Efficiency .....................................................................11
5.1. Vehicle Selection.................................................................................................................11
5.2. Overview: Modifications to Miller Cycle Engine.................................................................13
5.3. Reducing Parasitic Drag.......................................................................................................14
5.4. Engine Rebuilding Methods and Modifications..................................................................16
5.5. Engine Modifications ..........................................................................................................17
5.6. Transmission Modifications................................................................................................20
5.7. Wiring, Electrical, and ECU..................................................................................................21
5.8. Air Intake Manifold Fabrication and Modifications............................................................22
5.9. Air Charge System Enhancements ......................................................................................23
5.10. Exhaust System Enhancements ......................................................................................24
5.11. Fuel Enhancements and Control Methods .....................................................................24
5.12. Air Charge Cooling Enhancements..................................................................................25
6. Results............................................................................................................................................27
6.1. Testing Methodology and Equipment ................................................................................27
6.2. Performance........................................................................................................................27
6.3. Brake Specific Fuel Consumption (BSFC) ............................................................................29
6.4. Brake Mean Effective Pressure (BMEP) ..............................................................................29
6.5. Brake Thermal Efficiency.....................................................................................................30
6.6. Highway Cruising Efficiency ................................................................................................31
4. 6.7. Intake Air Charging Cooling Effectiveness ..........................................................................31
6.8. Turbocharger Compressor Efficiency..................................................................................32
6.9. Turbocharger Turbine Restriction.......................................................................................33
7. Discussion and Summary ...............................................................................................................34
7.1. Performance and Fuel Economy.........................................................................................34
7.2. Turbocharging The Miller Cycle Engine ..............................................................................35
7.3. Water injection Air Charge Cooling.....................................................................................35
7.4. Optimizations......................................................................................................................36
8. Acknowledgement.........................................................................................................................37
9. Nomenclature................................................................................................................................37
10. References .....................................................................................................................................37
5. Figure 1: The Miller Cycle Process
Intake:
25%
No Compression: Compression: Expansion: Exhaust:
75%
2. Introduction
Throughout Mazda's history, Mazda developed various non-conventional engines including the rotary
(Wankel) engine. In the mid-90's, Mazda developed the Miller cycle engine based off their K-series Otto
cycle engine. The Miller cycle follows similar principles to that of the Otto cycle, but with the addition of
a "fifth cycle" delaying the closing of the intake valves during compression cycle. The Miller cycle valve
timing effectively reduces the compression ratio while retaining the nominal expansion ratio determined
by the physical geometry of the engine. The intake cycle begins as the piston moves down pulling an air
charge and fuel mixture into the cylinder. The intake valve remains open during a portion of the
compression cycle causing a portion of the air charge and fuel mixture to exit the cylinder and renter the
intake manifold for the next cycle (see Figure 1). Compression begins when the intake valves close, this
equates to about 75 percent of the original maximum displacement. The expansion cycle is effectively
longer than the compression cycle, resulting in an increase in efficiency. According to Mazda, the
expansion ratio has a greater effect on efficiency than the compression ratio thus making the Miller
cycle engine ideal for efficiency.
The larger expansion ratio allows more of the combustion energy to be converted to mechanical energy.
The Miller cycle process improves efficiency by reducing "pumping losses" at the expense of reducing
power-density. Mazda overcomes the power-density issue by implementing a positive displacement
Lysholm supercharger to boost volumetric efficiency, resulting in 219 hp and 210 lbs-ft torque from 2.3L
of displacement. The Miller cycle process can be induced by early or late closing of the intake valves.
6. Both the Miller cycle and Atkinson cycle engine achieve a
greater expansion ratio than compression ratio by different
means. The Atkinson cycle is developed by James Atkinson in
1885 and used additional linkages in the crankshaft to allow
all four cycles to perform with one rotation of the crankshaft.
The geometry of the additional linkages allow for different
physical compression stroke length and expansion stroke
length. The additional linkages of the Atkinson cycle engine
are subjected to additional stresses and wear that make the
Miller cycle engine more appealing to the automotive
industry. Named after Ralph Miller, the Miller cycle engine
achieves a greater effective expansion stroke than the
effective compression strokes through either late or early
closing intake valve timing. During the intake cycle, a full air
charge is draw into the cylinder but then a portion is pushed
back into the intake manifold. Performance losses are
overcome with the addition of forced induction.
In Figure 2, the area under the curve of a pressure verse volume diagram corresponds to the work
performed. Figure 2 displays the additional work performed by the Miller cycle process by the
shaded/hatched area. The intake valve is open during compression from 6 and close at 1. The
compression cycle begins at 1 and ends at 2. Then the combustion cycle begins from 2 to ends at 3.
Expansion cycle begins from 3 and ends at 5.
3. Mazda's Miller Cycle Engine
3.1. Background
Mazda's K-series engines are 60° V6, 24 valve DOHC, short stroke, and all aluminum-alloy construction. A
unique feature of the Mazda's K-series engines is the ultra rigid split crankcase design usually found in
high performance engines. Many of the components are shared across the series. Features of the K-
series engines include internally balanced forged steel crankshaft, lightweight powder forged carbon
steel connecting rods, all-alloy, split-crankcase, 24 valves, and DOHC. See Table 1 for the K-series engine
specification.
Table 1: Mazda K-Series V6 Specification (America Market)
KL KJ KF K8
Cycle Otto Cycle Miller Cycle Otto Cycle Otto Cycle
Displacement (L) 2.497 2.254 1.995 1.845
Bore x Stroke (mm) 84.5 x 74.2 80.3 x 74.2 78.0 x 69.6 75.0 x 69.6
Effective Compression Ratio 9.2 7.6 9.4 9.4
Expansion Ratio 9.7 10 10 10
Int. Valve Open 8° BTDC 2° BTDC 5° BTDC 6° BTDC
Int. Valve Close 47° ABDC 70° ABDC 35° ABDC 37° ABDC
Exh. Valve Open 50° BBDC 47° BBDC 49° BBDC 49° BBDC
Exh. Valve Close 5° ATDC 5° ATDC 5° ATDC 6° ATDC
Power (hp), Torque (lbs-ft) 170, 160 217, 210 140, 132 130, 115
Configuration Aluminum block and heads , 60° V6, 24 valve DOHC
Figure 2: PV diagram Otto vs. Miller Cycle
Cylinder Pressure vs. Volume
Otto Cycle
Miller Cycle
7. The Mazda Miller cycle engine is derived from the Otto cycle versions of the K-series engine. A Lysholm
twin screw supercharger is developed by IHI and placed in the center of the engine's 'V' to supply the
engine with 14.7psi of boost pressure(see Figure 3). Dual intercoolers are fitted one per engine bank to
cool the hot air charge generated by forced induction. The intake valves of the Miller cycle engine closes
at 70° ABDC versus the Mazda Otto cycle's 35-47°. The resulting effective compression ratio of 7.6 and
expansion ratio of 10 was carefully chosen by Mazda.
During the development of the KJ Miller cycle engine, Mazda focused on revising various systems to
achieve high fuel economy, low emission, and proper driving performance for the 3.0L vehicle class. The
Miller cycle process is achieved through late closing timing of the intake valves rather than early closing
timing. Mazda fitted the air induction system with a Lysholm compressor, dual intercooler, and air-
bypass valve to improve drivability. Combustion chamber geometry and valve angle were optimized to
generate a targeted tumble ratio to achieve a desired balance between rate of combustion and
scavenging effects.
Figure 3: Mazda Miller cycle engine display, Lysholm supercharger, and intercoolers diagram
8. 3.2. Late Closing Intake Valve Timing
The Miller cycle process experiences a longer expansion stroke than effective compression stroke, which
contributes to high brake thermal efficiency. The Miller cycle process can be realized by using either
early closing intake valve timing (ECIVT) or later closing intake valve timing (LCIVT). The method of
closing valve timing results in different effects. ECIVT shortens the intake cycle period, greatly reducing
the volume efficiency as engine speed increases, thus requiring higher intake air charging. High pressure
intake air charging can increase thermal load beyond manageable range, making LCIVT appropriate for
automotive applications. LCIVT decreases in cylinder charge temperature during the compression
stroke, reduces knocking (irregular/premature combustion), which improves combustion quality. Due to
higher knock limit, ignition timing advancement can be realized.
According to Mazda's study, "At the same BMEP, the more the valve closing timing retards, the more the
exhaust gas temperature lowers. This will lead to superior fuel economy at high load and high speed
range since rich mixture to protect exhaust system is unnecessary."
3.3. Intake Air System
To overcome the reduced volumetric efficiency of the Miller cycle process, a Lysholm compressor
charges the air intake system up to 15 psi (1 atmosphere) gauge pressure. The compressed air heats up
and requires sufficient charge cooling. Dual intercoolers, arranged one per cylinder bank, are
implemented to provide adequate charge cooling. Air flow control is managed with a throttle valve
placed pre-compressor and a compressor air bypass valve (ABV) placed to divert excess charged air to
the inlet side of the supercharger. To achieve desired torque output during heavy throttle input, the
intake air charge is modulated by the ABV, reducing the peak charge pressure and tapering
instantaneous torque to
provide a more linear
torque curve. During light
throttle conditions, a
charge air cooler bypass
valve (CAC-BV) is used to
bypass the intercoolers to
reduce heat soak and
pumping losses(see Figure
4). Mazda claims the
intercoolers reduce the air
charge from 90-150°C
195-300°F) down to 50-
60°C ( 122-140°F) during
WOT acceleration providing
adequate air flow.
Figure 4: Mazda's Air Charger Routing Modes
Components: [1] Throttle body, [2] Supercharger, [4] Intercooler, and [5] Intake manifold
Possible Air Charge Routing Paths:
Regular mode: [1] [2] [3a] [4] [5]
Intercooler Bypass mode: [1] [2][3b][5]
Supercharger Bypass mode: [1],[3b] [5]
9. Mazda selected a Lysholm supercharger over other superchargers or turbochargers do the required
"near instant" response and high air charge pressure needed. Root-type blowers are able to produce a
"near instant" response but at only half the pressure. Mazda and Ishikawajima Harima Heavy Industry
(IHI) developed a Lysholm compressor for the Miller cycle engine. The Lysholm Compressor achieves
over 70% adiabatic efficiency which is considerably higher than Root-type blowers thus greatly reduces
air charge temperatures during wide open throttle conditions. When considering any form of forced
induction, careful attention to thermal load temperatures must be addressed.
3.4. Thermal Load Reduction
Forced induction engines generate additional heat that can be damaging to engine components
including pistons, valves, cylinder heads, etc. To address concerns, Mazda implemented oil cooled
pistons with a cooling channel near the top of the piston. Strategically placed oil jets fill the cooling
channels reducing piston temperature by 25°C (77°F). An additional 25°C reduction in piston
temperature is noted from the Miller cycle process. When compared to a NA Otto cycle engine, a total
of 50°C (122°F) temperature reduction of the piston is realized, thus further increasing knock limit.
3.5. Combustion Improvements
The "quality" of combustion can be influenced by various methods to generate a desired air flow
motion, whether it is turbulent, laminar, swirl (rotation parallel to axis of flow), or tumble (rotation
perpendicular to axis of flow). Mazda adopted a tumble air motion to further improve combustion by
accelerating initial and main stage mass burn rates. A designed tumble ratio of 2:1 (in-cylinder angular
air speed moves twice the angular speed of the crankshaft at BDC) is selected to improve in-cylinder
scavenging effects and optimize the burn rate. A shrouding mask surrounding the intake valves is added
to the combustion chamber to aid in-cylinder tumble motion while also concentrating air toward the
center of the combustion chamber. The angle of the intake runner with respect to intake valve, and
valve to mask clearance was studied and optimized to achieve the desired tumble motion of incoming
air introduced into the cylinder. Mazda notes "Combined with compact combustion chamber, this
optimum tumble ratio contributes to enhancing engine anti-knocking performance."
3.6. Performance
The supercharged Miller cycle engine has high torque performance especially in the low and mid range
engine speed. Due to its far superior torque performance over a NA 3.0L V6, displacement downsizing to
2.3L further reduces losses including piston ring friction and air pumping losses, thus improving fuel
efficiency over the larger displacement engines. The reduction of displacement reduces thermal and
frictional losses; including piston rings to cylinder wall interface, air intake pumping losses, etc. Mazda
claims a friction loss reduction of 25% when comparing the supercharged 2.3L KJ Miller cycle engine to a
NA 3.3L engine of similar performance. High efficiency is realized due to the over expanded stroke of
Miller cycle process.
Brake specific fuel consumption (BSFC) can be used as a quantitative comparison between different
engines and their ability to produce power with respect to its fuel consumption (fuel efficiency). Mazda
compared the BSFC of the Miller cycle gasoline engine, a NA Otto cycle gasoline engine, and a
turbocharged indirect-injection (IDI) diesel engine producing the same maximum torque. The study
10. shows that the Miller cycle engine has a BSFC 10-15percent lower than the NA gasoline engine through a
load range of 20-60 percent. Under normal load condition, the V6 Miller cycle gasoline engine has an
equivalent mass fuel consumption level to a turbocharged IDI diesel engine.
4. Problem Statement
This research aims to examine the potential of the gasoline turbocharged Miller cycle engine. Significant
revisions are made to Mazda's 2.3L V6 supercharged Miller cycle engine to attain the appropriate
configuration for this study. Superchargers are crankshaft driven air compressors that are parasitic in
nature, whereas a turbocharger utilizes wasted exhaust gas energy. The most notable revision is
substituting the supercharger with a turbocharger. Investigation of BSFC at various boost pressures will
be studied to determine the highest efficiency conditions and thus correlate an ideal engine sizing for
various automotive applications. The Miller cycle engine's effective charge volume is reduced from air
charge reversion during a portion of the compression cycle and is compensated with forced induction
turbocharging to achieve the same performance of a larger engine. In order to operate efficiently,
optimization of the turbocharger, operating boost levels, and fueling all must be considered. When
compared to the conventional Otto cycle engine, the Miller cycle engine's exhaust gas energy is reduced
and thus matching the turbocharger size will follow a different protocol. The flow rate of the original
Lysholm supercharger will be considered when selecting the turbocharger's compressor flow
characteristics. Real world testing must be performed to verify results and determine drivability. The
test engine will be fitted into a road legal vehicle and must comply with all emissions standards for a
1997 OBD-ii system. This becomes a real challenge when implementing full control of fueling and
ignition timing to achieve the desired goals. Careful considerations must be made to attain road legal
status. Drivability and durability must also be considered during all phases of the research in order to
evaluate the worthiness of the turbocharged V6 Miller cycle gasoline engine.
4.1. Limitations
The test engine must comply with standard DOE emission regulations. This allows the engine to be
tested on the roadways and real world road testing results can measured. The current fuel system flow
is limited to a total maximum of 1,602 cc/min thus limiting power production. Effective compression
ratio and expansion ratio are fixed at 7.6:1 and 10:1 respectively. Intake valve closing timing is fixed at
70°. Boost pressure is variable with a minimum of 3-4 psi and a "safe" 15 psi upper limit. The strength to
the internals will determine the upper boost limit. The stock pressure of the supercharger is set to 15 psi
by Mazda, thus selecting the current upper boost limit for the turbocharger. Concerns of the engine's
ability to withstand a 30 percent torque increase to 273 lbs-ft torque at the same boost levels are
prevalent.
The majority of the budget is self-funded; therefore limitations exist on abilities to acquire specialty test
equipment for data acquisition. The majority of the work is also self-performed other than welding and
chassis dynamometer testing. This research is currently an independent research project under the
guidance of Temple University College of Engineering.
11. 5. Approach: Modification to Further Boost Efficiency
In search of greater fuel efficiency, many automotive manufacturers are downsizing engines and
implanting modern technologies of turbocharging, variable valve timing and direct fuel injecting to
enhance performance and fuel efficiency. These technologies do improve fuel efficiency, but the cost to
implement them may not be worth the small fuel efficiency improvement. Direct fuel injection has fuel
pressure in the range of 1,500 to 3,000 psi compared to port fuel injection pressure of 40 to 100 psi.
These high fuel pressure systems have to deal with high pressure and high heat challenges which
increases manufacturing and servicing cost. Variable valve timing also has additional complexities that
equate to greater costs. A current trend that many engine manufactures are implementing is downsizing
displacement and turbocharging. In light load conditions, the turbocharger may provide a minute
amount of additional air while providing a reduction in "pump losses" due to the slightly more positive
pressure when compared to NA engine. This can reduce parasitic drag of pumping air past the throttle
body and intake valves.
Engine manufactures must evaluate the worth of these technologies with cost and fuel efficiency gains
in mind. These technologies are known to improve power production with respect to displacement but
in many cases show very little real world fuel consumption improvement. Turbochargers are essentially
an air pump (compressor) driven by a turbine impeller in the exhaust gas stream. The drive force to
spool the turbine does not come free. During high load condition an increased amount of exhaust
backpressure can be seen which essential puts more load on the engine to drive the exhaust gas out of
the cylinders. When compared to a supercharger, a compressor driven directly off the crankshaft, a
turbocharger is notably less parasitic. Additional manufacturing cost are associated forced induction
engines mainly due to the additional components and complexity. There are many factors of the
equation when manufacturing a fuel efficient engine.
In order to ensure that the test engine is "healthy", a full rebuild is performed to ensure proper
tolerances are achieved. Several modifications are made to various engine components to improve flow
and resist carbon build-up. The stock fuel injectors are replaced with modern units that provide
increased atomization of the delivered fuel. Additional control and hardware components are
implemented to provide drivability. Durability is equally important, thus influencing decisions
throughout this study.
5.1. Vehicle Selection
In order to exploit the fuel efficiency gains of the modified turbocharged Miller cycle engine, a proper
chassis must be selected. The Miller cycle engine is only found in the Mazda Millenia S (see Figure 6).
The Mazda Millenia S weighs in at a hefty 3,410 lbs, making it impractical for the desired performance
and fuel efficiency project goals. Since simplicity of installation and cost of vehicle are factors of concern,
the second generation Ford Probe GT is the prime candidate due to sharing of Mazda's K-series 60° V6
engines. The K-series engines share many similarities. The Ford Probe GT weighs approximately 2,900
lbs, has a drag coefficient of 0.33, and a relatively small frontal area making it a reasonable chassis for
the intended goals. The sporty appearance of the Ford Probe GT and the cross compatibility of the
12. Figure 5: Test vehicles fitted with turbocharged Miller cycle engine
Test Vehicle version 2
Test Vehicle version 1
Figure 6: Mazda Millenia S
components became positive to the vehicle selection process. Figure 5 show the Ford Probe fitted with
the turbocharged Miller cycle engine.
13. 5.2. Overview: Modifications to Miller Cycle Engine
Reducing Parasitic Drag
Replaced supercharger with properly sized turbocharger unit
Removed vacuum pump
Rebuilt engine to ensure OEM specification
Installed engine in a smaller, lighter, and more aerodynamic vehicle
Engine Rebuild Methods and Modifications
Full cleaning and rebuild of engine.
o Cylinder honing/deglazing
o Valvetrain reconditioning and cam journals clearance adjustments
o Tappets (Bucket lifter) shims adjustments
o New bearing and seals
Engine Modifications
o Installed uprated (stiffer) valve springs for 7000 rpm rev-limit
o Ported and polished cylinder head ports
o Polished all combustion surfaces to a mirror finish
o Oil pump fitment modification
o Revised and remounted belt driven for accessories
o Fabricated and installed engine V-valley plate and coolant route plumbing
Transmission Modifications
o Taller Final drive (3.85 vs. 4.388)
o Taller and wider gear ratios (3rd
, 4th
, and 5th
gears)
o Installed Limited slip differential to ensure even power delivery to drive wheels
Wiring, Electronics, and ECU
o Stock KJ engine control unit ECU wired for initial start and operation
o Fabricate and wire ECU patch adapter for plug and play to stock wire harness
o Acquired data from OBD-ii port for developing ignition timing map
o Established sensor calibration curves.
Air Intake Manifold, Air Charge System, and Exhaust System
Air Intake Manifold fabrication and Modifications
o Reduced intake air plenum volume by placing throttle body just before the
intake plenum to improve throttle response
o Reduced intake air restriction and improve engine packaging
Air Charge System Enhancements
o Simplified and designed suitable air charging system for turbocharger
o Removed twin-intercoolers
o Reduced intake air restriction and improve engine packaging
o Properly size air charge piping diameter
o Implemented smooth plumbing transitions
14. o Removed all air intake noise suppression control devices for improved racy
sound
Exhaust System Enhancements
o Implemented high flow tubular heads
o Modify Turbocharger pipes and plumbing for application
o Properly size exhaust piping diameter
o Implemented smooth plumbing transitions
o Installed a high flow quiet muffler
o Rerouted turbine bypass (wastegate) back into muffled exhaust system
Fuel Enhancements and Control Methods
Implemented and tuned a full standalone ECU.
Steady-state driving (highway) lean burn mode a:f → 18:1 and higher)
Modern high atomization 12-hole Denso fuel injectors (50 micron droplets)
Improved drivability by implementing variable boost pressure based on throttle input
Air Charge Cooling Enhancements
Implemented direct port water/alcohol injection to manage in-cylinder temperatures
Implemented pre-compressor water/alcohol injection to aid air charge cooling and
increase compressor efficiency
5.3. Reducing Parasitic Drag
The battle between turbocharging versus supercharging has been a topic of discussion in the automotive
world since their introduction. Selecting the proper method of forced induction became dependent on
application and desired results. Superchargers are well known for their instant throttle response and
massive low rpm torque. Directly driven off the crankshaft, the supercharger places an additional load
on the engine. This is known as parasitic drag. Since turbochargers are driven by high energy wasted
exhaust gases, parasitic drag is virtually eliminated other than the opposing force to drive the exhaust
gas out past the valves and through the turbine housing. Exhaust backpressure increases upstream the
turbine based on the loading condition of the engine.
The IHI Lysholm supercharger requires up to 50 hp to generate 15 psi of boost in the 2.3L V6 KJ engine.
Essentially, 50 hp worth of extra fuel is required to drive the supercharger, thus reducing BSFC.
Turbochargers do not contribute to high parasitic drag, but, as with all systems, tradeoffs can be found,
specifically turbo lag. The delay of pressure charging when the throttle is tipped-in under heavy load
conditions is known as turbo lag. During this transient state, turbo lag is influenced by pressure
differentials and rotational mass of the compressor and turbine system. When requesting additional
power from a properly sized turbo, a typical range of lag time of 0.25-3 seconds can be measured. Turbo
lag is significantly reduced as engine speed increases. Turbochargers are dependent on pressure
differentials in the intake and exhaust system. Basically, turbochargers do not pump additional air unless
the engine is under load, in which a pressure differential is created and thus the turbo spools generating
positive pressure in the air intake system. Turbochargers can provide "soft" power that is based on
loading conditions making it possible to replace larger displacement NA engine with small displacement
turbocharged engine.
15. With this concept in mind and the primary goal of fuel efficiency, turbocharging becomes a clear choice
for many automotive manufacturers. The turbocharger's compressor is selected based on mass flow
rate provide by the original supercharger. The selected turbocharger is a Garrett T3 0.42ar 45 trim
compressor and a 0.48ar turbine housing. Due to the over-expanded stroke of the Miller cycle engine,
exhaust gas pressure and temperature are reduced compared to the Otto cycle engine. This must be
accounted for when selecting the turbine housing size. Figure 7 show the turbocharger fitted to the V6
Miller cycle engine.
Mazda's supercharged Miller cycle engine is equipped with a mechanical vacuum pump that operates
various pneumatic diaphragm valves controlling the supercharger bypass valve and several electric
solenoids that switched air flow operating modes. In order to properly implement the turbocharger
swap, the air intake system, air charge system, and various other systems are revised including the
omission of the vacuum pump. Although minute, drag loss is reduced with the deletion of the vacuum
pump and simplifies belt arrangement.
A full engine rebuild was performed and various modifications were made during this stage to further
enhance efficiency gains. Pistons and connecting rods were reused and weight matched to ensure
proper balancing of the rotational assembly. All heat surfaces, including valves faces, combustion
chamber, and piston tops were polished to mirror-finish. Cylinder bores were deglazed with a honing
brush to provide proper piston ring seating surfaces. New main bearings, journal bearings, and seals
were installed. Porting and polishing of the cylinder heads were carefully performed to improve air
Figure 7: Turbocharger system
16. intake and exhaust flow. Valves were lapped to provide proper sealing to valve seat. The health of the
engine was restored to factory specification to ensure optimal performance.
5.4. Engine Rebuilding Methods and Modifications
The cost to rebuild engines can be very expensive based on the specialty tools, machines, and labor. Due
to financial constraints all engine rebuilding work is performed "in-house." A two stage (coarse and fine)
brush-honing technique is selected to ensure proper oil retention on the cylinder walls. With the proper
reciprocating rate and drill speed, cross hatching marks around 60° are generated. After coarse brush
hone work is performed, a finer hone brush is used to ensure a smooth wall surface. After the two-stage
brush honing, the cylinder wall surface will have a very smooth contact surface caused by the fine hone
brush and shallow groves for oil retention caused by the coarse hone brush. Appropriate cleaning
techniques were used to remove any remaining metals and abrasive brush hone materials. The cleaning
stage is critical and must be done properly. Each cylinder bore runout is checked with a dial bore gauge
after brush honing. All critical dimensions of the crankshaft and cylinder block were measured and
checked. High-quality tri-metal bearings were selected for the main and connecting rod bearings.
Cylinder heads have many moving parts that require tight tolerances. Extreme cleaning to the valves is
performed due to excessive carbon buildup. The valves showed signs of light pitting on the seating face,
thus requiring valve lapping to achieve proper valve and seat sealing. To check if valve lapping work was
performed correctly, a leak down test was performed. The valve and springs are installed in cylinder
heads. Gasoline is then poured into the ports and monitored the following day. Valve re-lapping is
required if any ports showed signs of leakage. Disassembling, re-lapping, assembling, testing, and
repeating as needed is a time intensive process that can easily be overlooked by many builders.
Overhead cams are held into place with aluminum cam journal caps that can wear and become loose. If
cam journals caps do not meet clearance specifications, oil pressure may drop below specification
resulting in inadequate oil that could led to engine failure. Cam journal caps clearance was checked with
Plastigauge. To tighten up clearance, material is removed off the bottom of the cam journal caps and
recheck with Plastigauge until specification is attained. This process is labor and time intensive and if
performed correctly can ensure proper oil pressure in all conditions. According to various KL engine
builders, if a KL engine suffers from low oil pressure conditions, it is usually caused by loose cam cap
clearances. Connecting rod bearing failure in cylinder number six is a common failure known to the KL
engines when oil pressure is low under high-load conditions. With this concept in mind, precision work is
performed to reduce risk of inadequate oil pressure.
Each of the 24 valves on the KJ engine cylinder heads contain 8 parts, totaling to 192 components
consisting of the valvetrain (not including camshafts, caps, and blots). Containing nearly 200
components, the valvetrain requires time intensive labor to recondition to proper specification. Mazda
opted to use solid bucket lifters for higher performance over the self-adjusting hydraulic lifters. Cam
lobe to bucket lifter clearance was measure with feeler gauges and proper thickness shims were
installed to provide proper clearance. The processes starts with a fully assembled head. Cams are then
rotated until the valve is in closed position and feeler gauge then measures and records the clearance.
Exact shim size can be calculated after measuring. After all 24 lifters are measured, the camshaft is
17. removed, calculated shims are installed, and camshaft reassembly are performed. Clearances are
checked and, if required, readjustments are performed. When comparing to similar engine, the Mazda
KJ engine has a very small lifter tolerance range making this adjusting process more strenuous.
The purpose of this section is to provide a general layout of the rebuilding process and approach to
achieving the best possible "home build" without accruing the cost of professional engine builders. A
typical rebuild for a V6 engine starts around $3,000 for parts and labor. By taking the "in-house build"
route, the cost of labor can be eliminated thus making the rebuilding process financially attainable.
5.5. Engine Modifications
Automotive enthusiasts who enjoy performance driving commonly want to extract every bit of power
from their engine by bolting on every available aftermarket part that they can afford. Typical power
boosting upgrades common to "aftermarket tuners" consist of air intakes, exhaust systems, cams, ECU
remapping, etc. More extreme modifications include turbocharging or supercharging. As the "build"
becomes more extreme in terms of generating power, it is common to see cylinder head alterations
including combustion chamber reshaping, enlarging the intake and exhaust ports, polishing the exhaust
ports, enlarged valves, etc. This stage can accumulate massive expenses in machining and components.
If done correctly, these aftermarket parts can work in unison with one another and greatly improve
performance over the stock system. Automotive manufacturers generally design car with cost,
durability, and drivability in mind; thus making a highly tuned engine too expensive to manufacture and
maintain. Also, the expected life of the engine shortens and the manufacturer warrantee support
diminishes. The cost of power comes with a high price and usually high fuel consumption. It seems that
the fuel efficiency and power are at the opposite ends of the spectrum. With the correct methodology, a
middle ground may be attainable.
During this phase, decisions are carefully made knowing that Mazda engineered the KJ engine to
perform in a desired manor. The
internal components of the engine
are design to handle high power
from forced induction. The
crankshaft is forged steel rather
than cast steel making it ideal for
high power reliability. The
connecting rods are also made of
forged steel. And finally the pistons
are oil cooled to reduce thermal
load. Basic analysis of the bottom
end and corresponding components
is performed resulting with no need
to upgrade due to the robustness of
each part. Only minor modifications
are performed to the bottom end.
Figure 8: Polished Combustion Chamber and Valves
18. An increased number of modifications are performed on the cylinder heads.
Each piston is weight matched to ensure proper balancing of the rotating assembly. All the surfaces
exposed to combustion gases are polished to a mirror finish including piston tops, combustion chamber,
valves faces, and exhaust ports (see Figure 8 and 9). The topic of polished surfaces and power gains is
controversial hence the reason to perform this process is to primarily reduce carbon buildup. Carbon
buildup can be detrimental to engine
performance and can be reduced if the surfaces
are slick making it difficult for byproducts to
stick. Additional benefits to polished surfaces
exist but are out of the scope of this study and
will not be discussed in this report. Engine
modifications that can be measured are more
evident to performance and can easily be
quantified including enlarging the intake and
exhaust ports.
The intake and exhaust ports are responsible for air delivery and exhausting wasted gases. Flow rates
through the ports are critical to the desired performance of the engine. The three factors of concern to
the cylinder ports include velocity, volume, and surface smoothness. Velocity is critical for proper
cylinder filling effects and rate of discharging exhaust gas. The volume component is critical for power
production potential. Volume verse velocity is a balancing act that must be considered when optimizing
the size of the ports. At the same mass flow rate, enlarged ports reduce charge velocity and may
decrease cylinder filling effects for a given condition while increasing maximum power potential.
Both intake and exhaust ports are enlarged by about 10%
over the original ports (see Figure 10). The surface finish
(roughness factor) of the ports contributes to flow but may
induce other adverse effects. The exhaust ports are
polished to a mirror finish to promote the reduction of
drag losses through the ports and reduction of possible
carbon buildup. Polished intake ports are ideal for dry air
charge but in this case fuel is injected in the intake ports
and wall wetting effects becomes a factor of concern.
When fuel droplets strike a smooth glass like surface, it is
likely to stick and "wet" the surface. If the surface is rough,
the fuel droplets are less likely to stick resulting in a better
charge mixture. Based on this concept, the intake ports
received less treatment than the mirrored finish exhaust ports.
Rerouting the belts became necessary with the removal of the supercharger and vacuum pump. The
vacuum pump, which is bolted to the oil pump housing, is driven by one of the two accessory belts.
After several revisions the simplest arrangement is to run a single belt setup. The vacuum pump portion
Figure 10: Porting cylinder heads
exhaust ports
Figure 9: Mirrored Polished Pistons
19. Figure 13: KJ valve spring (left)
verse uprated KL valve spring (right)
Figure 12: Reconfigured water coolant system
of the oil pump housing is removed so that the
alternator and air conditioning compressor
mounted tightly to the engine. The
crankshaft's outer pulley is removed to allow
single belt configuration and also provide
clearance to fit into the engine bay of the Ford
Probe. Mixing various Mazda parts and
fabrication, a proper belt alignment and pulley
warping is achieved while simplifying the belt
drive system. The current reconfiguration (see
Figure 11) resulted with removing the outer
pulley and harmonic damper. Knowing the
benefits of a harmonic damper, it is difficult to
make any statements about the benefits of
removal and thus easier to make statements
about the negatives effects. The degree of
benefits that a harmonic damper provides
vastly depend on engine design. In order to
move forward the removal of the harmonic
damper must be overlooked.
Revisions to the coolant system plumbing is
made in the "V" of the engine where the
supercharger is originally nested. The
supercharger bolts to a cradle that contained
integrated water passage and the supercharger
oil send and return ports. The approach is to
simplify and adapt the infrastructure to fulfill
requirements of the new system. A cover plate
with a coolant inlet port and oil send port are
fabricated and interfaced with various Mazda
components to achieve desired results (see Figure 12).
Higher engine speeds can be realized by increasing the rev-
limiter to 7200 rpm compared to the stock 6000 rpm. Valve
float is reduced by up uprated the valve springs by 15
percent (see Figure 13). Top end is enhanced.
Figure 11: Revised pulleys and belt routing
20. Gear Stock Revised
1 3.307 3.307
2 1.833 1.833
3 1.310 1.233
4 1.030 0.914
5 0.795 0.717
Final Drive 4.388 3.850
Gear Stock Revised
1 4.9 5.5
2 8.8 10.0
3 12.3 14.9
4 15.6 20.1
5 20.3 25.6
Stock 3,400 RPM
Revised 2,735 RPM
Gear Ratio
MPH/1000RPM
Engine Speed at 70 MPH in 5th Gear
Table 2: Stock and revised gear reatio
5.6. Transmission Modifications
Power deliver to the ground is increasingly difficult with higher power
engines. The differential is responsible for properly distributing power to
the drive wheels, thus gaining traction and accelerating the vehicle
rapidly. The stock transmission differential is an open type differential
that is known to be weak in terms strength in the Mazda "tuning scene."
Due to the 75 percent increase in torque, the differential was upgraded
to a high performance mechanical torque biasing helical type limited slip
differential (LSD). The LSD is essential to the transmission component
upgrading process. Another aspect of consideration during the
transmission reconfiguration and rebuilding stage is the optimization of
gear ratio selection. Table 2 shows both stock and revised gear ratios.
The original transmission gearing found in the Ford Probe GT consists of a 4.388:1 final drive ratio with
a 0.795 5th
gear. This stock gear configuration results in 20 MPH per 1000 RPM, which is relatively low,
resulting in excessive engine speed for
typical highway driving conditions. Fuel
economy may decrease with excessive
engine speeding thus selecting more
suitable gearing is considered. While
cruising at typical highway speeds of 70
MPH, the original gearing resulted in
engine speeds of 3,500 RPM. The new
finial drive of 3.85:1 and 5th gear of
0.717 results in reducing engine speed
by a notable 665 RPM. The new
configuration of gearing is
implemented to help realize fuel
economy increases for typical highway
driving speeds. Figure 14 show the
revised internals of the transmission. Figure 14: Transmission Rebuild, Revised Gear Ratio,
and limited slip differential
21. 5.7. Wiring, Electrical, and ECU
Modern engines contain complex electronics systems that commonly control fuel metering, ignition
timing, boost levels, throttle control, variable valve timing, etc. The engine control unit (ECU) is
responsible for controlling all engine parameters to ensure proper drivability, durability, emissions, fuel
efficiency etc. Various sensors feed signals to the ECU for processing and for controlling outputs. Mazda
implemented advance electrical components for the time era, establishing the "standard" for the
following decade. Highly developed, the KJ engine ECU precisely control various systems including;
sequential fuel injection, sequential coil-on-plug ignition, mass air flow sensor (MAF) , supercharger air
bypass control, intercooler bypass control, and intake air charge pressure control. The ignition timing of
the Miller cycle engine is unique and differs from the Otto cycle engine due to the overly expanded
combustion cycle and delayed closing of the intake valve. In order to unlock the potential of the engine
with a turbocharger and lean burn, full engine control is require, thus making the KJ ECU inadequate.
Furthermore, developing proper fuel maps and more so ignition maps requires considerable time and
testing.
Considering all options for wiring the engine
and tuning the standalone ECU properly, the
most reasonable approach was to wire the
stock ECU for the initial phase of developing
and tuning. The Ford Probe GT wire harness
was used with a custom built patch adapter to
provide correct signal routing (see Figure 15).
A near "plug-and -play" arrangement is
achieved with the patch adaptor and the
addition of a wire bundle (about 8 wires) for
the support of addition sensors (2 additional
intake air temperatures sensors) and 6 ignition
coil-on-plug. This wiring approach is time
consuming and complex, but ensures the engine will properly run with Mazda KJ ECU. Once running,
data will be recorded from the OBD-II port of the KJ ECU to help develop the initials ignition timing maps
and MAF calibration curve. The next step includes wiring the programmable standalone ECU in a parallel
arrangement splitting various signals to both KJ ECU and Megasquirt 3 ECU (MS3).
Once data is collected, and the ignition timing map established, and MAF flow curve calibrated, the MS3
can take full control of all the engine's parameters including; fueling, spark timing, boost control, rev-
limiting, water injection mapping, traction control, etc. In order to make use of the uprated valve
springs, the rev-limit is raised from 6000 RPM to 7000 RPM with the use of the MS3. The fuel injectors
seem to be the only outdated system found on the KJ engine. Controlled by MS3, modern 12-hole Denso
fuel injectors were implemented to provide the best possible fuel atomization. Full engine ECU control is
required for optimizing various modification.
Figure 15: Custom wire patch from ECU to wire harness
22. 5.8. Air Intake Manifold Fabrication and Modifications
The order of components in a supercharged system can differ greatly for turbocharged system, thus
effecting throttle response and boost control. The order of components found in the intake air path of
the supercharged system starts with the air filter and passes in order through the MAF, throttle body,
supercharger, dynamic chamber, intercoolers, and lastly into the intake manifold and engine. If the
turbocharger is placed in the same order as the supercharger, throttle input control would suffer, thus
reducing drivability.
Durability would also
suffer due to
subjecting the
turbocharger
compressor to high
vacuum found
between the valves
and throttle body.
Another notable issue
found with the original
system is the location
of the intercoolers. The
intercoolers were also
placed between the
throttle body and the
intake manifolds
(plenums), thus
increasing effective air charge volume
between the throttle body and intake
valves. As a result of the increased air
charge volume, throttle response will
decrease, hindering desire throttle
input. Full reconfiguration of the intake
manifolds and throttle body is
performed to accomplish the different
requirements of the turbocharged
system(see Figure 16).
Smooth piping transition are
implemented to reduce turbulent flow
within the air charge path (see Figure
17). The throttle body is placed just
before the intake manifold,
consequently reducing unwanted air
volume. A "clean" looking intake
Figure 16: Intake Manifold Fabrication and Layout
Figure 17: Smooth flowing end caps before welding
23. manifold is achieved by using the simpler front intake manifold plenum for both banks of the engine.
Extensive custom fabrication to effectively interface the components is required and ensure proper air
charge flow. Two factors of concern when sizing pipe diameter is velocity and mass flow rates. For a
fixed flow rate a larger diameter pipe will have slower velocity compared to the smaller pipe, but can
support higher mass flow rate which is required for making high power. Cylinder fill effects can benefit
with higher velocity up to mach speed of 0.40. Air becomes increasing turbulent above mach 0.40. A
needed to optimize the pipe diameter is require for desire results.
5.9. Air Charge System Enhancements
Deleting the intercoolers, reconfiguring the intake manifold, relocating the throttle body and fitting a
turbocharger requires fabricating a new intake air charge system. Pipe diameters are matched to fix
components of the system to enhance smooth flow. Fixed inlet and outlet diameters components
include; the MAF sensor, the turbocharger compressor housing, the throttle body, and the intake
plenum. Corresponding pipe diameters where matched and selected to attain proper flow. Selected
piping diameters can be seen in Figure 18.
1. Throttle Body...............................2.36"
2. Transition 1..................................2.65"
3. Transition 2..................................2.25"
4. Transition 3..................................2"
5. Intake Plenum..............................2"
Air Charge Components and Inside Diameters:
6. MAF Sensor..................................2.7"
7. MAF to Compressor Inlet Pipe......2.75"
8. Compressor Inlet..........................2.4"
9. Charge Pipe..................................2.1"
Figure18: Air Charge Piping Diameters
24. y = 0.0041x - 0.0013
R² = 0.9994
0
0.01
0.02
0.03
0.04
0 2 4 6 8 10
cc/pulse
PW
Fuel injector 'LATENCY'
Figure 19
5.10. Exhaust System Enhancements
Since the KJ engine shares similar architecture to the other K-series engines, the exhaust system
components are nearly identical, thus utilizing the available aftermarket turbocharger piping system.
One notable difference between the KJ engine and other K-series engine is the sizing of the exhaust
ports. The KJ engine has smaller exhaust ports possibly due to the reduced energy found in the
exhausted gas from the over expanded cycle. The turbo-pipes required minor fabrication to be fitted to
the KJ engine. The wastegate is routed back into the muffled exhaust system to reduce unwanted noise
and to ensure exhaust does not enter the cabin. The exhaust system consists of 2.5" pipe diameter
starting at Y-collector, through the turbocharger and continuing to the muffler. Surprisingly, the level of
engineering on these mass-produced turbo-pipes seem above average when viewing other available
systems for other engines. The diameters of the headers pipes, the y-collectors, the turbo up-pipe, and
the turbo down-pipe all seem to be sized properly and flow smoothly at transitions. Careful attention to
fitment is given to ensure gaskets do not interfere with flow and interior welds near flanges are
smoothed.
5.11. Fuel Enhancements and Control Methods
Reducing the size of the fuel droplets can potentially
improve fuel efficiency, emissions, and increase power. The
original fuel injectors found on the KJ engine can be
considered an outdated design. Improve fuel atomization
can be accomplished with modern 12-hole fuel injectors
manufacture by Denso, claiming mean droplets size of 50
microns. MS3 will control proper fueling to the upgraded
injectors. In order to have accurate fuel flow data calculated
from the MS3 ECU, direct data must be measured and
calculated from the injectors. A fuel injector is a mechanical
device that uses a magnetic coil to actuate a pintle valve. These valves have to overcome inertia, thus
delaying the time of reaction. This delayed time it takes the fuel injector to react is known as "injector
latency" or "injector dead time." Fuel flow rate must
also be measured.
Injector latency measurement test is performed on
the car with the test protocol enabled on the MS3
interface. By varying different pulse width (PW) and
number of injections, fuel is collected in a graduated
cylinder and recorded(see Table 3). These points are
graphed and a linear line is fitted. The line is
projected back to where it crosses the x-axis and in
used as the injector average latency value measure
in milliseconds. Based on the graph, latency
measures 0.32 milliseconds at 13.7 volts (see Figure
19). The voltage so the system must be accounted
Table 3: FUEL INJECTOR LATENCY
Denso 12-hole PINK 267 cc/min
Battery voltage 13.7 V
PW PULSE cc cc/pulse
2 7400 48 0.0065
4 3650 57 0.0156
6 2000 47 0.0235
8 1450 45 0.031
25. for and the test should be performed at various voltages to improve accuracy. Test are performed at 12
V and 13.2 V. The values are then plugged into MS3 and data is interpolated. At low PW values, fuel
injectors do not operate in a linear fashion. This can account for increased error in data logging during
low PW operation. Due to the small amount of error found below 2 ms PW values, the error will not be
seen during cruising conditions, thus omitting any real error in the range of operation.
Injector fuel flow must also be measure for calculated fuel flow data logging and initial configuration
table in MS3. A fuel pressure of 45 psi controlled by a regulator that references intake manifold pressure
to ensure the effective pressure differential remains constant for predictable and consistent fueling
under all operating conditions. Fuel flow testing is performed on the car with the MS3 test protocol
interface. Each injector ran for a fixed time at 100% duty cycle filling a graduated cylinder. All six
injectors are within flow specification to one another measuring a flow rate of 267 cc/min. The flow rate
of these updated injectors measures lower than the original KJ injectors (280 cc/min). Having a 5
percent reduction is in flow rate is not of any concern, especially when replacing the parasitic
supercharger with a turbocharger. If more power is required, the fuel system will reach the fueling limit
with the single fuel injectors per cylinder. The fuel system reaches the safe limit at boost pressures of 10
psi and 7000 rpm or 12 psi and 6300 rpm.
During spirited driving, boost pressure generated from the turbo became difficult to modulate by the
throttle, thus boost pressure rises rapidly, requiring the engine to be throttled back. Data is recorded
during highly loaded driving conditions, resulting in typical maximum required throttle position around
50 percent. Having near full power available at 50 percent made power modulation difficult. A method
of controlling boost pressure is required to improve drivability. The wastegate is a pressure control
device responsible for diverting excess exhaust gas around the turbine, thus controlling turbine speed
and limiting the work of the compressor and limiting boost pressure. The wastegate is configured with
the lowest spring available (3-4 psi). Boost pressure is controlled by MS3 and an air-bypass solenoid that
alters boost pressure to the wastegate, making boost pressure adjustable from 4 psi upwards. MS3
allows the throttle pedal position to modulate boost pressure, thus improving drivability. The pedal
position boost control is adjusted to allow 0-4 psi of boost below 30 percent and starts tapering to 12 psi
by 95 percent throttle. Boost modulation mapping can be adjusted for a desired result.
5.12. Air Charge Cooling Enhancements
Forced induction is achieved by "forcing" (compressing) air into the engine. This process of compressing
air increases temperature, adversely affecting the engine's performance by reducing air charge density
and increasing knocking potential. Ignition timing is then less than optimum if retarded timing is require
to avoid knock. A standard method to reduce charge temperature is to implement an intercooler.
Although an intercooler effectively reduces air charge temperatures, there are drawbacks to
intercoolers including; boost pressure reduction (pressure drop), packing size, extra weight, extra cost,
additional air piping, etc. Due to the confined spacing of the engine in the Ford Probe, a different
method of air charge cooling is implemented. Phase change cooling or water injection is the selected
method to address the excessive air charge temperatures generated from the turbocharger.
26. Phase change cooling a less common method of air charge cooling and may also be known as chemical
cooling, water injection, alcohol injection, water/alcohol injection, etc. The basic principle is that water
or alcohol-water mix is injected in the air inlet as a fine mist and the droplets changes phase (vaporizes),
thus cooling the air charge and increasing charge density. The amount of energy required to change
phase from a liquid to a vapor is known as latent heat of vaporization. Gasoline has a much lower latent
heat of vaporization than water, about 600 and 2260 respectively, thus making water better
at cooling air charge than gasoline. A common practice of forced induction engine is to add extra fuel to
the charge during heavy load conditions. Typically, nature aspirated (NA) engines under heavy load are
tuned with a air to fuel ratio (AFR) of about 13.2:1, while forced induction is about 11.5:1. The extra fuel
seen in the forced induction tune is for cooling effects in the air charge and combustion chamber
surfaces. In practice, water can be use to provide air charge cooling and "in-cylinder" cooling so that the
AFR can be leaned out towards the ideal torque producing ratio of 13.2. Knocking becomes suppressed
with the presence of water injection thus ignition timing can be further advanced. Water injection may
also improve combustion stability as the water droplets change to steam during combustion.
Direct port water injection nozzles are placed in each intake runner upstream the fuel injector to ensure
even distribution (see Figure 20). A 200 psi pump, a fast acting valve (fav), and MS3 is responsible for
metering water. A 10-20 percent water to fuel (by mass) is targeted for ideal results. An additional water
nozzle is placed
upstream the
compressor to
further increase air
charge cooling. Pre-
compressor water
injection is known
to help cool the air
charge rather than
in-cylinder cooling.
The system consist
of a total of seven
55 cc/min nozzles.
Water injection
systems can be
complicated and
difficult to properly
tune. Extensive
testing is required
to ensure validity of
this system. Figure 20: Direct Port Water Injection and Fuel System
water injection nozzle
27. 6. Results
6.1. Testing Methodology and Equipment
Engine's power was measured on a Mustang chassis dynamometer and various data was collected from
the dyno's software. MS3 data was recorded and overlaid with the dyno results to attain brake specific
fuel consumption (BSFC), brake mean effective pressure (BMEP), and brake thermal efficiency under
various boost pressures. The intake air temperature after the turbocharger is recorded and is used to
determine the working efficiency of the compressor and effectiveness of air charge cooling from the
pre-compressor water injection nozzle. A wideband oxygen sensor supplied from Ecotrons is used to
accurately fuel the engine and monitor the oxygen content in the exhaust gas. The MAF flow rate data
from the OBD-ii system is checked to verify mass flow rate and to properly match the turbocharger's
compressor size by overlay data on the compressor map to validate various states of operations.
Exhaust gas temperature is monitored to ensure temperatures do not exceed the upper limit. Exhaust
back pressure caused by the turbocharger is measured to verify proper turbine housing size. Fuel
injector flow rates and injector latency are measured to ensure precise fuel flow data is recorded.
6.2. Performance
In order to analyze the turbocharged Miller cycle engine, accurate measurements of power out is
necessary. A chassis dynamometer (dyno) is used to measure horse power and torque at the wheel.
Note that the measures power output is to the wheels and not at the engine. A 15 percent drivetrain
loss is used to calculate engine power. The Ford Probe test vehicle with the Miller cycle engine's power
output measures 217 hp and 228 lbs-ft to the wheels at 12 psi of boost. Torque is increased by 5 percent
with leaning the AFR from 11.5 to 12.5. In order to directly compare this engine's efficiency with other
engines displacement, power output, and fuel usage must be recorded. There are many factors of
efficiencies that must be determined such as highway cruising fuel economy, power production with
respect to fuel usage (BSFC), brake mean effective pressure (BMEP), and brake thermal efficiency. Once
data is collected, determining the most efficient boost pressures setting can be made and ideal engine
displacement can be determined for specific applications. See Figures 21 and 22 for engine performance
at various boost levels.
29. 6.3. Brake Specific Fuel Consumption (BSFC)
BSFC is the rate of fuel consumption divided by the power produced. This allows efficiencies of different
engines to be directly compared making it a valuable measurement. The loading condition and engine
speed influence engine efficiency, thus changing BSFC. Various conditions can be studied to determine
ideal load condition. Determining engine displacement (downsizing) and gear ratios can be derived to
achieve ideal loading for a given cruise speed to maximize cruise efficiency.
The equitation for BSFC is: ,
where is brake power produced at engine in horse power and is mass flow rate of fuel in pounds
per hour.
3 psi testing:
A typical BSFC for NA engines is
.40-.52, for turbocharged engines
.55-.60, and for supercharged
engines .60-.70. See Table 4 for
BSFC at various boost pressures.
6.4. Brake Mean Effective Pressure (BMEP)
BMEP can be viewed as the average pressure over the expansion cycle. This becomes an effective
parameter when comparing different engine's ability to do work. Engine speed and displacement are
factored into the equation. See Table 5 for BMEP at various boost pressures. Typical BMEP range for
naturally aspirated gasoline engines range 125-150 psi
(8.5-10 bar), for forced induction gasoline engines
range 180-250 psi (12.5-17 bar), and for forced
induction diesel range 200-269 psi (14-18
bar).[Haywood]
Break Mean Effective Pressure is:
where: = Torque
= Engine displacement
= use: four stroke=2 or two stroke=1
Boost
Pressure
(psi)
Engine
Cranksha
ft Power
(HP)
Fuel
Flow
total
(cc/min)
Fuel
Flow
total
(lbs/hr)
BSFC
(lbs/HP*
hr)
BSFC
(g/kW*hr
)
3 147 577.1 54.83 0.037 227
6 187 779.5 74.06 0.400 241
12 224 1216.7 115.58 0.520 315
Table 4: BSFC at Peak Torque at Various Boost Pressures
(measured at peak torque)
Boost
pressure
(PSI)
Engine
Speed
(RPM)
Peak
Torque
(lbs-ft)
BMEP
(PSI)
BMEP
(Bar)
3 4,100 188 206.5 14.2
6 4,300 224 250.9 17.3
12 4,400 268 294 20.3
Table 5: BMEP at Various Boost Pressures
(measured at Peak Torque)
30. for English units:
for SI units:
6.5. Brake Thermal Efficiency
Brake thermal efficiency is the ratio of brake power output to power input. It is used to evaluate how
efficient an engine converts the heat from a fuel to mechanical energy.
Brake thermal efficiency is:
where: = brake power
= theoretical power (HP)
= heating value of fuel (1900 BTU/lbs)
= volume of fuel (gallon/hour)
= mass of fuel (lbs/hour)
= fuel conversion efficiency (assume = 1)
At 3 psi boost, peak torque measures 188 lbs-ft with 150 hp and fuel flow rate of 9.147 gallons per hour
resulting with a Brake thermal efficiency of 36.4 percent.
Therefore;
Brake thermal efficiency for various boost pressure are tested to determine efficiency at various load
conditions (see Table 6). Note that brake thermal efficiency is measured at peak torque for each boost
pressure test. Analysis of brake thermal efficiency is also calculated and graphed for various engine
speed (see Figure 23).
31. 23%
25%
27%
29%
31%
33%
35%
37%
2500 3000 3500 4000 4500 5000 5500 6000 6500
BreakThermalEfficiency(%)
Engine Speed (RPM)
Brake Thermal Efficiency at Various Boost Pressure
3 psi 6 psi 12 psi
Figure 23: Efficiency and RPM
6.6. Highway Cruising Efficiency
Highway driving efficiency is a real world measure of low load efficiency. A measured 28 mpg is recorded
during a 200 mile trip traveling at speeds or 70-75 mph in cold weather (35°F). Another 200 mile test is
performed during warmer temperatures and speeds of 65 mph resulting in a measured 34 mpg. More
data is needed to further validate the low load efficiency of this particular configuration.
6.7. Intake Air Charging Cooling Effectiveness
A recorded 10 hp increase is measure with pre-compressor water injection during initial dyno testing at
8 psi boost pressure. Air intake temperature shows a 90°F drop when pre-compressor nozzle is active
during high load conditions. No ignition timing advancement is made during this test. A further power
increase may be realized if ignition timing is advanced to exploit the increased knock resistances caused
by cooler air charge effects of water injection. Data recording the effects of water injection on exhaust
gas temperature (EGT) became inaccurate do to a sensor failure, leaving this area of study unattainable.
Boost
pressure
Engine
Speed
Engine Peak
Torque
Theoretica
l Power
Power
(PSI) (RPM) (lbs-ft) (HP)
(HP)
3 4100 147 188 9.147 54.15 1028855 404 36.40%
6 4300 190 224 12.356 73.15 1389803 546 34.80%
12.5 4400 225 268 19.284 114.16 2169064 852 26.40%
Table 6: Brake Thermal Efficiency at Various Boost Pressures
(measured at Peak Torque )
Fuel Flow
Rate
(gal/hr)
Fuel Mass
Rate
(lb/hr)
Fuel
Energy
Rate
(BTU/hr)
Brake
Thermal
Efficiency
(%)
32. Figure 24: Compression efficiency map and performance output at various boost
pressures
Garrett T3 .42 A/R 45 trim
6.8. Turbocharger Compressor Efficiency
The compressor side of the turbocharger is relatively small for the required flow for the application.
Based on the compressor map of the Garrett T3 .42 A/R 45 trim, the compressor supports a maximum
flow up to about 23-25 lbs/min of air at boost pressures above 15 psi (see Figure 24). Compressor
efficiency calculations are used to check if the compressor is operating with in efficient range between
60 to 74 percent. During a typical 12 psi (pressure ratio 1.82) boost run without water injection, the
intake air temperature peaks up to 235°F with ambient air temperatures near 90°F. Based on these
recorded valves compressor efficiency can be calculated to ensure that the turbocharger is not outside
the proper working range.
Compressor efficiency is:
where: = compressor efficiency
= theoretical compressor outlet temperature (Rankin)
= compressor inlet temperature (Rankin)
= compressor outlet temperature (Rankin)
= Pressure ratio
33. The compressor efficiency is calculated to be 71 percent efficient at 12 psi of boost, therefore the
compressor is not over working causing unusually high air temperatures. See Table 7 for compressor
efficiency recorded during various conditions with and without pre compressor water injection. Pre-
compressor water injection cools air charge temperatures, thus effectively improving compressor
efficiency. Pre-compressor water injection shows to improve compressor efficiency over 100 percent.
This is due the high latent heat
energy of water causing the
heat energy to be removed
from the system during the
phase change from liquid to
vapor as the water passed
through the compressor.
Compressor efficiency seems to
increase with water injection at
higher boost pressures.
6.9. Turbocharger Turbine Restriction
The turbocharger is driven by exhaust gas flowing past a turbine. The turbine housing reduces the cross
sectional area of exhaust flow thus increasing the velocity of the exhaust gas and powering the turbine
wheel. The turbine housing is a point of restriction in the exhaust system and can be measure to ensure
proper sizing. A typical range that many automotive manufactures follow for exhaust back pressure to
intake manifold pressure ratio is between 1.5 to 2:1. A pressure gauge is connected to the exhaust
system upstream the turbine and a 3 foot 1/4 inch copper line to aid in cooling the exhaust charge (see
Figure 25). During a 10 psi boost pressure test, 15 psi of exhaust turbine backpressure is recorded, thus
making this particular condition a 1.5:1 pressure differential. A pressure differential ratio under most
conditions is measured between 1 to 1.5. These numbers verify that the turbine housing is not overly
restrictive. Based on numerous data and measured parameters, evidences of the compressor sizing
seems to be near the limit and not the turbine section. Concerns of the turbine being the limitation is
now ruled out as a concern and seems to be nicely matched for the intended application.
Boost T_out T_in PR T_out_th ηc
8 273 130 1.54 668 54.50%
8.7 273 130 1.59 674 58.6%
12 270 100 1.81 664 61.2%
15.5 220 75 2.054422 657.14 84.2%
13.5 280 115 1.918367 692.5851 71.3%
12.5 222 100 1.85034 667.5981 88.2%
13 265 135 1.884354 713.0227 90.8%
16.5 288 158 2.122449 766.1842 114.0%
Table 7: Compressor Efficiency
With Water Injection
34. 7. Discussion and Summary
To gain a further understanding of the turbocharged 2.3L V6 Miller cycle engine, several forms of
comparison are required to reveal its performance abilities. The original NA 2.5L V6 Otto cycle engine
found in the 1997 Ford Probe GT (test vehicle) is used to compare several performance parameters. It
seems reasonable to compare these engines for the reason that the displacement is similar and shares
near identical infrastructure. Therefore this discussion can aim on the focus of directly comparing the
Miller cycle engine verse the Otto cycle engine in a real world testing environment. The Miller cycle
engine is comparable to an indirect injection Diesel cycle engine in terms of brake thermal efficiency.
7.1. Performance and Fuel Economy
Various efficiency parameters of the turbocharged V6 Miller cycle engine are improved compared to the
"cousin" engine and thus by nature should have seen better highway mpg compared to base 2.5L V6
engine. Several factors of concern during testing are apparent and require additional testing to verify
results. The majority of the highway efficiency data is collected during the winter season, thus
decreasing mpg potential. During the cold season, aerodynamic drag forces increases from denser air,
lubricating fluids become more viscous, and the composition of gasoline differs. Even with these
concerns in mind, a slight improve in mpg are recorded when compared to the original engine. In similar
driving conditions, winter season, traveling speeds of 70 mph, the base engine averaged 26 mpg while
the Miller cycle engine recorded a 28.5 mpg.
Although this is an increase, higher mpg numbers were expected. Note the Miller cycle engine has taller
gearing than the base engine, this could have been the main contributing factor to the mpg increase.
Figure 25: Exhaust Back Pressure Monitoring Method
Copper line used to cool hot exhaust
charge for gauge monitoring.
35. The base engine can achieve 30 mpg at sustained cruise speeds of 65 mph during summer weather. The
Miller cycle engine recorded 34 mpg during a 200 mile trip. Under these condition a 13-14 percent gain
can be realized. In terms of power, output measures an impressive 265 hp and 285 lbs-ft torque at 15
psi of boost pressure. Turbocharging the Miller cycle engine is an effective method to provide adequate
power and at the same time excellent fuel efficiency.
7.2. Turbocharging The Miller Cycle Engine
Figure 21 and 22 shows the output performance curve of the turbocharged V6 Miller cycle gasoline
engine. Despite of its relatively small displacement, the engine's torque is that of a NA V6 3.7L engine. A
23 percent gain hp and a 30 percent gain in torque is realized by replacing the supercharger with a
turbocharger. Mazda's concern when developing the Miller cycle engine is the response delay of the
turbocharger thus Mazda's engineers selected the Lysholm supercharger for its near instant response to
throttle input. Turbo lag is minimized by properly selecting a turbine housing that matches the exhaust
flow characteristics of the Miller cycle engine. Boost threshold occurs around 1500 rpm while 10 psi of
boost occurs as early as 2700 rpm. Boost builds rapidly with increased engine speed, therefore selecting
the ideal gear can deliver required power with minimal delay of throttle input.
A sudden drop in torque is noticed between 4600-5400 rpm, rises from5400-6000, and drops again (see
Figure 22). A possible reason for the decrease torque output may be due to the air charge reversion in
the intake manifold caused by the Miller cycle process and/or improper ignition timing maps. This rapid
reversion of the air charge becomes more turbulent with increasing engine speeds, thus further
reducing cylinder air charge filling, therefore reducing VE and reducing performance. It seems that this
effect can be reduced with increased boost pressure. With increase boost pressure, sufficient air charge
cooling is required to reduce knocking and optimize ignition timing. It seems that the effectiveness of
water injection increases with greater temperature differentials, thus making higher boost feasible.
7.3. Water injection Air Charge Cooling
During phase change from liquid to vapor, water removes energy from the system, thus proving a
cooling effect to the surroundings. Under high load condition, the compressed air charge can reach up
to 300°F (150°C). Water is injected into the turbocharger's compressor, vaporizes, thus removing energy
from the system and effectively cooling the air charge. The benefits of this air charge cooling method
can be measures as a temperature decrease and ultimately power gains on the dyno. Compressor
efficiency is increased with pre-compressor water injection. In some cases, compressor efficiency can
exceed 100 percent. To understand what this means, when air is compressed to a given pressure ration
(PR), it will heat up to calculated theoretical value based on the conditions. A typical compressor
efficiency range between 65-75 percent, thus heating the air charge above the theoretical temperature
calculated.
Water injection is an effective cooling method to decrease air charge temperatures, thus calculated
compressor efficiency can exceed 100 percent. The compressed air charge temperature increase can be
lower than the theoretical temperature increase when water injection is used. A possible reason for is
the increase in compressor efficiency as boost pressures increases may be due to increase temperatures
differentials , thus causing a greater percentage of water to undergoing a phase change.
36. Without a working EGT sensor no accurate conclusion can be made on the effects of direct port water
injection on air charge cooling and in-cylinder cooling. If the system is effective then small decrease in
EGT can be measured. Normally, for every degree increase/decrease in air temperature, a
corresponding result can be measured in the exhaust gas, thus making cooler air charge beneficial on
reducing thermal loading. Forced induction methods increase air charge temperatures and power-
density, resulting in higher in cylinder temperatures. During heavy load conditions, retarding ignition
timing and additional fueling is required to avoid irregular combustion known as knocking and/or pre-
ignition. Compared to NA engine, forced induction engines require additional fuel to reduce in-cylinder
temperatures. AFR of 11.5 is typical on forced induced engines to ensure durability, whereas AFR for an
NA engine are usually near 13. Maximum torque output can be realized with AFR of 13.2, therefore
richer mixture are implemented for in-cylinder cooling rather than optimizing performance. Direct port
water injection is implanted to provide in-cylinder cooling and reduce extra fuel used for cooling under
high loads. Water injection effectively cools the air charge, thus allows a leaner 13:1 AFR and more
advance ignition timing resulting is a 5-10 percent gain in power.
When comparing power output with respect to displacement, it is an accomplishment to produce this
power-density without an intercooler and no apparent irregular combustion. The most notable reason
for this achievement is realized by the Miller cycle's low compression ratio and high expansion ratio.
Power gains are likely seen with the implantation of an intercooler.
7.4. Optimizations
The power produced from the turbocharged 2.3L V6 Miller cycle is adequate for most production cars.
To further improve efficiency, displacement downsizing can be used to match required power output.
Downsizing displacement as much as 50 percent will result with a 130 hp 1.15L three cylinder engine.
The power of the three cylinder engine is comparable to most production engines for the compact/mid-
size class vehicle. Too many variables are involved to calculate the potential mpg of the three cylinder
Miller cycle engine without real world testing. Another possible method to improve highway cruising
efficiency to select a taller 5th
gear. The 2.3L V6 has sufficient torque and can handle a an approximately
30 percent taller 5th
gear to reduce engine speed thus reducing fuel consumption.
37. 8. Acknowledgement
In this section I would like to acknowledge all those who supported the work and to those that provide
their highly specialized talents. A special thanks to Dr. Richard Cohen at Temple University College of
Engineering for excellent guidance throughout the work.
Thanks to Mitch Davis, the best welder I know, for your quality welding work and your willingness to
provide your skills in times of need.
Thanks to Jim Gannon from Temple University College of Engineering, for your advice and your
enthusiasm to share your expertise in the machine shop.
Thanks to Dane Pellicone, Temple University College of Engineering Alumni and Senior Engineer at
A.W.E. Tuning, for your support, your expert dyno services, and your valuable insight.
Thanks A.W.E. Tuning for your excellent dyno services.
Thanks Matt Lee, Manager at Ecotrons, for your support, endorsement, and product sponsorship.
9. Nomenclature
CR = Compression ratio
NA = Naturally aspirated
AFR = Air fuel ratio
TDC = Top dead centre
BDC = Bottom dead centre
ABDC = After bottom dead center
BBDC = Before bottom dead center
SI = Spark ignition
BMEP = Brake mean effective pressure
BSFC = Brake specific fuel consumption
ECU = Engine control unit
MS3 = Megasquirt 3 standalone ECU
DOHC = Dual overhead camshafts
ECIVT = Early closing intake valve timing
LCIVT = Later closing intake valve timing
10. References
Heywood, John B. Internal Combustion Engine Fundamentals. New York: McGraw-Hill, 1988.
Pulkrabek, Willard W. Engineering Fundamentals of the Internal Combustion Engine. Upper Saddle River,
NJ: Pearson Prentice Hall, 2004.
Goto, T., Hatamura, K., Takizawa, S., Hayama, N. et al., "Development of V6 Miller Cycle Gasoline." SAE
Technical Paper Series 940198, 1994, doi:10.4271/940198.
Sakono, T., Takizawa, S., Harada, S., Ikeda, T. et al., "Mazda New Lightweight and Compact V6 Engines,"
SAE Technical Paper 920677, 1992, doi:10.4271/920677.
